GRUNDFOS PUMP HANDBOOK
PUMP HANDBOOK
Copyright 2008 GRUNDFOS Pumps Corporation. All rights reserved. Copyright law and international treaties protect this material. No part of this material may be reproduced in any form or by any means without prior written permission from GRUNDFOS Pumps Corporation. Trademarks and tradenames mentioned herein are the property of their respective owners. Disclaimer All reasonable care has been taken to ensure the accuracy of the contents of this material; however, GRUNDFOS shall not be liable or responsible for any loss whether direct, indirect, incidental or consequential arising out of the use of or reliance upon any of the contents of this material.
Foreword Today’s processes place heavy demand on pumps when it comes to optimum operation, high reliability and low energy consumption. Therefore, we have developed the Grundfos Pump Handbook which, in a simple manner, deals with various considerations when sizing pumps and pump systems. This handbook, developed for engineers and technicians who work with design and the installation of pumps and pump systems, includes answers to a wide range of technical questions. The handbook can either be read from cover-to-cover or in part on specific topics. The handbook is divided into five chapters which deal with different phases when designing pump systems. Chapter 1 includes a general presentation of different pump types and components. Also described are precautions to consider when dealing with viscous liquids. Further, the most used materials, as well as different types of corrosion, are presented. Terminologies in connection with reading pump performance are presented in Chapter 2. Chapter 3 deals with system hydraulics and some of the most important factors to consider for optimum operation of the pump system. Pump performance adjustment methods are discussed in Chapter 4. Chapter 5 describes life cycle costs, as energy consumption plays an important role in today’s pumps and pump systems. We sincerely hope that you will find this handbook useful in your daily work.
Grundfos Pumps Corporation
Table of Contents
Chapter 1 Design of pumps and motors ......................7 Section 1.1 Pump construction ............................................................ 8 1.1.1 The centrifugal pump .................................................................. 8 1.1.2 Pump curves ................................................................................................ 9 1.1.3 Characteristics of the centrifugal pump ......................... 11 1.1.4 Most common end-suction and in-line pump types .............................................................................. 12 1.1.5 Impeller types (axial forces) ....................................................... 14 1.1.6 Casing types (radial forces) ......................................................... 15 1.1.7 Single-stage pumps............................................................................ 15 1.1.8 Multistage pumps ............................................................................... 16 1.1.9 Long-coupled and close-coupled pumps ........................ 16 Section 1.2 Types of pumps ................................................................... 17 1.2.1 Standard pumps .................................................................................... 17 1.2.2 Split-case pumps................................................................................... 17 1.2.3 Hermetically sealed pumps ........................................................ 18 1.2.4 Sanitary pumps ......................................................................................20 1.2.5 Wastewater pumps ............................................................................ 21 1.2.6 Immersible pumps .............................................................................. 22 1.2.7 Groundwater pumps......................................................................... 23 1.2.8 Positive displacement pumps ...................................................24 Section 1.3 Mechanical shaft seals .................................................. 27 1.3.1 The mechanical shaft seal’s components and function ............................................................29 1.3.2 Balanced and unbalanced shaft seals................................ 30 1.3.3 Types of mechanical shaft seals.............................................. 31 1.3.4 Seal face material combinations ...........................................34 1.3.5 Factors affecting the seal performance ...................... .....36 Section 1.4 Motors ....................................................................................39 1.4.1 Standards .................................................................................................... 40 1.4.2 Motor start-up ....................................................................................... 46 1.4.3 Voltage supply ........................................................................................ 47 1.4.4 Frequency converter .......................................................................... 47 1.4.5 Motor protection ................................................................................. 49
Section 1.5 Liquids ....................................................................................... 53 1.5.1 Viscous liquids ........................................................................................ 54 1.5.2 Non-Newtonian liquids .................................................................. 55 1.5.3 The impact of viscous liquids on the performance of a centrifugal pump ....................... ............. 55 1.5.4 Selecting the right pump for a liquid with antifreeze ....................................................................................... 56 1.5.5 Calculation example .......................................................................... 58 1.5.6 Computer-aided pump selection for dense and viscous liquids ......................................................................................... 58 Section 1.6 Materials ................................................................................ 59 1.6.1 What is corrosion? .............................................................................. 60 1.6.2 Types of corrosion ................................................................................ 61 1.6.3 Metal and metal alloys.................................................................... 65 1.6.4 Ceramics ....................................................................................................... 71 1.6.5 Plastics............................................................................................................ 71 1.6.6 Rubber ............................................................................................................ 72 1.6.7 Coatings ........................................................................................................ 73
Chapter 2 Installation and performance reading............................................................................................................. 75 Section 2.1 Pump installation ............................................................76 2.1.1 New installation .................................................................................... 76 2.1.2 Existing installation-replacement ........................................ 76 2.1.3 Pipe flow for single-pump installation ............................. 77 2.1.4 Limitation of noise and vibrations ....................................... 78 2.1.5 Sound level ................................................................................................ 81 Section 2.2 Pump performance ........................................................83 2.2.1 Hydraulic terms .....................................................................................83 2.2.2 Electrical terms ...................................................................................... 90 2.2.3 Liquid properties ...................................................................................93
Chapter 3 System hydraulics ..................................................95
Chapter 5 Life cycle costs calculation
Section 3.1 System characteristics .................................................96 3.1.1 Single resistances ................................................................................. 97 3.1.2 Closed and open systems............................................................. 98
Section 5.1 Life cycle costs equation ............................................128 5.1.1 Initial cost, purchase price (Cic) .............................................. 129 5.1.2 Installation and commissioning costs (Cin) ................ 129 5.1.3 Energy costs (Ce) ................................................................................. 130 5.1.4 Operating costs including labor (Co) ................................ 130
Section 3.2 Pumps connected in parallel and series ...................1 01
3.2.1 Pumps in parallel ................................................................................101 3.2.2 Pumps connected in series ....................................................... 103
5.1.5 5.1.6 5.1.7 5.1.8
Chapter 4 Performance adjustment of pumps ..................................................................................................... 105 Section 4.1 Adjusting pump performance ..............................106 4.1.1 Throttle control ....................................................................................107 4.1.2 Bypass control .......................................................................................107 4.1.3 Modifying impeller diameter ................................................. 108 4.1.4 Speed control ........................................................................................ 108 4.1.5 Comparison of adjustment methods............................... 110 4.1.6 Overall efficiency of the pump system..................... ...... 111 4.1.7 Example: Relative power consumption when the flow is reduced by 20% ....................... ................. 111 Section 4.2 Speed-controlled pump solutions ....................114 4.2.1 Constant pressure control .......................................................... 114 4.2.2 Constant temperature control ............................................... 115 4.2.3 Constant differential pressure in a circulating system ............................................................................. 115 4.2.4 Flow-compensated differential pressure control ...................................................................................116
........................1 27
Environmental costs (Cenv) ......................................................... 130 Maintenance and repair costs (Cm) ...................... ............... 131 Downtime costs (loss of production) (Cs) ..................... 131 Decommissioning or disposal costs (Cd) ..................... ... 131
Section 5.2 Life cycle costs calculation – an example ................................................................................................ 132 Appendix.......................................................................................................... 133 A) Notations and units .........................................................................134 B) Unit conversion tables ................................................................... 135 C) SI-prefixes and Greek alphabet ............................................ 136 D) Vapor pressure and specific gravity of water at different temperatures ................................................................. 137 E) Orifice .......................................................................................................... 138 F) G) H) I) J) K)
Change in static pressure due to change in pipe diameter ................................................................................. 139 Nozzles ........................................................................................................ 140 Nomogram for head losses in bends, valves, etc. ..................................................................... 141-150 Periodic system .................................................................................... 151 Pump standards .................................................................................. 152 Viscosity for typical liquids as a function of liquid temperature ............................................................. 153-157
Section 4.3 Advantages of speed control..................................1 17 Index ......................................................................................................... 158-162 Section 4.4 Advantages of pumps with integrated frequency converter............................................................................... 118 4.4.1 Performance curves of speed-controlled pumps ...........................................................................................................119 4.4.2 Speed-controlled pumps in different systems ........119 Section 4.5 Frequency converter .................................................... 122 4.5.1 Basic function and characteristics ...................................... 122 4.5.2 Components of the frequency converter ..................... 122 4.5.3 Special conditions regarding frequency converters ................................................................................................124
Chapter 1. Design of pumps and motors
Section 1.1: Pump construction 1.1.1 1.1.2 1.1.3 1.1.4
The centrifugal pump Pump curves Characteristics of the centrifugal pump Most common end-suction and in-line pump types
1.1.5 1.1.6 1.1.7 1.1.8 1.1.9
Impeller types (axial forces) Casing types (radial forces) Single-stage pumps Multistage pumps Long-coupled and close-coupled pumps
Section 1.2: Types of pumps 1.2.1 1.2.2 1.2.3 1.2.4 1.2.5
Standard pumps Split-case pumps Hermetically sealed pumps Sanitary pumps Wastewater pumps
1.2.6 Immersible pumps 1.2.7 Groundwater pumps 1.2.8 Positive displacement pumps
Section 1.1 Pump construction
1.1.1 The centrifugal pump In 1689, the physicist Denis Papin invented the centrifugal pump. Today, this kind of pump is the most commonly used around the world. The centrifugal pump is built on a simple principle: Liquid is led to the impeller hub and is flung towards the periphery of the impeller by means of centrifugal force. Fig. 1.1.1: The liquids flow through the pump
The construction is fairly inexpensive, robust and simple, and its high speed makes it possible to connect the pump directly to an asynchronous motor. The centrifugal pump provides a steady liquid flow, and it can easily be throttled without causing any damage to the pump. See figure 1.1.1 for liquid flow through the pump. The inlet of the pump leads the liquid to the center of the rotating impeller from where it is flung towards the periphery. This construction provides high efficiency and is suitable for handling pure liquids. Pumps which have to handle impure liquids, such as wastewater pumps, are fitted with an impeller that prevents objects from getting lodged inside the pump, see section 1.2.5.
Radial flow pump
Mixed f l ow pu mp
A xi a l f l o w p u m p
Fig. 1.1.2: Different kinds of centrifugal pumps
If a pressure difference occurs in the system while the centrifugal pump is not running, liquid can still pass through due to its open design. As you can tell from figure 1.1.2, the centrifugal pump can be categorized in different groups: Radial flow pumps, mixed flow pumps and axial flow pumps. Radial flow pumps and mixed flow pumps are the most common. These types of pumps are discussed on the following pages with a brief presentation of a positive displacement pump in section 1.2.8.
H [ft]
10000
1000
The different demands on the centrifugal pump’s performance, especially with regard to head, flow, and installation, together with the demands for economical operation, are only a few of the reasons why so many types of pumps exist. Figure 1.1.3 shows the different pump types with regard to flow and head.
100
10
H [m] 10000 6 4 2 1000 6 4
Multistage radial flow pumps
2 100 6 4
Single-stage radial flow pumps
2 10 6 4
Mixed flow pumps
Axial flow pumps
2 1
2
10
4 6 10 2
100
4 6 100 2
4 6 1000 2
1000
4 6 10000 100000 Q [m3 /h]
10000
Q [GPM] 100000
Fig. 1.1.3: Flow and head for different types of centrifugal pumps
8
1.1.2 Pump curves The performance of a centrifugal pump is shown by a set of performance curves. The performance curves for a centrifugal pump are shown in figure 1.1.4. Head, power consumption, efficiency and NPSH are shown as a function of the flow.
In some pump types with integrated motors and possibly integrated frequency converters, e.g. canned motor pumps (see section 1.2.3), the power consumption curve and the η-curve cover both the motor and the pump. In this case the P1-value has to be taken into account, see figure 1.1.5.
[%]
60 50 40 70 60 50 40
30
Efficiency
20
30 10
20 10
P200
Normally, pump curves in Grundfos product guides only cover the liquid end hydraulic performance. Therefore, the power consumption, the P2-value which is listed in the product guides as well, only covers the power going into the pump – see figure 1.1.4. The same applies for efficiency value, which only covers the liquid end (η = ηP).
η
H
[ft]
5
10
15
20 2 5
3 0
35
40 Q [GPM]0
NPSH (ft)
[hp]
Power consumption
0.6
20 15 10
0.4
NPSH
0.2
5
0
0
Fig. 1.1.4: Typical performance curves for a centrifugal pump. Head, power consumption, efficiency and NPSH are shown as a function of the flow
Q P1
M 3~
P2
H
In general, pump curves are designed according to Hydraulic Institute test standards or ISO 9906 Annex A, which specifies the tolerances of the curves.
ηM
ηP
Fig. 1.1.5: The curves for power consumption and efficiency will normally only cover the pump part of the unit – i.e. P2 and ηP
Following is a brief presentation of the different pump performance curves.
Head, the QH-curve H
[ft]
The QH-curve shows the head, identifying where the pump is able to perform at a given flow, see figure 1.1.6. Head is measured in feet liquid column [ft]; normally the unit feet [ft] is applied. The advantage of using the unit [ft] as the unit of measurement for a pump’s head is that the QH-curve is not affected by the type of liquid the pump has to handle, see section 2.2 for more information.
60 50 40 30
Efficiency
20 10 0 0
5
10
15
20
25
30
35
40 Q [GPM]
Fig. 1.1.6: A typical QH-curve for a centrifugal pump; low flow results in high head and high flow results in low head
9
Section 1.1 Pump construction
Efficiency, the η-curve The efficiency is the relationship between thesupplied power and the utilized amount of power. In the world of pumps, the efficiency ηp is the relationship between the power which the pump delivers to the water (PH) and the power input to the shaft (P2 ):
ηp = PH = QH . SG P2 3960 x P2
side of the pump to avoid cavitation (see section 2.2.1). The NPSH r value is measured in [ft] and depends on the flow. When flow increases, the NPSHr value increases, see figure 1.1.9. For more information concerning cavitation and NPSH, go to section 2.2.1.
η
[%] 80 70 60
where: SG is the specific gravity of the liquid. Q is the flow in GPM and H is the head in ft. ηp is the pump efficiency o
For water at 68 F and with Q measured in GPM and H in ft, the hydraulic power can be calculated as:
50 40 30 20 10 0 0
25
50
7 5
100
125
150
1 75
2
00
225
250
275 300
325
350
375
Q [GPM]
Fig. 1.1.7: The efficiency curve of a typical centrifugal pump
PH = lb of liquid per minute . H 33,000
Power consumption, the P2-curve The relationship between the power consumption of the pump and the flow is shown in figure 1.1.8. The P2-curve of most centrifugal pumps is similar to the one in figure 1.1.8 where the P2 value increases when the flow increases.
Q . H . SG P2 = 3960 x ηp
P2 [hp] 14 12 10 8 6 4 2 0 0
25
50
75
1
00 125
150
175
200
225
250
275
300
325
Q [GPM]
Fig. 1.1.8: The power consumption curve of a typical centrifugal pump
As it appears from the efficiency curve shown in figure 1.1.7, the efficiency depends on the duty point of the pump. It is important to select a pump that fits the flow requirements and ensures the pump is working in the most efficient flow area. NPSH [ft] 20
15
10
NPSH - curve (Net Positive Suction Head Required)
5
0 0
The NPSHr value of a pump is the minimum absolute head pressure that has to be present at the suction
10
25
50
75
100 125
150 1 75
200 22 5 2 50
275 3 00 325
Q [GPM]
Fig. 1.1.9: The NPSH curve of a typical centrifugal pump
1.1.3 Characteristics of the centrifugal pump The centrifugal pump has several characteristics and the most important ones are presented in this chapter. A more thorough description of the different pump types are given at the end of the chapter.
• The number of stages Depending on the number of impellers in the pump, a centrifugal pump can be either a single-stage pump or a multistage pump.
• The position of the pump shaft Single-stage and multistage pumps come with horizontal or vertical pump shafts and are normally designated as horizontal or vertical pumps. For more information, go to section 1.1.4.
• Single-suction or double-suction impellers Depending on the construction of the impeller, a pump can be fitted with either a single-suction impeller or a double-suction impeller. For more information, go to section 1.1.5.
• Construction of the pump casing Two types of pump casings are discussed: Volute casing and return channels. For more information, go to section 1.1.6. Fig 1.1.10: Example of multiple stage pump
11
Section 1.1 Pump construction
1.1.4 Most common end-suction and in-line pump types
End-suction
Horizontal
Single-stage
Long-coupled
Close-coupled
Close-coupled
End-suction pump
=
Liquid runs directly into the impeller. Inlet and outlet have a 90° angle. See section 1.1.9
In-line pump
=
Liquid runs directly through the pump in-line. The suction pipe and the discharge pipe are placed opposite one another and can be mounted directly in the piping system
Split-case pump
=
Pump with an axially divided pump housing. See section 1.2.2
Horizontal pump
=
Pump with a horizontal pump shaft
Vertical pump
=
Pump with a vertical pump shaft
Single-stage pump
=
Pump with a single impeller. See section 1.1.7
Multistage pump
=
Pump with several series-coupled stages. See section 1.1.8
Long-coupled pump
=
Pump connects to the motor by means of a flexible coupling. The motor and the pump have separate bearing constructions. See section 1.1.9
Close-coupled pump =
12
Multistage
Pump connects to the motor by means of a rigid coupling. See section 1.1.9
In-line
Horizontal
Vertical
Split-case Single-stage
Multistage
Single-stage Long-coupled
Long-coupled
Close-coupled
Close-coupled
13
Section 1.1 Pump construction
1.1.5 Impeller types There are three common types of pump impellers: open, enclosed and semi-open, see figure 1.1.11. The open impeller has a series of vanes attached to the center hub and is commonly chosen for low horsepower applications of clean, non-abrasive fluids or fluids with large solids. The enclosed impeller has vanes sandwiched between two shrouds. While the shrouds result in a slightly lower mechanical efficiency, they decrease the amount of pump casing wear caused by dirty or abrasive liquids. This design usually includes replaceable wear rings so critical clearances can be renewed. The semi-open impeller has a single shroud on one side of the vanes and it leaves one side open. This design can handle abrasives or solids well and often allows for simple axial adjustment of critical impeller-to-casing clearances without pump disassembly.
Axial Force Balancing
Open Fig. 1.1.11: Impeller types
Semi-open
Enclosed
Axial forces
Fig. 1.1.12: Single-suction impeller Fig. 1.1.13: Standard pump with single-suction impeller
A centrifugal pump generates pressure, exerting forces on both stationary and rotating parts of the pump. Pump parts are made to withstand these forces. If axial and radial forces are not counterbalanced in the pump, the forces have to be taken into consideration when selecting the driving system for the pump, such as angular contact bearings in the motor. In pumps fitted with a single-suction impeller, large axial forces may occur, see figures 1.1.12 and 1.1.13. These forces are balanced or avoided as follows: • Mechanically via thrust bearings. • Via balancing holes on the impeller, see figure 1.1.14 • Via throttle regulation from a seal ring mounted on the back of the impellers, see figure 1.1.15 • Via blades on the back of the impeller, see figure 1.1.16 • Through the use of double-suction impellers, see figure 1.1.17
14
Fig. 1.1.14: Balancing the axial forces in a single-stage centrifugal pump with balancing holes only
Fig. 1.1.15: Balancing the axial forces in a single-stage centrifugal pump with seal ring gap at discharge side and balancing holes
Fig. 1.1.16: Balancing the axial forces in a single-stage centrifugal pump with blades on the back of the impellers
Fig. 1.1.17: Balancing the axial forces in a double-suction impeller arrangement
1.1.6 Casing types
Fig. 1.1.18: Single-suction impeller
Radial forces
Radial forces are a result of the static pressure in the casing. Therefore, axial deflections may occur and lead to interference between the impeller and the casing. The magnitude and the direction of the radial force depend on the flow rate and the head. When designing the casing for the pump, it is possible to control the hydraulic radial forces. Two casing types worth mentioning are the single-volute and the double-volute. As seen in figure 1.1.19, both casings are shaped as a volute. The double-volute has a guide vane. The single-volute pump is characterized by a symmetric pressure in the volute at the optimum efficiency point, which leads to zero radial load. At all other points, the pressure around the impeller is not symmetrically equal and consequently a radial force is present.
Fig. 1.1.19: Single-volute casing
Double-volute casing
e c r o f l ia d a R
Single-volute casing
Double-volute casing 1.0
Q/Qopt
Fig. 1.1.20: Radial force for single and double-volute casing
As seen in figure 1.1.20,the double-volute casing develops a constant low radial reaction force at any capacity. Return channels (figure 1.1.21) are used in multistage pumps and have the same function as volute casings. Liquid is led from one impeller to the next. At the same time, water rotation is reduced and the dynamic pressure is transformed into static pressure. Because of the return channel casing’s circular design, no radial forces are present.
Fig. 1.1.21: Vertical multistage in-line pump with return channel casing Return channel
1.1.7 Single-stage pumps Generally, single-stage pumps are used in applications that do not require a total head of more than 450 ft. Normally, single-stage pumps operate in the range of 6-300 ft. Single-stage pumps are characterized by a low head relative to the flow, see figure 1.1.3. Single-stage pumps come in both a vertical and horizontal design, see figures 1.1.22 and 1.1.23.
Fig. 1.1.22: Horizontal single-stage Fig. 1.1.23: Vertical single-stage end-suction close-coupled pump in-line close-coupled pump
15
Section 1.1 Pump construction
1.1.8 Multistage pumps
Close-coupled pumps
Multistage pumps are used in installations where a high head is needed. Several stages are connected in series and the flow is guided from the outlet of one stage to the inlet of the next. The final head that a
These pumps can be constructed as follows: The pump’s impeller can be mounted directly on the extended motor shaft or the pump can have a standard motor and a rigid or a spacer coupling, see figures 1.1.28 and 1.1.29.
multistage pump is equal to the sum of the pressure that eachdelivers of the stages provide. Multistage pumps provide high head relative to the flow and have a steeper curve that is more advantageous for variable speed drive, also known as variable frequency drive (VFD) applications. Like the single-stage pump, the multistage pump is available in both vertical and horizontal versions, see figures 1.1.24 and 1.1.25.
Horizontal, Multistage Pumps This type of pump is somewhat unique. With the same benefits mentioned in 1.1.8, horizontal multistage pumps meet flow and head requirements of single-stage end-suction pumps but with significant reductions in required horsepower. In general, multistage pumps offer higher efficiencies when compared to single-stage end-suction pumps resulting in energy savings. Due to design, horizontal multistage pumps do not encounter the same vibration problems often associated with single-stage end-suction pumps.
Fig. 1.1.24: Vertical multistage in-line pump
Fig. 1.1.25: Horizontal multistage end-suction pump
Fig. 1.1.26: Long-coupled pump with basic coupling
Fig. 1.1.27: Long-coupled pump with spacer coupling
1.1.9 Long-coupled and close-coupled pumps Long-coupled pumps Long-coupled pumps have a flexible coupling (basic or spacer) that connects the pump and the motor. If the pump is connected to the motor by a basic coupling, the motor must be disconnected when the pump is serviced. The pump must therefore be aligned upon mounting, see figure 1.1.26. If the pump is fitted with a spacer coupling, the pump can be serviced without removing the motor and alignment is less of an issue, see figure 1.1.27.
16
Fig. 1.1.28: Close-coupled pump with rigid coupling
Fig. 1.1.29: Close-coupled pump with impeller directly mounted on motor shaft
Section 1.2 Types of pumps
1.2.1 Standard pumps Few international standards deal with centrifugal pumps. In fact, many countries have their own standards, which more or less overlap one another. A standard pump is a pump that complies with official regulations pertaining to the pump’s duty point. A couple of examples of international standards for pumps follow: •
Fig. 1.2.1: Long-coupled standard pump
ANSI B73.1 standard covers centrifugal pumps of horizontal end-suction single-stage, centerline design. This standard includes dimensionalinterchangeability requirements and certain design features to facilitate installation and maintenance.
• DIN 24255 applies to end-suction centrifugal pumps, also known as standard water pumps, with a rated pressure (PN) of 145 psi. The standards mentioned above cover the installation dimensions and the duty points of the different pump types. The hydraulic parts of these pump types vary according to the manufacturer – so, no international standards are set for these parts.
Fig. 1.2.2: Bare shaft standard pump
Pumps designed according tostandards provide end users with advantages in installation, service, spare parts and maintenance.
1.2.2 Split-case pumps A split-case pump is designed with the pump housing divided axially into two parts. Figure 1.2.4 shows a single-stage split-case pump with a double-suction impeller. The double-inlet construction eliminates the axial forces and ensures a longer life span of the bearings. Usually, split-case pumps have a rather high efficiency, are easy to service and have a wide performance range.
Fig. 1.2.3: Long-coupled split-case pump
Fig. 1.2.4: Split-case pump with double-suction impeller
17
Section 1.2 Types of pumps
1.2.3 Hermetically sealed pumps
Liquid Seal Atmosphere
The penetration point of the pump liquid by the shaft that allows it to connect to the impeller has to be sealed. Usually, this is addressed by a mechanical shaft seal, see figure 1.2.5. The disadvantage of the mechanical shaft seal is its poor handling of toxic and aggressive liquids, which consequently leads to leakage. This problem can often be solved by using a double mechanical shaft seal. Another solution is to use a hermetically sealed pump. There are two types of hermetically sealed pumps: Canned motor pumps and magnetic-driven pumps. In the following two sections, additional information about these pumps is provided. A disadvantage of hermetically sealed pumps is that they can handle very little, if any, solids in the pumped liquid.
Canned motor pumps A canned motor pump is a hermetically sealed pump with the motor and pump integrated in one unit without a seal, see figures 1.2.6 and 1.2.7. The pumped liquid is allowed to enter the rotor chamber that is separated from the stator by a thin rotor can. The rotor can serves as a hermetically sealed barrier between the liquid and the motor. Chemical pumps are made of materials, such as plastics or stainless steel, that can withstand aggressive liquids.
Fig. 1.2.5: Example of a standard pump with mechanical shaft seal
Motor can
Fig. 1.2.6: Chemical pump with canned motor
Motor can
The most common canned motor pump type is the circulator pump. This type of pump is typically used in heating or cooling applications because the construction provides low noise and maintenancefree operation.
Fig. 1.2.7: Circulator pump with canned motor
18
Magnetic-driven pumps In recent years, magnetic-driven pumps have become increasingly popular for transferring aggressive and toxic liquids. As shown in figure 1.2.8, the magnetic-driven pump is made of two groups of magnets: An inner magnet and an outer magnet. A non-magnetic can separate these two magnets. The can serves as a hermetically sealed barrier between the liquid and the atmosphere. As it appears from figure 1.2.9, the outer magnet is connected to the pump drive and the inner magnet is connected to the pump shaft. The torque from the pump drive is transmitted to the pump shaft by means of attraction between the inner and outer magnets. The pumped liquid serves as lubricant for the bearings in the pump. Therefore, sufficient venting is crucial for the bearings.
Outer magnets
Inner magnets
Can
Fig. 1.2.8: Construction of magnetic drive
Inner magnets Can Outer magnets
Fig. 1.2.9: Magnetic-driven multistage pump
19
Section 1.2 Types of pumps
1.2.4 Sanitary pumps Sanitary pumps are mainly used in food, beverage, pharmaceutical and bio-technological industries where liquid is pumped gently and pumps are easy to clean using clean-in-place (CIP)techniques. In order to meet process requirements in these industries, the pumps have to have a surface roughness less than 32 µ-in (0.8 µ-m) or better. This can be best achieved by using forged or deep-drawn rolled stainless steel as the material of construction, see figure 1.2.12. These materials have a compact pore-free surface finish that can be easily worked up to meet the various surface finish requirements. The U.S. recommended interior surface finishes range from 32 µ-in for food and beverage applications down to 10 µ-in for bioprocessing applications.
Fig. 1.2.10: Sanitary pump
The main features of a sanitary pump are ease of cleaning and ease of maintenance. The leading U.S. manufacturers of sanitary pumps have designed their products to meet the material specifications of the U.S. Food and Drug Administration (FDA) and the voluntary standards developed by 3-A Sanitary Standards Inc., as well as other well known globally-recognized standards such as:
Fig.1.2.11: Sanitary self-priming side-channel pump
EHEDG – European Hygienic Engineering Design Group QHD – Qualified Hygienic Design Sand casting
Precision casting
Rolled steel Fig.1.2.12: Roughness of material surfaces
20
1.2.5 Wastewater pumps Wastewater pumps can be classified as submersible and dry pit pumps. In submersible installations with sliderail systems, double rails are normally used. The auto-coupling system facilitates maintenance, repair and replacement of the pump. It is not necessary to enter the pit to perform service. In fact, it is possible to connect and disconnect the pump automatically from the outside of the pit. Wastewater pumps can also be installed dry, like conventional pumps, in vertical or horizontal installations. This type of installation provides easy maintenance and repair as well as uninterrupted operation of the pump in case of flooding of the dry pit, see figure 1.2.14. Normally, wastewater pumps must be able to handle large particles (i.e. 3-inch solids) and are fitted with special impellers to avoid blockage and clogging. Different types of impellers include: Single-channel impellers, double-channel impellers, three and fourchannel impellers and vortex impellers. Figure 1.2.15 shows the different designs of these impellers.
Fig. 1.2.13: Detail of a sewage pump for wet installations
Fig. 1.2.14: Wastewater pump for dry installations
Wastewater pumps with submersible motors shall carry the Underwriters Laboratories Inc label for class I, Divison I, Group D environment. Submersible wastewater pump motors are hermetically sealed and have a common extended shaft with a tandem mechanical shaft seal system in an intermediate oil chamber, see figure 1.2.13. Wastewater pumps are able to operate either intermittently or continuously, depending on the installation in question.
Vortex impeller
Single-channel impeller
Double-channel impeller
Fig. 1.2.15: Impeller types for wastewater
21
Section 1.2 Types of pumps
1.2.6 Immersible pumps An immersible pump is a pump type where the pump part is immersed in the pumped liquid and the motor is kept dry. Normally, immersible pumps are mounted on top of or in the wall of tanks or containers. Immersible pumps are used in the machine tool industry, in chip conveyor systems, grinding machines, machining centers, cooling units or in other industrial applications involving tanks or containers, such as industrial washing and filtering systems. Pumps for machine tools can be divided into two groups: Pumps for the clean side of the filter and pumps for the dirty side of the filter. Pumps with closed impellers are normally used for the clean side of the filter because they provide a high efficiency and a high pressure if necessary. Pumps with open or semi-open impellers are normally used for the dirty side of the filter because they can handle metal chips and particles. Refer to page 14 for more discussion on impeller types.
Fig. 1 .2.16: Immersible pump
22
1.2.7 Groundwater pumps There are two primary types of pumps used for groundwater applications: The submersible turbine pump type, which features a pump directly attached to a submersible motor and are completely submerged in thewith groundwater, and the line shaft turbine type a motor mounted at the top of thepump well which is connected to the submerged pump by a long shaft. Both pump types are used to pump groundwater from a well, typically for water supply and irrigation. Because these pump types must fit into deep, narrow wells, they have a reduced diameter compared to above-ground pumps making them long and thin compared to most other pump types. Submersible turbine pumps are specially designed to be fitted to a submersible motor, and the entire assembly is submerged in a liquid. The submersible motor is sealed to prevent water intrusion, and generally no regular maintenance is required on these pumps. Submersible pumps are preferred in deep installations and those requiring low to medium flow rates, generally up to 2,500 GPM. The liquid surrounding the submersible motor cools it, so submersible pumps are not suitable for hot water applications. Line shaft turbine pumps have been replaced in many applications by submersible turbine pumps but are preferred for certain applications such as shallow wells and those applications requiring higher flow rates. The long shaft is a drawback in deep settings making installation difficult and requiring frequent service. Because the line shaft turbine’s motor is aircooled, it is often used in industrial applications to pump hot water.
A
B
Fig. 1.2.17: Submersible turbine pump (A) and Line s haft turbine (B)
23
Section 1.2 Types of pumps
1.2.8 Positive displacement pumps
Fig. 1.2.19: Typical relation between flow and head for 3 different pump types: 1) Centrifugal pumps 2) Rotary pumps 3) Reciprocating pumps
H
The positive displacement pump provides an approximate constant flow at fixed speed, despite changes in the back pressure. Two main types of positive displacement pumps include:
1 H
• Rotary pumps • Reciprocating pumps The difference in performance between centrifugal, rotary and reciprocating pumps is illustrated in figure 1.2.19. Depending on the pump type, a small change in the pump’s back pressure results in differences in the flow. The flow of a centrifugal pump will change considerably with back pressure. Changing back pressure on rotary pumps will result in a minimal flow change. However, the flow of the reciprocating pump is almost constant with the back pressure change. The performance difference between reciprocating pumps and rotary pumps is due to the rotary pump’s larger seal surface area. Even though the two pumps are designed with the same tolerances, the loss due
Fig. 1.2.18: Rotary Lobe pump
24
3
2
2
1
Q
to the larger seal area of the rotary pump is greater. The pumps are typically designed with the finest tolerances possible to obtain the highest possible efficiency and suction capability. However, in some cases, it is necessary to increase the tolerances, for example, when the pumps must handle highly viscous liquids, liquids containing large particles or liquids of high temperature.
Metering pumps The metering pump belongs to thepositive displacement pump family and is typically of the diaphragm type. Diaphragm pumps are leak-free, because the diaphragm forms a seal between the liquid and the surroundings.
electrical parts caused by the solenoid operation, stepper motor-driven diaphragm pumps enable a more steady dose of additive.
The diaphragm pump is usually fitted with two or three non-return valves; one or two on the suction side and one on the discharge side of the pump. On smaller diaphragm pumps, the diaphragm is activated by the connecting rod, which is connected to a solenoid, permitting the coil to receive the exact amount of strokes needed, see figure 1.2.21. On larger diaphragm pumps, thediaphragm is typically mounted on the connectingrod, which is activated by a camshaft. The camshaft is turned by way of a standard asynchronous motor, see figure 1.2.22. Fig. 1.2.20: Dosing pump
The flow of a diaphragm pump is adjusted by changing the stroke length and/or the frequency of the strokes. If it is necessary to expand the operating area, frequency converters can be connected to the larger diaphragm pumps, see figure 1.2.22. Yet another kind of diaphragm pump exists. In this case, the diaphragm is activated by means of an eccentrically driven connecting rod powered by a stepper motor or a synchronous motor, figures 1.2.20 and 1.2.23. A stepper motor drive increases the pump’s dynamic range, thus improving its accuracy. This construction no longer requires stroke length adjustment because the connecting rod is mounted directly on the diaphragm. The result is optimized suction and operation due to full suction. Stepper motor drive design simplifies control of both the suction side and the discharge side of the pump. Compared to traditional electromagneticdriven diaphragm pumps which provide undesirable pulsations as well as fast wearing of mechanical and
Fig.1.2.21: Solenoid spring return +
1.2.22: Cam-drive assembly spring return
+
1.2.23: Stepper motor drive
25
Chapter 1. Design of pumps and motors
Section 1.3: Mechanical shaft seals 1.3.1 The mechanical shaft seal’s components and function 1.3.2 Types of mechanical shaft seals 1.3.3 Balanced and unbalanced shaft seals 1.3.4 Seal face material combinations 1.3.5 Factors affecting the seal performance
Section 1.3 Mechanical shaft seals
From the middle of the 1950s, mechanical shaft seals gained ground in favor of the traditional sealing method - the stuffing box. Compared to stuffing boxes, mechanical shaft seals provide the following advantages:
• None or minimal leakage of the fluid being pumped.
• No adjustment required • Seal faces provide a small amount of friction, minimizing power loss
• The shaft does not slide against any of the seal’s components and therefore reduces wear and associated repair costs. The mechanical shaft seal is the part of a pump that separates the liquid from the atmosphere. Figure 1.3.1 illustrates mechanical shaft seal mounting in different types of pumps. Before choosing shaft seal material and type, consider the following:
• Determine the type of liquid • Determine the pressure that the shaft seal is exposed to
• Determine the speed that the shaft seal is exposed to
• Determine the shaft-seal housing dimensions The following pages present how a mechanical shaft seal works, the different types of seals, materials used in mechanical shaft seals, and the factors that affect the mechanical shaft seal’s performance.
28
Fig. 1.3.1: Pumps with mechanical shaft seals
1.3.1 The mechanical shaft seal’s components and function The mechanical shaft seal is made of two main components: A rotating part and a stationary part. The parts of a shaft seal are listed in figure 1.3.2. Figure 1.3.3 shows where the different parts are placed in the seal.
• The stationary component of the seal is fixed in the pump housing. The rotating component of the seal is fixed on the pump shaft and rotates when the pump operates. • The two primary seal faces are pushed against each other by the spring (or other devices such as a metal bellows) and the liquid pressure. During operation, a liquid film is produced in the narrow gap between the two seal faces. This film evaporates before it enters the atmosphere making the mechanical shaft seal leak-free, see figure 1.3.4.
• The hydrodynamic lubricating film is created by pressure generated by the shaft’s rotation.
Mechanical shaft seal
Designation
Seal face (primary seal) Secondary seal Rotating component Spring Spring retainer (torque transmission) Seat (seal faces, primary seal) Static seal (secondary seal)
Stationary component
Fig. 1.3.2: The mechanical shaft seal’s components
Spring
Secondary seal Primary seal
Stationary part Rotating part
Spring retainer
Shaft
• Secondary seals prevent leakage from occurring between the assembly and the shaft. Secondary seal
• The spring or metal bellows press the seal faces together mechanically. • The spring retainer transmits torque from the shaft to the seal. In connection with mechanical bellows shaft seals, torque is transferred directly through the bellows.
Fig. 1.3.3: Main components of the
Primary seal
mechanical shaft seal
Vapor
Evaporation begins
Lubrication film Liquid force Spring force
Seal gap During operation, the liquid forms a lubricating film between the seal faces. This lubricating film consists of a hydrostatic and a hydrodynamic film.
Fig. 1.3.4: Mechanical shaft seal in operation
hydrostatic element is generated by the pumped • The liquid which is forced into the gap between the seal faces.
29
Section 1.3 Mechanical shaft seals
Start of evaporation 1 atm
Exit into atmosphere
1.3.2 Balanced and unbalanced shaft seals To obtain an acceptable face pressure between the primary seal faces, two kinds of seal types exist: A balanced shaft seal and an unbalanced shaft seal.
Balanced shaft seal Stationary seal face
Rotating seal face
Pump pressure
Entrance in seal
Pressure Liquid
Va p o r
A t m o s ph e r e
Fig. 1.3.5: Optimum ratio between fine lubrication properties and limited leakage
The thickness of the lubricating film depends on the pump speed, the liquid temperature, the viscosity of the liquid and the axial forces of the mechanical shaft seal. The liquid in the seal gap is continuously renewed due to:
• evaporation of the liquid to the atmosphere • Recirculation of the liquid Figure 1.3.5 shows the optimum ratio between fine lubrication properties and limited leakage. The optimum ratio occurs when the lubricating film covers the entire seal gap, except for a very narrow evaporation zone close to the atmospheric side of the mechanical shaft seal. Deposits on the seal faces may cause leakage. When using coolant agents, deposits build up quickly from evaporation at the atmosphere side of the seal. When the liquid evaporates in the evaporation zone, microscopic solids in the liquid remain in the seal gap as deposits, causing wear. These deposits are seen with most types of liquid. When the pumped liquid crystallizes, it canbecome a problem. The best way to prevent wear is to select seal faces made of hard material such as WC (tungsten carbide) or SiC (silicon carbide). The narrow seal gap between these materials (approx. Ra 0.3 µin) minimizes the risk of solids entering the seal gap, resulting in less buildup of deposits.
30
Figure 1.3.6 shows a balanced shaft seal indicating where the forces impact on the seal.
Unbalanced shaft seal Figure 1.3.7 shows an unbalanced shaft seal indicating where the forces impact the seal.
Contact area of seal faces Contact area of seal faces Spring forces
Hydraulic forces
Hydraulic forces
A
B
Fig. 1.3.6: Impact of forces on the balanced shaft seal
A
B
Fig. 1.3.7: Impact of forces on the unbalanced shaft seal
Several different forces have an axial impact on the seal faces. The spring and the hydraulic forces from the pumped liquid press the seal together while the force from the lubricating film in the seal gap counteracts this. With high liquid pressure, the hydraulic forces can be so powerful that the lubricant in the seal gap cannot counteract the contact between the seal faces. Because the hydraulic force is proportionate to the area that the liquid pressure affects, the axial impact can only be reduced by obtaining a reduction of the pressure-loaded area. The balancing ratio (K) of a mechanical shaft seal is
defined as the ratio between the area A and the area B : K=A/B
Comparative wear rates valid for water
K = 1.15 K = 1.00 K = 0.85
K = Balancing ratio A = Area exposed to hydraulic pressure B = Contact area of seal faces The balancing ratio for balanced shaft seals is around K=0.8 and for unbalanced shaft seals is around K=1.2.
1.3.3 Types of mechanical shaft seals
68
104
401
212
230
TemperatureoF) (
Fig. 1.3.8: Wear rate for different balancing ratios
The main types of mechanical shaft seals include: Oring, bellows, cartridge single-unit seal.
Advantages and disadvantages of O-ring seal
O-ring seals Sealing between the rotating shaft and the rotating seal face is affected by an O-ring’s movement (see figure 1.3.9). The O-ring must be able to slide freely in the axial direction to absorb axial displacements as a result of changes in temperature and wear. Incorrect positioning of the stationary seat may result in rubbing, which can cause wear on the O-ring and shaft. O-rings are made of different types of rubber material, such as NBR, EPDM, Buna -N and FKM, depending on operating conditions.
176
Advantages: Suitable in hot liquid and high pressure applications Disadvantages: Deposits on the shaft, such as rust, may prevent the O-ring shaft seal from moving axially causing leakage and premature failure Fig. 1.3.9: O-ring seal
Bellows seals Common to bellows seals is a rubber or metal bellows which functions as a dynamic sealing element between the rotating ring and the shaft.
Rubber bellows seal with folding bellows geometry
Advantages and disadvantages of rubber bellows seal
Rubber bellows seals The bellows of a rubber bellows seal (see figure 1.3.10) can be made of different types of rubber, such as NBR, EPDM, Buna-N and FKM, depending on the operating conditions. Two designs are used for rubber bellows:
Advantages: Not sensitive to deposits, such as rust, on the shaft Suitable for pumping solid-containing liquids
• Folding bellows
Disadvantages: Not suitable in hot liquid and high pressure applications
• Rolling bellows Fig. 1.3.10: Rubber bellows seal
31
Section 1.3 Mechanical shaft seals
Metal bellows seals
Advantages and disadvantages of cartridge metal bellows seal
In an ordinary mechanical shaft seal, the spring produces the closing force required to close the seal faces. In a metal bellows seal, the spring is replaced by a metal bellows with a similar force (see figure 1.3.11). Metal bellows act both as a dynamic seal between the rotating ring and the shaft and as a spring. The bellows have a number of corrugations to provide the desired spring force.
Advantages: Not sensitive to deposits, such as rust and lime, on the shaft Suitable in hot liquid and high-pressure applications
Fig. 1.3.11: Cartridge metal bellows seal
Low balancing ratio leads to low wear rate and consequently longer life Disadvantages: Fatigue failure of the mechanical shaft seal may occur when the pump is not aligned correctly Fatigue may occur as a result of excessive temperatures or pressures
Cartridge seals In a cartridge mechanical shaft seal, all parts form a compact unit on a shaft sleeve and are ready to be installed. A cartridge seal offers many benefits compared to conventional mechanical shaft seals, see figure 1.3.12.
Advantages of the cartridge seal: • Easy and fast service • The design protects the seal faces • Preloaded spring • Safe handling
Flushing In certain applications it is possible to extend the performance of the mechanical shaft seal by installing a flushing device, see figure 1.3.13. Flushing can lower the temperature of the mechanical shaft seal and prevent deposits from occurring. A flushing device can be installed internally or externally. Internal flushing is done when a small flow from the pump’s discharge side is bypassed to the seal area. Internal flushing is primarily used to prevent further heat generation from the seal in heating applications. External flushing is done by a flushing liquid and is used to ensure trouble-free operation when handling liquids that are abrasive or contain clogging solids.
32
Fig. 1.3.12: Cartridge seal
Fig 1.3.13: Flushing device of a single mechanical shaft seal
Double mechanical shaft seals Double mechanical shaft seals are used when the life span of a single mechanical shaft seal is insufficient due to wear caused by solids, or too high/low pressure and temperature. Double mechanical shaft seals help protect the surroundings when aggressive and explosive liquids are pumped. Two types of double mechanical shaft seals include: The double seal ina tandem arrangement and the double seal in a back-to-back arrangement.
Quench liquid
•
Pumped liquid
•
Double seal in tandem This seal consists of two mechanical shaft seals mounted in tandem, one behind the other, and placed in a separate seal chamber, see figure 1.3.14. The tandem seal arrangement must be fitted with an external barrier liquid system which:
• Absorbs leakage • Monitors the leakage rate • Lubricates and cools the outboard seal to prevent icing • Protects against dry-running • Stabilizes the lubricating film • Prevents air from entering the pump in case of vacuum Pressure of the external barrier liquid must always be lower than the pumped liquid pressure.
•
•
Fig. 1.3.14: Tandem seal arrangement with external barrier •
•
liquid circulation
Quench liquid
•
•
•
Pumped liquid
•
•
•
Fig. 1.3.15: Tandem seal arrangement with external barrier liquid dead end
Tandem - circulation For external barrier liquid circulation via a pressureless tank, see figure 1.3.14. External barrier liquid from the elevated tank circulates by thermosiphon action and/or by the pumping action in the seal. •
•
Tandem - dead end For external barrier liquid from an elevated tank, see figure 1.3.15. No heat is dissipated from the system.
Tandem - drain The external barrier liquid runs throughthe seal chamber to be collected for reuse or directed to drain, see figure 1.3.16.
•
•
Pumped liquid
•
•
Fig. 1.3.16: Tandem seal arrangement with external barrier liquid to drain
33
Section 1.3 Mechanical shaft seals
Barrier pressure liquid
Seal chamber with barrier pressure liquid
industrial applications: Tungsten carbide/tungsten carbide, silicon carbide/silicon carbide and carbon/tungsten carbide or carbon/silicon carbide.
Tungsten carbide/tungsten carbide Pumped liquid
•
Fig. 1.3.17: Back-to-back seal arrangement
Double seal in back-to-back This type of seal is the optimum solution for handling abrasive, aggressive, explosive or stickyliquids which would wear out, damage or block a mechanical shaft seal. The back-to-back double seal consists of two shaft seals mounted back-to-back in a separate seal chamber, see figure 1.3.17. The back-to-back double seal protects the surrounding environment and the people working with the pump. The pressure in the seal chamber must be 14.5-29 psi higher than the pump pressure. The pressure can be generated by:
• An existing, separate pressure source. Many applications incorporate pressurized systems.
Cobalt-bonded (Co) WC is only corrosion resistant in water if the pump incorporates base metal, such as cast iron. Chromium-nickel-molybdenum-bonded WC has a higher corrosion resistance. Sintered binderless WC has the highest corrosion resistance. However, its resistance to corrosion in liquids, such as hypochlorite, is not as high. The material pairing WC/WC has the following features:
• Extremely wear resistant • Very robust; resists roughhandling • Poor dry-running properties. In case of dry-running, the temperature increases to several hundred degrees Fahrenheit in just a few minutes and consequently damages the O-rings. If a certain pressure and temperature are exceeded, the seal may generate noise. Noise is an indication of poor seal operating conditions that, in the long term, may cause wear of the seal. The limits of use depend on seal face diameter and design.
• A separate pump, e.g. a metering pump
A WC/WC seal face pairmight be noisy during the break in period. Usually the noise dissapears after a couple of days
1.3.4 Seal face material combinations
of operation. In some cases noise may last up to 3-4 weeks.
What follows is a description of the most important material combinations used in mechanical shaft seals for
34
Cemented tungsten carbide covers the type of hard metals that are based on a hard tungsten carbide (WC) phase and usually a softer metallic binder phase. The correct technical term is cemented tungsten carbide; however, the abbreviated term tungsten carbide (WC) is used by Grundfos for convenience.
Silicon carbide/silicon carbide Silicon carbide/silicon carbide (SiC/SiC) is an alternative to WC/WC and is used where higher corrosion resistance is required.
P
P
In warm water, the Q1 / Q 1 face material pair generates less noise than the WC/WC pair; however, noise from porous SiC seals must be expected during the running-in wear period of 3-4 days. Q 1G self-lubricating, sintered SiC
The SiC/SiC material pair has the following features:
• Very brittle material requiring careful handling • Extremely wear resistant
Several variants of SiC materials containing dry lubricants G are available on the market. The designation Q applies 1 to a SiC material which is suitable for use in distilled or demineralized water, as opposed to theabove materials.
• High resistance to corrosion. SiC (Q1s, Q 1P and Q 1G ) hardly corrodes, independent of the pumped liquid type with the exception of water with very poorconductivity, such as the SiC variants Q1s demineralized water, which attacks and Q 1P. Q 1G is also corrosion - resistant in demineralized water
Pressure and temperature limits of Q1G / Q 1G are similar to those of Q 1P / Q 1P.
• In general, these material pairs have poor dry-running properties. However, the Q1G / Q 1G material withstands a limited period of dry-running due to the graphite content of the material
Carbon/tungsten carbide or carbon/ silicon carbide features
For different purposes, SiC/SiC variants include: Q 1s, dense-sintered, fine-grained SiC
The dry lubricants, such as graphite, reduce the friction in case of dry-running and are critical to the durability of a seal during dry-running.
Seals with one carbon seal face have the following features:
• Brittle material requiring carefulhandling • Are worn by liquids containing solid particles
A dense-sintered, fine-grained SiC with a small amount of tiny pores. For a number of years, this SiC variant was used as a standard mechanical shaft seal material. Pressure and temperature limits are slightly below those of WC/WC. Q 1P, porous, sintered, fine-grained SiC This porous-sintered SiC variant has large circular closed pores. The degree of porosity is 5-15% and the size of the pores is Ra 10-50 µin.The pressure and temperature limits exceed those of WC/WC.
• Good corrosion resistance • Good dry-running properties (temporary dry-running) • Self-lubricating properties (of carbon) make the seal suitable for use even with poor lubricating conditions (high temperature) without generating noise. However, such conditions will cause wear of the carbon seal face leading to reduced seal life. The wear depends on pressure, temperature, liquid diameter and seal design. Low speeds reduce the lubrication between the seal faces resulting in possible increased wear However, sincethe distance that the seal faces have to move is reduced, a shorter seal life may not be experienced
35
Section 1.3 Mechanical shaft seals
• Metal-impregnated carbon (A) offers limited corrosion resistance, but improved mechanical strength and heat conductivity, thus reducing wear
• The centrifugal pumping action of the seal’s rotating parts increases power consumption dramatically with the speed of rotation (to the third power)
• With reduced mechanical strength, but higher corrosion resistance, synthetic resin-impregnated carbon (B) covers a wide application field. Synthetic
• The seal face friction Friction between the two seal faces consists of – friction in the thin liquid film and
resin-impregnated carbon is suitable for drinking water
• The use of carbon/SiC for hot water applications may cause heavy wear of the SiC, depending on the quality of the carbon and water. This type of wear primarily applies to Q1 S/carbon. The use of Q1 P, Q 1G or a carbon/WC pair causes far less wear. Thus, carbon/WC, carbon/Q1P or carbon/Q1G are recommended for hot water systems
– friction due to points of contact between the seal faces The amount of power consumed depends on seal design, lubricating conditions and seal face materials.
0.25
Power loss (hp)
0.2
0.15
0.1 3600 0.05
1.3.5 Factors affecting the seal performance As mentioned previously, no seal is completely tight. On the next pages, factors which have an impact on the seal performance, such as energy consumption, noise and leakage, will be presented. While these factors will be presented individually, it is important to stress that they are closely interrelated and should be considered as a whole.
Energy consumption The following factors contribute to the power consumption of a mechanical shaft seal:
36
0 0
2000
4000
6000
8000
10000
12000
Speed (rpm)
Fig. 1.3.18: Power consumption of a 1/2 inch mechanical shaft seal
Pumping action Friction
Figure 1.3.18 is a typicalexample of the power consumption of a mechanical shaft seal. The figure shows that up to 3600 rpm friction is the major reason for the mechanical shaft seal’s energy consumption.
Energy consumption is, especially in connection with packed stuffing box, an important issue. Replacing a stuffing box with a mechanical shaft seal leads to considerable energy savings, see figure 1.3.19.
Standard pump 50 ft WCH; 2 inch shaft Energy consumption
Stuffing box 2.0 kwh Mechanical shaft seal 0.3 kwh Leakage
Noise The choice of seal face materials is critical for the function and the life of the mechanical shaft seal. Noise is generated as a result of the poor lubricating conditions in seals handling low viscosity liquids. The viscosity of water decreases with increasing temperature. This means that the lubricating conditions decrease as the temperature rises. If the pumped liquid reaches or exceeds boiling temperature, the liquid on part of the seal face evaporates resulting in decreased lubricating conditions. A speed reduction has the same effect, see figure 1.3.20.
Stuffing box .02 GPD (when mounted correctly) Mechanical shaft seal .005 GPD Fig. 1.3.19: Stuffing box versus mechanical shaft seal
psi 350
300
Noise
Duty range
250
200
Speed at 3600 rpm
150
Speed at 3000 rpm
100
Speed at 1800 rpm 50
Leakage
Speed at 1200 rpm
0 0
75
50
100
125
150
175
200
°F
2 25
Fig. 1.3.20: Relationship between duty range and speed
The pumped liquid lubricates the seal face of a mechanical shaft seal, providing improved lubrication resulting in less friction and increased leakage. Conversely, less leakage means poor lubricating conditions and increased friction. In practice, the amount of leakage and power loss occurring in mechanical shaft seals can vary because leakagedepends on factors which are impossible to quantify theoretically due to seal face type, liquid type, and spring load.
Exit into atmosphere
1 atm
Start of evaporation
Figure 1.3.21 shows how the lubricating film of fluid is evaporated into the atmosphere. Stationary seal face
Rotating seal face
Pump pressure
Pressure liquid
Entrance in seal vapor
Atmospheric
Fig. 1.3.21: Sealing gap
37
Chapter 1. Design of pumps and motors
Section 1.4: Motors 1.4.1 1.4.2 1.4.3 1.4.4 1.4.5
Standards Motor start-up Voltage supply Frequency converter Motor protection
Section 1.4 Motors
Motors are used in many applications all over the world. The purpose of the electric motor is to create rotation, that is to convert electric energy into mechanical energy. Pumps are operated by means of mechanical energy which is provided by electric motors.
Fig. 1.4.1: Electric motor
1.4.1 Standards
Fig. 1.4.2: NEMA and IEC standards
40
NEMA
IEC
The National Electrical Manufacturers Association (NEMA) sets standards for a wide range of electric products, including motors. NEMA is primarily associated with motors used in North America. The standards represent general industry practices and are supported by the manufacturers of electric equipment. The standards can be found in NEMA Standard Publication No. MG1. Some large motors may not fall under NEMA standards.
The International Electrotechnical Commission (IEC) sets standards for motors used in many countries around the world. The IEC 60034 standard contains recommended electrical practices that have been developed by the participating IEC countries.
Introduction to potentially explosive atmospheres
Area Classification
Potentially explosive atmospheres exist where there is a risk of explosion due to mixtures of gas/air, vapor/ air, dust/air or other flammable combinations. In such areas there is a need to eliminate ignition sources such as sparks, hot surfaces or static electricity which may ignite these mixtures. When electrical equipment is used where there is risk of explosion, the area must be so designed and constructed to avoid sources of ignition capable of igniting these mixtures. Before electrical equipment can be used in a potentially explosive atmosphere, a represenative sample must be fully tested and certified by an independent authority such as UL in the U.S.A. This information is intended as a guide only, and further expert guidance should be sought before placing the equipment into service or before maintaining or repairing any item of equipment in a potentially explosive atmosphere. Where showing comparisons, i.e., North American and European practices, these may be approximations and individual standards/codes of practice should be observed for precise details. E u r o p ean & IE C Cl a ss if ic at io n
Process plants are divided into Divisions (North American method) or Zones (European and IEC method) according to the likelihood of a potentially explosive atmosphere being present. Note: North American legislation now allows Zones to classify areas, and when used, the IEC Zone method is followed. See figure 1.4.3.
Gas Groups (plus dusts and fibers) There are two main gas groups: Group I - Mining only and Group II - Surface Industries. These categories are used in European and I.E.C. groupings. Group I gases relate to underground mining where methane and coal dust are present. Group II gases relate to surface industries and are sub-grouped according to their volatility. This enables electrical equipment to be designed with less onerous tolerances if it is to be used with the least volatile gases. See figure 1.4.4.
Def ini t io n o f z o ne o r di visi on
No r t h A m er ic an C l a ssi f ic at i on
Zone 0 (gases) Zone 2 0 (dusts)
A n area in which an explosive mixture is continuously present or present for long periods
C la ss I Division 1 (gases) C la ss II Division 1 (dusts)
Zone 1 (gases) Zone 2 1 (dusts)
A n area in which an explosive mixture is likely to occur in norma l opera tion
C la ss I Division 1 (gases) C la ss II Division 1 (dusts)
Zone 2 (gases) Zone 2 2 (dusts)
A n a rea in which an explosive mixture is not likely to occur in normal operation a nd if it occurs it will exist only for a short time
C la ss I Division 2 (ga ses) C la ss II Division 2 (dusts) Class III Division 1 (fibers) Class III Division 2 (fibers)
Fig. 1.4.3: Area Classification
Typica l ga s/ ma teria l Metha ne A cetylene Hydrogen Ethylene Propa ne
N or th A merica n G as Group -
Europea n/ I.E.C . G as G roup
A
IIC
B
IIC
I
C
IIB
D
IIA
E
-
C oa l dust
F
-
G ra in dust
G
-
Meta l dust
Fig. 1.4.4: Gas Groups
41
Section 1.4 Motors
Types of electrical equipment suitable for use in potentially explosive atmospheres Different techniques are used to prevent electrical equipment from igniting explosive atmospheres. See fig 1.4.5 for restrictions as to where these different types of equipment can be used.
IEC USA Area of use Area of use Designation Designation Standard Standard Zones 1 & 2 Class I Divisions 1 & 2 Exd – IEC60079-1 UL1203
Flameproof Enclosure – An enclosure used to house electrical equipment which, when subjected to an internal explosion, will not ignite a surrounding explosive a tmosphere. Intrinsic Safety– A technique whereby electrical energy is limited such that any sparks or heat generated by electrical equipment is sufficiently low as to not ignite an explosive atmosphere.
Class I Divisions 1 & 2 –
Zones 1 & 2
Exi IEC60079-11 Zones 1 &2 – Exi – IEC 6007 9-7 – Zones 1 & 2 Class l Divisions 1 & 2 – Exp NFPA 496 IEC 6007 9-2 Zone 1 & 2 – – EC 6Exm 007 9-1 8 – Zones 1 & 2 Class l Division 2 E xo – EC 6007 9-6 UL698 Zones 1 & 2 – UL1203
Increased Safety – This equipment is so designed as to eliminate spa rks a nd hot surfa ces capable of igniting an explosive atmosphere. Purged and Pressurized – Electrical equipment is housed in an enclosure which is initially purged to remove any explosive mixture then pressurizedto prevent ingress of the surrounding atmosphere prior to energization. Encapsulation – A method of exclusion of the explosive atmosphere by fully encapsulating the electrical components in an approved ma terial. Oil Immersion – The electrical components are immersed in oil, thus excluding the explosive atmosphere from any sparks or hot surfaces. Powder Filling – Equipment is surrounded with a fine powder, such as quartz, which does not allow the surrounding atmosphere to come into conta ct with any spa rks or hot surfa ces. Non-sparking – Sparking contacts are sealed against ingress of the surrounding atmosphere;, hot surfa ces are eliminated. Special Protection– Equipment is certified for use in a Potentially Explosive Atmosphere but does not conform to a type of protection listed above.
– – – – –
Exq IE C 6007 9-5 Zone 2
–
Exn EC 6007 9-1 5 Zones 0, 1 & 2
–
Exs
European Area of use Designation Standard Zones 1 & 2
EExd EN50018 Zones 0, 1 & 2
EE xi EN50020 Zones1&2 EExe EN 5 001 9 Zones 1 & 2 EExp E N 5 001 6 Zones 1 & 2
E NEExm 5 002 8 Zone 1 & 2 EE xo EN50015 Zones 1 & 2 EExq E N 5 001 7 Zone 2 EExN EN50021 Zones 0, 1 & 2 * Exs
Fig 1.4.5: Standards and methods of protection
42
North American practice
for general electrical requirements, e.g. light fittings.
Sample equipment and supporting documentation are submitted to the appropriate authority, e.g U.L., F.M., C.S.A. Equipment is tested in accordance with relevant standards for explosion protection and also
After successful testing, a listing is issued allowing the manufacturer to place the product on the market. The product is marked with the certification details such as the gas groups A,B,C,D and the area of use, e.g. Class 1 Division 1.
Temperature Hot surfaces can ignite explosive atmospheres. To prevent this from happening, all electrical equipment intended for use in a potentially explosive atmosphere is classified according to the maximum surface temperature it will reach while in service. This maximum temperature is normally based on a surrounding ambient temperature of 104° F (40° C). This temperature can then be compared to the ignition temperature of the gas(es) which may come into contact with the equipment and a judgement can be reached as to the suitabillity of the equipment to be used in that area, see figure 1.4.6.
NEMA Motor Enclosures The following describes NEMA Motor Enclosures:
• Open Drip Proof (ODP) Internal fan pulls air in, blows it across windings inside motor and exits opposite drive end. Motor is protected from drops of liquid or particles fallingat any angle from 0-15 degrees.
Temperature Classification North America European/IEC T1 T1 T2 T2 T2 A T2 B T2 C T2 D T3 T3 T3 A
T3 B T3 C T4 T4A T5 T6
842 ° F 5 7 2° F 5 3 6° F 5 00° F 4 4 6° F 41 9° F 3 92 ° F 3 5 6° F 3 2 9° F 3 2 0° F 27 5 ° F 2 48° F 21 2° F 185° F
T4 T5 T6
Fig 1.4.6:: Temperature classification
and the second digit stands for protection against ingress of water, see figure 1.4.7. Drain holes enable the escape of water entering the starter housing, i.e., through condensation. When the motor is installed in a damp environment, the bottom drain hole should be opened. Opening the drain hole changes the motor’s enclosure class from IP55 to IP44. F ir stdig i t
Protection against contact and ingress of solid objects
• TEFC-Totally Enclosed External fan pulls airin throughfan coverand blows it over the exterior (only) surface ofthe motor. More resistant to the liquid and particles. • Washdown - Totally Enclosed Spray Proof Corrosion-resistant. There can be a HP limit for rolled steel frame motors. Cast Iron finned motors do not meet FDA requirements. • Explosion Proof (xp) Enclosed motor designed to withstand an explosion of a specified dust, gas or vapor according to explosive environment standards.
Maximum Surface Temperature
Seco nddig it
Protection against ingress of water
0 No special protection
0 No special protection
1 The motor is protected against solid objects bigger than 55 mm, e.g. a hand
1 The motor is protected against vertically falling drops of water, such as condensed water
2 The motor is protected against objects bigger than 12 mm, e.g. a finger
2 The motor is protected against vertically falling drops of water, even if the motor is tilted at an angle of 15 degrees
3 The motor is protected against solid objects bigger than 25 mm, i.e. wires, tools, etc. 4 The motoris protectedagainst solid objects bigger than 1 mm, e.g. wires 5 The motoris protectedagainst ingress of dust 6 The motor iscompletely
3 The motoris protectedagainst water spray falling at an angle of 60 degrees from vertical 4 The motoris protectedagainst water splashing from any direction 5 The motoris protectedagainst water being projected from a nozzle from any direction
dust-proof 6 The motor isprotected against
IEC Motor Enclosures
heavy seas or high-pressure water jets from any direction
7 The motor is protected when
The IP rating states the degrees of protection of the motor against ingress of solid objects and water. The rating is stated by the letters “IP” followed by two digits, for example IP55. The first digit stands for protection against contact and ingress of solid objects,
submerged from 15 cm to 1 by m in water for a period specified the manufacturer 8 The motoris protectedagainst continuous submersion in water under conditions specified by the manufacturer
Fig 1.4.7: Two-digit IP enclosure class identification (IEC)
43
Section 1.4 Motors
Frame size Figure 1.4.8 gives an overview of the relationship between frame size, shaft end, and motor power. The figure shows where the different values that make up the frame size are measured on the motor. Flanges and shaft end comply with NEMA standards or EN 50347 and IEC 60072-1 for IEC. Some pumps have a coupling which requires a smooth motor shaft end or a special shaft extension which is not defined in the standards.
D
2F Distance between holes
Fig 1.4.8: Frame size
Insulation class Hot-spot overtemperature [°f] 356
The insulation class isabout defined in robust the NEMA and tells something how the standard insulation system is relative to motor operating temperatures. The life of an insulation material is highly dependent on the temperature to which it is exposed. The various insulation materials and systems are classified into insulation classes depending on their ability to resist high temperatures, see figure 1.4.9.
15
311
10
266 248
10
Maximum temperature increase 1 76
22 1 257
104
Maximum ambient temperature 104 104 104 B Class
F
H
Maximum Maximum Hot-spot Maximum ambient temperature temperature increase overtemperature winding temperature (Tmax) °(F) (°F) (°F) (°F) 266
B
104
144
18
F
104
189
18
311
H
104
225
27
356
Fig 1.4.9: Different insulation classes and temperature increases at nominal voltage and load
44
1 Frame Size (C-facemotors) 42C 48C 56C 66C
2 Shaft end diameter [in] 0.375 0.5 0.625 0.75
3 Rated power (TEFC Motors) 2-pole 4-pole 6-pole 8-pole [HP] [HP] [HP] [HP] In these fractional size motors, specific frame assignments have not been made by horsepower and speed. It is possible for more than one HP and speed combination to be f ound in a given frame size.
143TC 145TC 182TC 184TC 213TC 215TC 254TC 256TC 284TC 286TC 284TSC 286TSC 324TC 326TC 324TSC 326TSC 364TC 365TC 364TSC
0.875 0.875 1.125 1.125 1.375 1.375 1.625 1.625 1.875 1.875 1.625 1.625 2.125 2.125 1.875 1.875 2.375 2.375 1.875
1.5 2 3 5 7.5 10 15 20
365TSC 404TC 405TC 404TSC 405TSC 444TC 445TC 444TSC 445TSC
1.875 2.875 2.875 2.125 2.125 3.375 3.375 2.375 2.375
75
1 1.5, 2.0 3 5 7.5 10 15 20 25 30
1 1.5 2 3 5 7.5 10 15 20
1 1.5 2 3 5 7.5 10 15
40 50
25 30
20 25
60 75
40 50
30 40
100
60 75
50 60
125 150
100 125
75 100
25 30
40 50
60
100
125 150
Fig 1.4.10: The relationship between frame size and power input
45
Section 1.4 Motors
1.4.2 Motor start-up Methods of starting referred to in this section include: Direct-on-line starting, star/delta starting, autotransformer starting, soft starter and frequency converter starting, see figure 1.4.11. Starting method
Pros
Cons
Direct-on-line starting (DOL)
Simple and cost-efficient. Safe starting.
High locked-rotor current.
Star/delta starting (SD) (Y/∆)
Reduction of starting current by a factor of 3.
Current pulses when switching over from star to delta. Not suitable if the load has a low inertia. Reduced locked-rotor torque.
Autotransformer starting
Reduction of locked-rotor current and torque.
Current pulses when switching from reduced to full voltage. Reduced locked-rotor torque.
Softstarter
"Soft"starting.Nocurrentpulses. Less water hammer when starting a pump. Reduction of locked-rotor current as required, typically 2-3 times.
Reduced locked-rotor torque.
Frequency converter starting
No current pulses. Less water hammer when starting a pump. Reduction of locked-rotor current as required, typically 2 to 3 times. Can be used for continuous feeding of the motor.
Reduced locked-rotor torque. Expensive
Fig 1.4.11: Starting method
Direct-on-line starting
Autotransformer starting
As the name suggests, direct-on-line starting (DOL) means that the motor is started by connecting it directly to the supply at rated voltage. Direct-online starting is suitable for stable supplies as well as mechanically stiff and well-dimensioned shaft systems, i.e. pumps. Whenever applying the directon-line starting method, it is important to consult local authorities.
As the name states, autotransformer starting makes use of an autotransformer. The autotransformer is placed in series with the motor during start and varies the voltage up to nominal voltage in two to four steps.
Star/delta starting The objective of this starting method, which is used with three-phase induction motors, is to reduce the starting current. Current supply to the starter windings is connected in star (Y) configuration for starting. Current supply is reconnected to the windings in delta ( ) configuration once the motor has gained speed.
46
Soft starter A soft starter is a device which ensures a soft start of a motor. This is done by raising the voltage within a preset voltage rise time.
Frequency converter starting Frequency converters are designed for continuous feeding of motors, but they can also be used for soft starting.
1.4.3 Voltage supply The motor’s rated voltage lies within a certain voltage range. Figure 1.4.12 shows typical voltage range examples for 60 Hz motors. According to the NEMA standard, the motor has to be able to operate with a main voltage tolerance of ± 10% from the lowest and highest voltage in the range.
Typical North America voltage examples
60 Hz 60 Hz motors come with the following voltages: • 1 x 115 – 230 ∆ / 346 – 400 Y • 1 x 115/208-230 • 1 x 208-230 • 1 x 230 • 3 x 208-230/460 • 3 x 230/460 • 3 x 575 Fig 1.4.12: Typical voltages
1.4.4 Frequency converter Frequency converters are often used for speed controlled pumps, see chapter 4. The frequency converter converts the main voltage into a different voltage and frequency, causing the motor to run at a different speed. This way of regulating the frequency might result in some problems:
• Acoustic noise from the motor which is sometimes transmitted to the system as noise
• High voltage peaks on the output from the frequency converter to the motor
47
Section 1.4 Motors
Insulation for motors with frequency converters
Phase insulation also referred to as phase paper
The discussion below highlights different kinds of motors with frequency converters and how differentkinds of insulation affect the motor.
Motors without phase insulation For motors constructed without phase insulation, continuous voltages known as Root Mean Square voltages (RMS) above 460 V can increase the risk of disruptive discharges in the windings and destroy the motor. This applies to all motors constructed according to these principles. Continuous operation with voltage peaks above 650 V can cause damage to the motor.
Motors with phase insulation Phase insulation is normally used in three-phase motors. Specific precautions are not necessary if the voltage supply is less than 500 V.
Motors with reinforced insulation With supply voltages between 500 V and 690 V, the motor has to have reinforced insulation or be protected with delta U /delta T filters. For supply voltages of 690 V and higher, the motor has to be fitted with both reinforced insulation and delta U /delta T filters.
Motors with insulated bearings In order to avoid harmful current flows through the bearings, the motor bearings have to be electrically insulated. This generally applies to motors >- 40 hp run with variable frequency drives. Motor manufacturers will use special ceramic coatings to insulate one or both bearings.
48
Fig 1.4.13: Stator with phase insulation
Motor efficiency In general, electric motors are quite efficient. Some have electricity-to-shaft power efficiencies of 8093% depending on the motor size and sometimes even higher for bigger motors. There are two types of energy losses in electric motors: Load-dependent and load-independent losses. Load-dependent losses vary with the square of the current and cover:
• Stator winding losses (copper losses) • Rotor losses (slip losses) • Stray losses (in different parts of the motor) Load-independent losses in the motor refer to:
• Iron losses (core losses) • Mechanical losses (friction) Motors are categorized according to efficiency. The most important classifications are Environmental Protection Act in the US (EPact) and CEMEP in the European Union (EFF1, EFF2 and EFF3). 1
100
0.8
80
0.6 s o C
0.4
Motors can fail due to overload for long periods of time so are often intentionally oversized and operate at 75% to 80% of their full load capacity. At this level of loading, motor efficiency and power remain relatively high, but when motor load is less than 25%, efficiency and power decrease. Motor efficiency drops quickly below a certain percentage of rated load. Therefore, it is important to size the motor so that losses associated with running the motor too far below its rated capacity are minimized. It is common to choose a motor that meets the power requirements of the pump.
1.4.5 Motor protection Motors are usually protected against high temperatures that can damage the insulation system. Depending on motor construction and application, thermal protection can also prevent damaging temperatures in the frequency converter if it is mounted on the motor. Thermal protection varies with motor type. Motor construction and its power consumption must be considered when choosing thermal protection. Generally, motors must be protected against the following:
60
t n e c r e P
40
0.2
20
Fig 1.4.14: Efficiency vs. load and power vs. load (schematic drawing)
Efficiency Power factor 0
25
50 75 100 125 Per cent of rated load
150
100
100 hp
90
10 hp 80
1 hp
70
%60 y c n 50 e i ff ic E
Fig 1.4.15: The relationship between efficiency and rated load of different sized motors (schematic drawing)
40 30 20 10
Errors causing slow temperature increase in the windings: • Slow overload • Long start-up periods • Reduced cooling / lack of cooling • Increased ambient temperature • Frequent starts and stops • Frequency fluctuation • Voltage fluctuation Errors causing fast temperature increase in the windings: • Blocked rotor • Phase failure
0 0
25
50
75
100
125
150
175
Percent of rated load
49
Thermal protection A motor’s thermal protection (TP) is provided by a temperature-sensing device that is built in to the motor. When motor temperature becomes excessively hot due to failure-to-start or overloading, the sensor device shuts off the motor. This is especially important for motors that start automatically, are unattended, or for motors that are located remotely or operated off-sight. The basic types of temperature sensing devices include: • Automatic Reset - The thermal protector automatically restores power after the motor cools. Note: This should not be used where unexpected restarting would be hazardous.
• Manual Reset - Power to the motor is restored by pushing an external button. This type is preferred where unexpected restarts would be hazardous.
• Impedance Protected - The motor is designed to
Technical overload with Number of levels and Category 1 variation (1 digit) function area (2 digits) (3 digits) Only slow (i.e. constant overload) Slow and fast (i.e. constant overload and blocked condition )
Only fast PTC thermistors (i.e. blocked condition)
1 level at cutoff 2 levels at emergency signal and cutoff 1 level at cutoff 2 levels at emergency signal and cutoff 1 level at cutoff
1 2 1 2 1 2 1 2 1 2
Indication of the permissible temperature level when the motor is exposed to thermal overload. Category 2 allows higher temperatures than category 1 does.
Fig 1.4.16: TP designations
50
Thermal switch and thermostats Thermal switches are small bi-metallic switches that change state due to the temperature. They are available with a wide range of trip temperatures; normally open and closed types, with closed being the most common. One or two, in series, are usually fitted in the windings like thermistors and can be connected directly to the circuit of the main contactor coil, requiring no relay. This type of protection is less expensive than thermistors; however, it is less sensitive and is not able to detect a locked rotor failure. Thermal switches are also referred to as Klixon thermal switches and Protection thermal overload (PTO). Thermal
Single-phase motors
According to the IEC 60034-11 standard, the thermal protection (TP) of the motor has to be indicated on the nameplate with a TP designation. Figure 1.4.16 shows an overview of the TP designations.
TP 111 TP 112 TP 121 TP 122 TP 211 TP 212 TP 221 TP 222 TP 311 TP 312
when it reaches its trip temperature.
switches always carry a TP111 designation.
protect itself under locked rotor (stalled) conditions, in accordance with UL standards.
Symbol
Positive Temperature Coefficient (PTC thermistors) can be fitted into the windings of a motor during production or afterwards. Usually three PTCs are fitted in series; one in each phase of the winding.They can be purchased with trip temperatures ranging from 194°F to 356°F. PTCs have to be connected to a thermistor relay which detects the rapid increase in resistance of the thermistor
Single-phase motors normally come with thermal protection. Thermal protection usually has an automatic reclosing. This implies that the motor has to be connected to the main voltage supply in a way to ensure that accidents caused by the automatic reclosing are avoided.
Three-phase motors Three-phase motors have to be protected according to local regulations. This kind of motor usually has contacts for resetting in the external control circuit.
Space Heater
The fixed bearing in the drive end can be a deep-groove ball bearing or an angular contact bearing.
A heating element ensures the standby heating of the motor and is used with applications that struggle with humidity and condensation. By using the space heater, the motor is warmer than the surroundings, and thereby, the relative air humidity inside the motor is always lower than 100%.
Bearing clearances and tolerances are stated according to ISO 15 and ISO 492. Because bearing manufacturers must fulfill these standards, bearings are internationally interchangeable.
Fig 1.4.17: Space heater
In order to rotate freely, a ball bearing must have a certain internal clearance between the raceway and the balls. Without this internal clearance, the bearings can be difficult to rotate or they may seize up and be unable to rotate. Conversely, too much internal clearance will result in an unstable bearing that may generate excessive noise or allow the shaft to wobble. Depending on the pump type to which the motor is fitted, the deep-groove ball bearing in the drive end must have C3 or C4 clearance. Bearings with C4 clearance are less heat sensitive and have increased axial load-carrying capacity. The bearing carrying the axial forces of the pump can have C3 clearance if:
Maintenance The motor should be checked at regular intervals. It is important to keep the motor clean to ensure adequate ventilation. If the pump is installed in a dusty environment, the pump must be cleaned and checked regularly.
• The pump has completeor partial hydraulic relief • The pump has manybrief periods ofoperation • The pump has longidle periods C4 bearings are used for pumps with fluctuating high axial forces. Angular contact bearings are used if the pump exerts strong one-way axial forces.
N o n -d r i ve n d
D ri v e n d
Bearings There are several types of bearing designs. Normally, motors have a locked bearing in the drive end and a bearing with axial play in the non-drive end. Axial play is required due to production tolerances, thermal expansion during operation, and other factors. The motor bearings are held in place by wave spring washers in the non-drive end, see figure 1.4.18.
Spr i n gwa sh e r
Non-drive end bearing D ri vee n dbea r i n g
Fig 1.4.18: Cross-sectional drawing of motor
51
Section 1.4 Motors
Axialforces
Bearingtypesandrecommendedclearance Drive nd
Moderate to strong forces. Primarily outward pull on the shaft end Strong outward pull on the shaft end Moderate forces. Primarily outward pull on the shaft end (partly hydraulically relieved in the pump)
Non-drivend
Fixed deep-groove ball bearing (C4)
Deep-groove ball bearing (C3)
Deep-groove ball bearing (C3) Fixed angular contact bearing Fixed deep-groove ball bearing (C3)
Deep-groove ball bearing (C3)
Small forces (flexible coupling)
Fixed deep-groove ball bearing (C3)
Deep-groove ball bearing (C3)
Strong inward pressure
Deep-groove ball bearing (C4)
Fixed angular contact bearing
Fig:1.4.19: Typical types of bearings in pump motors
Motors with permanently lubricated bearings For closed permanently lubricated bearings, one of the following high temperature resistant types of grease are normally used: • Lithium-based grease • Polyurea-based grease
The grease zerks are visible and are easily accessible. The motor is designed so that: • there is a flow of grease around the bearing • new grease enters the bearing • old grease is removed from the bearing
Motors with lubrication system
Motors with lubricating systems are normally labeled on the fan cover and are supplied with a lubricating instruction. Apart from that, instructions are given in the installation and operating instructions.
Many integral size motors have lubricating nipples for the bearings both in the drive end and the nondrive end. This may vary by manufacturer.
The lubricant is often a lithium-based, high temperature grease. The basic oil viscosity must be: • Higher than 50 cSt at 104°F • 8 cSt at 212°F
52
Chapter 1. Design of pumps and motors
Section 1.5: Liquids 1.5.1 Viscous liquids 1.5.2 Non-Newtonian liquids 1.5.3 The impact of viscous liquids on the performance of a centrifugal pump 1.5.4 Selecting the right pump for a liquid with antifreeze 1.5.5 Calculation example 1.5.6 Computer-aided pump selection for dense and viscous liquids
53
Section 1.5 Liquids
1.5.1 Viscous liquids While water is the most common liquid that pumps handle, in a number of applications, pumps have tohandle other types of liquids, e.g. oil, propylene glycol, gasoline. Compared to water, these types of liquids have different densities and viscosities. Viscosity is a measure of the resistance of a substance to flow. The higher the viscosity, the more difficult the liquid will flow on its own. Propylene glycol and motor oil are examples of thick or high viscous liquids. Gasoline and water are examples of thin, low viscous liquids. Two kinds of viscosities exist: • The dynamic viscosity (μ), which is normally measured in Poise (1 Poise) •
ν=
The kinematic viscosity ν( ), which is normally measured in centiStokes (cSt)
μ ρ
ρ = density of liquid
The relationship between the dynamic viscosityμ() and the kinematic viscosity (ν) is shown in the formula at right. On the following pages, we will focus on kinematic viscosity (ν). The viscosity of a liquid changes considerably with the change in temperature; hot oil is thinner than cold oil. As you can tell from figure 1.5.1, a 50% propylene glycol liquid increases its viscosity 10 times when the temperature changes from +68 to –4oF. For more information concerning liquid viscosity, go to Appendix K.
54
Liquid
Water Gasoline Olive oil 50% Propylene glycol 50% Propylene glycol
Liquid Density temperature ρ [lb/ft3] t [°f] 68 68 68 68 -4
62.4 45.75 56.18 65.11 66.23
Kinematic viscosity ν [cSt] 1.004 0.75 93 6.4 68.7
Fig. 1.5.1: Comparison of viscosity values for water and a few other liquids. Density values and temperatures are also shown
1.5.2 Non-Newtonian liquids The liquids discussed so far are referred to as Newtonian fluids. The viscosity of Newtonian liquids is not affected by the magnitude and the motion that they are exposed to. Mineral oil and water are typical examples of this type of liquid. On the other hand, the viscosity of non-Newtonian liquids does change when agitated. A few examples of non-Newtonion liquids include: • Dilatant liquids, like cream, exhibit a viscosity increase when agitated • Plastic fluids, like ketchup, have a yield value which
must be exceeded before the flow starts. From that point on, the viscosity decreases with an increase in agitation • Thixotropic liquids, like non-drip paint, exhibit a
freezing. When glycol or a similar antifreeze agent
is added to the pumped liquid, the liquid obtains properties different from those of water. The liquid will have a: • • • • • • •
Lower freezing point, tf [°F] Lower specific heat, cp [btu/lbm °F] Lower thermal conductivity, λ [btu ft/h ft2 °F] Higher boiling point, tb [°F] Higher coefficient of expansion, β [ft/°F] Higher density, ρ [lb/ft3] Higher kinematic viscosity, ν [cSt]
These properties must be considered when designing a system and selecting pumps. As mentioned, the higher density requires increased motor power and the higher viscosity reduces pump head, flow rate and efficiency resulting in a need for increased motor power, see figure 1.5.2.
H, P, η
P
decrease in viscosity with an increase in agitation The non-Newtonian liquids are not covered by the viscosity formula described earlier in this section.
H
1.5.3 The impact of viscous liquids on the performance of a centrifugal pump Liquid with higher viscosity and/or higher density than water affects the performance of centrifugal pumps in different ways:
η
Q
Fig. 1.5.2: Changed head, efficiency and power input for liquid with higher viscosity
• Power consumption increases, i.e. a larger motor
may be required to perform the same task • Head, flow rate and pump efficiency are reduced
For example, when a pump is used for pumping a liquid in a cooling system with a liquid temperature below 32oF, an antifreeze agent like propylene glycol is added to the water to prevent the liquid from
55
Section 1.5 Liquids
1.5.4 Selecting the right pump for a liquid with antifreeze Pump characteristics are usually based on water temperature at around 68°F, i.e. a kinematic viscosity of approximately 1 cSt and is 1.0 specific gravity. When pumps are used for liquids containing antifreeze below 32°F, it is necessary to determine, most importantly, that the pump can meet the required performance or if a larger motor is required. The following section presents a simplified method used to determine pump curve corrections for pumps in systems that must handle liquids with a viscosity between 5 cSt - 100 cSt and (specific gravity of 1.0). Please notice that this method is not as precise as the computer-aided method described later in this section.
Pump curve corrections for pumps h andling high viscous liquid Based on knowledge about required duty point, flow (QS,), head (HS,) and kinematic viscosity of the pumped liquid, the correction factors of H and P 2 can be found, see figure 1.5.3. To get the correction factor for multistage pumps, the head of one stage has to be used.
56
Fig. 1.5.3: It is possible to determine the correction factor for head and power consumption at different flow, head and viscosity values
Figure 1.5.3 is read in the following way:
H
When kH and kP2 are found i n the figure, the equivalent head for clean water H W and the corrected actual shaft power P2S can be calculated by the following formula
Hw
H =k .H w
H
Water
2
S
Hs
1
Mixture
H W = kH . H S ρ P2S = kP2 . P2w . s
( )
Qs
ρw
Q
3
P
where HW : is the equivalent head of the pump if the pumped liquid is “clean” water P2W : is the shaft power at the duty point (QS,HW) when the pumped liquid is water HS : is the desired head of the pumped liquid with agents P2S : is the shaft power at the duty point (Q s,Hs) for the viscous pumped liquid water (with agents) ρs : is the specific gravity of the pumped liquid ρw : is the specific gravity of water = 1.0
The pump selection is based on the normal data sheets/curves applying to water. The pump should cover the duty point flow and head, and the motor should be powerful enough to handle the power input on the shaft. Figure 1.5.4 shows how to proceed when selecting a pump and testing whether the motor is within the power range allowed.
P2s P
2S
=K
P2
.P
2w
.
( ) ρs ρw
P2w
5
4
Mixture
Water
Q
Fig. 1.5.4: Pump curve correction when choosing the right pump for the system
The pump and motor selection procedure contains the following steps: • Calculate the corrected head H w (based on Hs and kH ), see figure 1.5.4 lines 1 and 2 • Choose a pump capable of providing performance according to the corrected duty point (Q s, Hw) • Read the power input P2w at the duty point (Qs,Hw), see figure 1.5.4 lines 3 and 4 • Based on P2w, kp2, ρw, and ρs calculate the corrected required shaft power P 2s, see figure 1.5.4, lines 4 and 5 2s 2 max • the Check if Pthe ismotor less than of the motor. Ifselect that isa case, can Pbe used. Otherwise more powerful motor
• Ensure NPSHr < NPSHa
57
Section 1.5 Liquids
1.5.5 Calculation example A circulator pump in a refrigeration system is to pump a 40% (weight) propylene glycol liquid at 14°F. The desired flow is QS = 260 GPM, and the desired head is HS = 40 ft. If the required duty point is known, it is possible to find the QH-characteristic for water and choose a pump to cover the duty point. Once the pump type and size is determined, the pump is fitted with a motor which can handle the specific pump load. The liquid has a kinematic viscosity of 20 cSt and a specific gravity of 65.48 lb/ft3. With QS = 260 GPM, HS = 40 ft and ν = 20 cSt, the correction factors can be found in figure 1.5.3.
1.5.6 Computer-aided pump selection for dense and viscous liquids Some computer-aided pump selection tools include a feature that compensates for the pump performance curves based on input of the liquid density and viscosity. Figure 1.5.5 shows the pump performance curves from the example at left. The figure shows both the performance curves for the pump when it handles viscous liquid (the full lines) and the performance curves when it handles water (the broken lines). As indicated, head, flow and efficiency are reduced resulting in an increase in power consumption. The value of P 2 is 4.5 hp which corresponds to the result as shown in the calculation example in section 1.5.4.
kH = 1.03 kP2 = 1.15 HW = k H · H S = 1.03 · 12 = 40 ft
η
H
[ft]
QS = 260 GPM
[%]
60 50
The pump selection has to cover a duty point equivalent to Q,H = 260 GPM, 40 ft. Once the necessary pump size is selected, the P2 value for the duty point is determined, which in this case is P 2W = 3.8 hp. It is now possible to calculate the required motor power for the propylene glycol mixture:
40
20
70 60 50 40
10
20
30
30 10 0 0
P2 [hp]
50
100
15 0
200
250
3 00
350
400 Q [GPM]0 NPSH (ft)
6 4
P2S = kP2 . P2w .
ρS ρw
2 0 Q [GPM]
Fig. 1.5.5: Pump performance curves
1049 P2S = 1.15 . 3.8 . = 4.6 hp 998 The calculation shows that the pump has to be fitted with a 5 hp motor, which is the smallest motor size able to cover the calculated P 2S = 4.6 hp.
58
Chapter 1. Design of pumps and motors
Section 1.6: Materials 1.6.1 1.6.2 1.6.3 1.6.4 1.6.5
What is corrosion? Types of corrosion Metal and metal alloys Ceramics Plastics
1.6.6 Rubber 1.6.7 Coatings
Section 1.6 Materials
This section discusses the different materials used for pump construction, including the features that every single metal and metal alloy have to offer. Corrosion will be defined, and the different types will be identified, as well as what can be done to prevent corrosion from occurring.
1.6.1 What is corrosion? Corrosion is usually referred to as the degradation of the metal by chemical or electrochemical reaction with its environment, see figure 1.6.1. Considered broadly, corrosion may be looked upon as the tendency of the metal to revert to its natural state similar to the oxide from which it was srcinally melted. Only precious metals, such as gold and platinum, are found in nature in their metallic state. Some metals produce a tight protective oxide layer on the surface which hinders further corrosion. If the surface layer is broken, it is self-healing. These metals are passivated. Under atmospheric conditions, the corrosion products of zinc and aluminum form a fairly tight oxide layer and further corrosion is prevented. Likewise, on the surface of stainless steel, a tight layer of iron and chromium oxide is formed, and on the surface of titanium, a layer of titanium oxide is formed. The protective layers of these metals demonstrate their good corrosion resistance. Rust, on the other hand, is a non-protective corrosion product on steel. Rust is porous, not firmly adherent and does not prevent continued corrosion, see figure 1.6.2.
Environmental variables that affect the corrosion resistance of metals and alloys pH (acidity) Oxidizing agents (such as oxygen) Temperature Concentration of solution constituents (such as chlorides) Biological activity Operating conditions (such as velocity, cleaning procedures and shutdowns)
Fig. 1.6.1: Environmental variables that affect the corrosion resistance of metals and alloys
Rust on steel
Non-protective corrosion product Oxide layer on stainless steel
Protective corrosion product Fig. 1.6.2: Examples of corrosion products
60
1.6.2 Types of corrosion Generally, metallic corrosion involves the loss of metal at a spot on an exposed surface. Corrosion occurs in various forms ranging from uniform attacks over the entire surface to severe local attacks. The environment’s chemical and physical conditions determine both the type and the rate of corrosion attacks. The conditions also determine the type of corrosion products that are formed and the control measures that must be taken. In many cases, it is impossible or rather expensive to completely stop the corrosion process; however, it is usually possible to control the severity to acceptable levels. On the following pages, different forms of corrosion and their characteristics will be discussed.
Uniform corrosion Uniform or general corrosion is characterized by corrosive attacks spreading evenly over the entire surface or on a large part of the total area. General thinning continues until the metal is broken down. Uniform corrosion results in waste of most of the metal.
Fig. 1.6.3: Uniform corrosion
Examples of metals subject to uniform corrosion include: • Steel in aerated water • Stainless steel in reducing acids [such as AISI 304 (EN 1.4301) in sulfuric acid]
Pitting corrosion Pitting corrosion is a localized form of a corrosive attack. Pitting corrosion forms holes or pits on the metal surface. It perforates the metal while the total corrosion, measured by weight loss, might be rather minimal. The rate of penetration may be 10 to 100 times that of general corrosion depending on the
Fig. 1.6.4: Pitting corrosion
aggressiveness of the medium. Pitting occurs more often in a stagnant environment. An example of metal subject to pitting corrosion: • Stainless steel in seawater
61
1. Design Sectionof 1.6pumps and motors Materials 1.1 Pump construction, (10)
Crevice corrosion Crevice corrosion, like pitting corrosion, is a localized form of corrosion attack. However, crevice corrosion is more aggressive. Crevice corrosion occurs at narrow openings or spaces between two metal surfaces or between metals and non-metal surfaces and is usually associated with a stagnant condition in the crevice. Crevices, such as those found at flange joints or at threaded connections, are often the most critical spots for corrosion.
Fig. 1.6.5: Crevice corrosion
An example of metal subject to crevice corrosion: • Stainless steel in seawater
Intergranular corrosion Intergranular corrosion occurs at grain boundaries. Intergranular corrosion, also called intercrystalline corrosion, typically occurs when chromium carbide precipitates at the grain boundaries during the welding process or in connection with insufficient heat treatment. A narrow region around the grain boundary may become deplete in chromium and become less resistant to corrosion than the rest of the material. This is unfortunate because chromium plays an important role in corrosion resistance.
Fig. 1.6.6: Intergranular corrosion
Examples of metals subject tointergranular corrosion include: • Insufficiently welded or heat-treated stainless steel • Stainless steel AISI 316 (EN 1.4401) in nitric acid
Selective corrosion
Brass
Selective corrosion attacks one single element of an alloy and dissolves the element in the alloy structure. Consequently, the alloy’s structure is weakened. Examples of selective corrosion: • The dezincification of unstabilized brass producing a weakened, porous copper structure • Graphitization of gray cast iron leaving a brittle graphite skeleton due to the dissolution of iron.
62
Zinc corrosion products Copper
Fig. 1.6.7: Selective corrosion
Erosion corrosion Erosion corrosion is a process whereby the rate of corrosion attack is accelerated by the relative motion of a corrosive liquid and a metal surface. The attack is localized in areas with high velocity or turbulent flow. Erosion corrosion attacks are characterized by grooves with a directional pattern.
Examples of metals subject to erosion corrosion: • Bronze in seawater • Copper in water
Flow
Fig. 1.6.8: Erosion corrosion
Cavitation corrosion Cavitation corrosion occurs when a pumped liquid with high velocity reduces the pressure, and it drops below the liquid vapor pressure forming vapor bubbles. In the areas where the vapor bubbles form, the liquid boils. When the pressure rises again, the vapor bubbles collapse and produce intensive shockwaves. Consequently, the collapse of the vapor bubbles remove metal or oxide from the surface. Examples of metals that are subject to cavitation: • Cast iron in water at high temperature • Bronze in seawater
Fig. 1.6.9: Cavitation corrosion
Stress corrosion cracking Stress corrosion cracking (SCC) refers to the combined influence of tensile stress (applied or internal) and corrosive environment. The material can crack without any significant deformation or obvious deterioration of the material. Often, pitting corrosion is associated with SCC. Examples of metals that are subject to SCC: • Stainless steel AISI 316 (EN 1.4401) in chlorides • Brass in ammonia
Fig. 1.6.10: Stress corrosion cracking
63
1. Design Sectionof 1.6pumps and motors Materials 1.1 Pump construction, (10)
Corrosion fatigue < Pure mechanical fatigue occurs when a material subjected to a cyclic load far below the ultimate tensile strength fails. If the metal is simultaneously exposed to a corrosive environment, the failure can take place at an even lower stress and after a shorter period of time. Contrary to a pure mechanical fatigue, there is no fatigue limit in corrosion-assisted fatigue.
Fig. 1.6.11: Corrosion fatigue
An example of a metal subject to corrosion fatigue: • Aluminium structures in a corrosive atmosphere
Galvanic corrosion Galvanic corrosion occurs when a corrosive electrolyte and two metallic materials are in contact (galvanic cell) and corrosion increases on the least noble material (the anode) and decreases on the noblest material (the cathode). The tendency of a metal or an alloy to corrode in a galvanic cell is determined by its position in the galvanic series. The galvanic series indicates the relative nobility of different metals and alloys in a given environment (e.g. seawater, see figure 1.6.13).The farther apart the metals are in the galvanic series, the greater the galvanic corrosion effect will be. Metals or alloys at the upper end are more noble than those at the lower end.
Aluminium -lessnoble Fig. 1.6.12: Galvanic corrosion
Copper-mostnoble
Examples of metals that are subject to galvanic corrosion include: • Steel in contact with AISI 316 (EN 1.4401) • Aluminum in contact with copper The principles of galvanic corrosion are used in cathodic protection. Cathodic protection is the reduction or prevention of the corrosion of a metal surface through the use of sacrificial anodes (zinc or aluminum) or impressed currents. Fig. 1.6.13: Galvanic series for metals and alloys in seawater
64
1.6.3 Metal and metal alloys On the following pages, the features of different metals and metal alloys used for construction of pumps are discussed.
Ferrous alloys
Cavitation corrosion of bronze impeller
Ferrous alloys are alloys where iron is the prime constituent. Ferrous alloys are the most common of all materials because of their availability, low cost, and versatility.
Steel Erosion corrosion of cast iron impeller
Pitting corrosion of AISI 316 (EN 1.4401)
Steel is a widely used material primarily composed of iron alloyed with carbon. The amount of carbon in steel varies in the range from 0.003% to 1.5% by weight. The content of carbon has an important impact on the material’s strength, weldability, machinability, ductility, and hardness. Generally, an increase in carbon content will lead to an increase in strength and hardness but to a decrease in ductility and weldability. The most common type of steel is carbon steel. Carbon steel is grouped into four categories, see figure 1.6.14. Steel is available in wrought and cast grades. Cast steel is closely comparable to wrought; both are relatively inexpensive to make, form, and process but have low corrosion resistance compared to alternative materials such as stainless steel.
0.0394 inch
Intergranular corrosion of stainless steel
Crevice corrosion of SAF 2205 (EN 1.4462)
Type of steel
Content of carbon
Low carbon or mild steel
0.003% to 0.30% of carbon
Medium carbon steel
0.30% to 0.45% of carbon
High carbon steel
0.45% to 0.75% of carbon
Very high carbon steel
0.75% to 1.50% of carbon
Fig 1.6.14: Four types of carbon steel
65
1. Design Sectionof 1.6pumps and motors Materials 1.1 Pump construction, (10)
Cast iron
Nodular (ductile) iron
Cast iron is an alloy of iron, silicon and carbon. Typically, the concentration of carbon is between 3-4% by weight, most of which is present in insoluble form (e.g. graphite flakes or nodules). The twomain types are grey cast iron and nodular (ductile) cast iron. The corrosion resistance of cast iron is comparable tothat of steel; and sometimes even better. Cast iron can be alloyed with 13-16% (by weight) silicon or 15-35% (by weight) nickel (Ni-resist) to improve corrosion resistance. Various types of cast irons are widely used in industry, especially for valves, pumps, pipes and automotive parts. Cast iron has good corrosion resistance to neutral and alkaline liquids (h igh pH) but has poor resistance to acids (low pH).
Nodular iron contains around 0.03-0.05% (by weight) of magnesium. Magnesium causes the flakes to become globular, so the graphite is dispersed throughout aferrite or pearlite matrix in the form of spheres or nodules. The round shape of nodular graphite reduces the stress concentration and consequently, the material is much more ductile than grey iron. Figure 1.6.16 shows that the tensile strength is higher for nodular iron than for grey iron. Nodular iron is normally used for pump parts with high strength requirements (high pressure or high temperature applications).
Stainless steel Grey iron In grey iron, the graphite is dispersed throughout a ferrite or pearlite matrix in the form of flakes. Fracture surfaces take on a grey appearance (hence the name). The graphite flakes act as stress concentrators under tensile loads making the material weak and brittle in tension, but strong and ductile in compression. Grey iron is used for the construction of motor blocks because of its high vibration damping ability. Grey iron is an inexpensive material and is relatively easy to cast with a minimal risk of shrinkage. That is why grey iron is often used for pump parts with moderate strength requirements.
Stainless steel is composed of chromium containing steel alloys. The minimum chromium content in standardized stainless steel is 10.5%. Chromium improves the corrosion resistance of stainless steel. This is due to a chromium oxide film that is formed on the metal surface. This extremely thin layer is self-repairing under the right conditions. Molybdenum, nickel and nitrogen are other examples of typical alloying elements. Alloying with these elements brings out different crystal structures which enable different properties in connection with machining, forming, welding and corrosion resistance. In general, stainless steel has a higher resistance to chemicals (i.e. acids) than steel and cast iron.
0
ASTM
150 172
EN-GJL-150
-
-
G G-1 5
50
-
200
EN-GJL-200
207
-
-
400
EN-GJS-400-18
GGG-40
40 0-1 8
400
EN-GJS-400-15
GGG-40.3
400-15
-
-
-
450-10
-
-
A 48 Gr 25A
GG-2 0
-
200
-
-
430 450
241
E N - G J L - 2 50
GG -25
2 50
A 48 Gr 35A
-
500 575
Fig 1.6.15: Comparison and designations of grey iron
66
EN-GJS-450-10
A 48 Gr 30A 460
250
-
EN-GJS-500-7
-
GGG-50
-
ASTM
A 536 Gr 60-40-18
A 536 Gr 65-45-12
-
500-7
-
A 536 Gr 80-55-06
Fig 1.6.16: Comparison and designations of nodular iron
In environments containing chlorides, stainless steel can be attacked by localized corrosion, such as pitting corrosion and crevice corrosion. The resistance of stainless steel to these types of corrosion is highly dependent on its chemical composition. It is common to use the so-called Pitting Resistance Equivalent (PRE) values as a measure of pitting resistance for stainless
resistance is taken into consideration. The higher the PRE, the higher the resistance to localized corrosion. Be aware that the PRE value isa rough estimate of the pitting resistance of a stainless steel and should only be used for comparison/classification of different types of stainless steel. To follow, the four major types of stainless steel: ferritic, martensitic, austenitic and duplex are presented.
steel. PRE values are calculated by formulas where the relative influence of a few alloying elements (chromium, molybdenum and nitrogen) on the pitting Fig 1.6.17: Chemical composition of stainless steel
Chemical composition of stainless steel [w%] Microstructure
Designation % EN/AISI/UNS Carbonmax.
Nickel
Ferritic
1.4016/430/ S43000
0.08
16-18
Martensitic
1.4057/431/ S43100
0.12-0.22
15-17
% % Molybdenum
%
PRE 5) 17
1.4305/303/ S30300
0.1
17-19
8-10
Austenitic
1.4301/304/ S30400
0.07
17-19.5
8-10.5
Austenitic
1.4306/304L/ S30403
0.03
18-20
10-12
Austenitic
1.4401/316/ S31600
0.07
16.5-18.5
10-13
Austenitic Austenitic S31635
1.4404/316L/ S31603 1.4571/316Ti/
0.03 0.08
16.5-18.5 16.5-18.5
Austenitic
1.4539/904L/ N08904
0.02
19-21
24-26
Austenitic
1.4547/none /
0.02
20
18
S 31254
Other
1.5-2.5
Austenitic
16 S 0.15-0.35
18 18 18
2-2.5
24
10-13 2-2.5 10.5-13.5 2-2.5 Ti < 0.70
Ti > 5 x carbon
24 24
4-5
Cu 1.2-2
34
6.1
N 0.18-0.22 Cu 0.5-1
43
3)
Ferritic/ austenitic
1.4462/ none/ S32205 2)
0.03
21-23
4.5-6.5
2.5-3.5
N 0.10-0.22
34
Ferritic/ austenitic
1.4410/none/ S 32750 4)
0.03
25
7
4
N 0.24-0.32
43
Microstructure
Designation EN/ASTM/UNS
% Carbon max.
% Chromium
% Nickel
% Molybdenum
% Other
PRE
18-20
8-11
1) Austenitic 1.4308/CF8/ J92600
0.07
1) Austenitic1.4408/CF8M/ J92900
0.07
Austenitic 1) 1.4409/CF3M/ J92800
1)
% Chromium
18-20
0.03
19
9-12
18-20
2-2.5
9-12
26
2-2.5
N max. 0.2
35
2.5-3.5
N 0.12-0.2
35
2.5-3.5
N 0.12-0.22 Cu 2.75-3.5
38
1.4584/none/ none
0.025
19-21
24-26
4-5
Ferritic/ austenitic
1.4470/CD3MN/ J92205
0.03
21-23
4.5-6.5
Ferritic/ austenitic
1.4517/CD4MCuN/ J93372
0.03
24.5-26.5
Contains some ferrite 2) Also known as SAF 2205, 3) Also known as 254 SMO, 5) Pitting Resistance Equivalent (PRE): Cr% + 3.3xMo% + 16xN%.
2.5-3.5 4)
26
N max. 0.2 Cu 1-3
Austenitic
Also known as SAF 2507
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1. Design Sectionof 1.6pumps and motors Materials 1.1 Pump construction, (10)
Ferritic (magnetic) Ferritic stainless steel is characterized by good corrosion properties, resistance to stress corrosion cracking, and moderate toughness. Low alloyed ferritic stainless steel is used in mild environments (teaspoons, kitchen sinks, washing machine drums, etc.) where maintenance-free and non-rusting is required.
and AISI 316 (EN 1.4401), see illustration 1.6.17.
Martensitic stainless steel is characterized by high strength and limited corrosion resistance. Martensitic steels are used for springs, shafts, surgical instruments and for sharp-edged tools, such as knives and scissors.
The stabilized grade AISI 316Ti (EN 1.4571) contains a small amount of titanium. Because titanium has a higher affinity for carbon than chromium, the formation of chromium carbides is minimized. The content of carbon is generally low in modern stainless steel, and with the easy availability of ‘L’ grades, the use of stabilized grades has declined significantly.
Austenitic (non-magnetic)
Ferritic-austenitic or duplex (magnetic)
Austenitic stainless steel is the most common type of stainless steel and is characterized by a high corrosion resistance, good formability, toughness and weldability. Austenitic stainless steel, especially
Ferritic-austenitic (duplex) stainless steel is characterized by strength, toughness, high corrosion resistance and excellent resistance to stress corrosion cracking and corrosion fatigue. Ferritic-austenitic
the AISI 304 and AISI 316, are used for almost any type of pump components. This kind of stainless steel can be either wrought or cast.
stainless steel is typically used in applications that require high strength, high corrosion resistance and low susceptibility to stress corrosion crackingor a combination of these properties. Stainless steel SAF 2205 is widely used for making pump shafts and pump housings.
Martensitic (magnetic)
AISI 303 is one of the most popular stainless steel types of all the free machining stainless steel types. Due to its high sulphur content (0.15-0.35 w%), the machinability improves considerably but corrosion resistance and weldability decrease. Over the years, free machining grades with a low sulphur content and a higher corrosion resistance have been developed. If stainless steel is heated up to 932°F - 1472°F for a relatively long period of time during welding, the chromium may form chromium carbides with the carbon in the steel. This reduces chromium’s capability to maintain the passive film and may lead to intergranular corrosion, also referred to as sensitization (see section 1.6.2).
68
If low carbon grades of stainless steel are used, the risk of sensitization is reduced. Stainless steel with a low content of carbon is referred to as AISI 316L (EN 1.4306), or AISI 304L (EN 1.4404). Both grades contain 0.03% of carbon compared to 0.07% in the regular type of stainless steel, AISI 304 (EN 1.4301)
Nickel alloys
Copper alloys
Nickel based alloys are defined as alloys in which nickel is present in greater proportion than any other alloying element. The most important alloying constituents are iron, chromium, copper, and molybdenum. The alloying constituents make it possible to form a wide range of alloy classes. Nickel and nickel alloys have the ability to withstand a wide variety of severe operating conditions, including corrosive environments, high temperatures, high stresses or a combination of these factors.
Pure copper has excellent thermal and electrical properties but is a very soft and ductile material. Alloying additions result in different cast and wrought materials suitable for use in the production of pumps, pipelines, fittings, pressure vessels and for many marine, electrical and general engineering applications.
HastelloyTM alloys are commercial alloys containing nickel, molybdenum, chromium, and iron. Nickel based alloys - such as InconelTM Alloy 625, HastelloyTM C-276 and C-22 - are corrosion resistant, not subject to pitting or crevice corrosion in low velocity seawater and do not suffer from erosion at high velocity. The price of nickel based alloy limits its use in certain applications. Nickel alloys are available in both wrought and cast grades. However, nickel alloys are more difficult to cast than the common carbon steels and stainless steel alloys. Nickel alloys are often used for pump parts in the chemical process industry.
1) Lead can be added as an alloying element to improve machinability. 2) Bronze can be alloyed with aluminium to increase strength. Fig 1.6.18: Common types of copper alloys
Brasses are the most widely used of the copper alloys because of their low cost and easy or inexpensive fabrication and machining. However, they are inferior in strength to bronzes and must not be used in environments that cause dezincification. Red brass, bronze and copper nickels, compared to cast iron, have a high resistance to chlorides in aggressive liquids, such as seawater. In such environments, brass is unsuitable because of its tendency to desincificate. All copper alloys have poor resistance to alkaline liquids (high pH), ammonia, and sulfides and are sensitive to erosion. Brass, red brass and bronze are widely used for making bearings, impellers and pump housings.
69
1. Design Sectionof 1.6pumps and motors Materials 1.1 Pump construction, (10)
Aluminum
Titanium
Pure aluminum is a light and soft metal with a density of about a third of that of steel. Pure aluminum has a high electrical and thermal conductivity. The most common alloying elements are silicon (silumin), magnesium, iron and copper. Silicon increases the material’s castability, copper increases its machinability, and magnesium increases its corrosion resistance and strength.
Pure titanium has a low density, is quite ductile and has a relatively low strength. When a limited amount of oxygen is added, it will strengthen titanium and produce commercial-pure grades. Additions of various alloying elements, such as aluminum and vanadium, increase its strength significantly but at the expense of ductility. The aluminum and vanadium-alloyed titanium (Ti-6Al-4V) is the “workhorse” alloy of the titanium industry. It is used in many aerospace engine and airframe components. Because titanium is a high-price material, it is seldom used for making pump components.
An advantage of aluminum is its ability to generate a protective oxide film that is highly corrosion resistant if it is exposed to the atmosphere. Treatment, such as anodizing, can further improve this property. Aluminum alloys are widely used in structures where a high strength to weight ratio is important, such as in the transportation industry. The use of aluminum in vehicles and aircrafts reduces weight and energy consumption. A disadvantage of aluminum is its instability at low or high pH or in chloride-containing environments. This property makes aluminum unsuitable for exposure to aqueous solutions, especially under conditions with high flow.
Titanium is a reactive material. Like stainless steel, titanium’s corrosion resistance depends on the formation of an oxide film. Titanium’s oxide film is more protective than stainless steel’s. Therefore, titanium performs much better than stainless steel in aggressive liquids, such as seawater, wet chlorine or organic chlorides, where pitting and crevice corrosion can occur.
This is further emphasized by the fact that aluminum is a reactive metal, i.e. has a low position in the galvanic series and may easily suffer from galvanic corrosion if coupled to nobler metals and alloys (see section on galvanic corrosion pg. 64). Designation
Major alloying element CP: commercial pure (titanium content above 99.5%)
1000-series
Unalloyed (pure) >99% Al
2000-series
Copper is the principal alloying element, though other elements (magnesium) may be specified
3000-series
Manganese is the principal alloying element
4000-series
Silicon is the principal alloying element
5000-series 6000-series
Magnesium is the principal alloying element Magnesium and silicon are principal alloying elements
7000-series 8000-series
Zinc is the principal alloying element, but other elements, such as copper, magnesium, chromium, and zirconium may be specified
Other elements (including tin and some lithium compositions)
Fig 1.6.19: Major alloying elements of aluminum
70
Fig 1.6.20: Titanium grades and alloy characteristics
1.6.4 Ceramics
Thermoplastics
Ceramic materials are composed of metallic and non-metallic elements and are typically crystalline in nature. Common technical ceramics are aluminum oxide (alumina - Al2O3), silicon carbide (SiC), tungsten carbide (WC), and silicon nitride (Si3N4). Ceramics are suitable for applications requiring high thermal stability, strength, wear resistance, and corrosion resistance. Disadvantages of ceramics include low ductility and high tendency for brittle fractures. Ceramics are mainly used for making bearings and seal faces for shaft seals.
Thermoplastic polymers consist of long polymer molecules that are not cross-linked to each other. They are often supplied as granules and heated to permit fabrication by methods such as molding or extrusion. A wide range is available from low-cost commodity plastics (e.g. PE, PP, PVC) to high cost engineering thermoplastics (e.g. PEEK) and chemical resistant fluoropolymers (e.g. PTFE, PVDF). PTFE is one of the few thermoplastics that is not meltprocessable. Thermoplastics are widely used for making pump housings or for lining of pipes and pump housings.
1.6.5 Plastics
Thermosets
Some plastics are derived from natural substances like plants but most types are synthetic. Most synthetic plastics come from crude oil, but coal and natural gas are also used. There are two main types of plastics: Thermoplastics and thermosets (thermosetting plastics), with thermoplastics being the most common used worldwide. Plastics often contain additives which transfer additional properties to the material. Furthermore, plastics can be reinforced with fiberglass or other fibers. These plastics, together with additives and fibers, are also referred to as composites.
Thermosets harden permanently when heated, as cross-linking hinders bending and rotations. Cross-linking is achieved during fabrication using chemicals, heat, or radiation; a process called curing or vulcanization. Thermosets are harder, more dimensionally stable and brittle than thermoplastics and cannot be remelted. Some thermosets include epoxies, polyesters, and polyurethanes. Thermosets are, among other things, used for surface coatings.
Examples of additives found in plastics: • Inorganic fillers for mechanical reinforcement • Chemical stabilizers, e.g. antioxidants • Plasticizers • Flame retardants Abbreviation PP PE PVC PEEK PVDF PTFE*
Linear polymer chains
Branched polymer chains
Elastomers
Polymer name Polypropylene Polyethylene Polyvinylchloride Polyetheretherketone Polyvinylidene fluoride Polytetrafluoroethylene
*Trade name: Teflon®
Fig 1.6.21: Overview of polymer names
Thermoplastics
Weakly cross-linked polymer chains
Thermosets Strongly cross-linked polymer chains
Fig 1.6.22: Different types of polymers
71
1. Design Sectionof 1.6pumps and motors Materials 1.1 Pump construction, (10)
Ethylene-propylelediene rubber
1.6.6 Rubber The term rubber includes both natural rubber and synthetic rubber. Rubbers, also known as elastomers, are flexible long-chain polymers that can be stretched easily to several times their length. Rubbers are cross-linked (vulcanized) but have a low cross-link density, see figure 1.6.22. The crosslink is the key to the elastic or rubbery properties of these materials. The elasticity provides resilience in sealing applications. Different components in a pump are made of rubber, such as gaskets and Orings (see section 1.3 on shaft seals). In this section, the different kinds of rubber qualities and their main properties, in regards to temperature and resistance to different kinds of liquid groups, will be presented.
Nitrile rubber At temperatures up to about 212°F, nitrile rubber (NBR) is an inexpensive material that has a high resistance to oil and fuel. Different grades of nitrile rubber exist - the higher the acetonitrile (ACN) content, the higher the oil resistance but the poorer the low-temperature flexibility. Nitrile rubbers have high resilience and high-wear resistance but only moderate strength. Further, this rubber has limited weathering resistance and poor solvent resistance. It can generally be used at about -22°F, but certain grades can operate at lower temperatures.
of Abbreviation Common Common types ofname copper alloys Examples trade name
Nitrile rubber
EPDM, EPM
Ethylene-propylelediene Fl u o r o e l a s t o m e r s
MQ, VMQ, PMQ, FMQ FFKM
Fig 1.6.23: Rubber types
72
Silicone rubber Perfluoroelastomers
Fluoroelastomers Fluoroelastomers (FKM) cover a whole family of rubbers designed to withstand oil, fuel and a wide range of chemicals including non-polar solvents. FKM offers excellent resistance to high temperatures (up to 392°F depending on the grade) in air and different types of oil. FKM rubbers have limited resistance to steam, hot water, methanol, and other polar fluids. This type of rubber also has poor resistance to amines, strong alkalis and many freons. There are standard and special grades - the latter have improved low-temperature properties or chemical resistance.
Silicone rubber Silicone rubbers (Q) have outstanding properties, such as low compression set in a wide range of temperatures (from -76°F to 392°F in air), excellent electrical insulation and non-toxic. Silicone rubbers are resistant to water, some acids and oxidizing chemicals. Concentrated acids, alkalines and solvents should not be used with silicone rubbers. In general, these rubber types have poor oil and fuel resistance. However, the FMQ silicone rubber resistance to oil and fuel is better than that of types MQ, VMQ, and PMQ.
Buna-N
NBR
FKM
Ethylene propylelediene (EPDM) has excellent water resistance which is maintained to approximately 248-284°F. This rubber type has good resistance to acids, strong alkalis and polar fluids such as methanol and acetone. It has very poor resistance to mineral oil and fuel.
Vi t o n
Nordel® ®
Siloprene® Chemraz® Kalrez®
Perfluoroelastomers Perfluoroelastomers (FFKM) have very high chemical resistance, almost comparable to that of PTFE (polytetrafluorethylene, e.g. Teflon®). They can be used in high temperatures, but their disadvantages are difficult processing, very high cost and limited use at low temperatures.
1.6.7 Coatings Protective coatings such as metallic, non-metallic (inorganic) or organic coatings, are a common method of corrosion control. The main function of coatings, aside from galvanic coatings such as zinc, is to provide a barrier between the metal substrate and its environment. They allow for the use of normal steel or aluminum instead of more expensive materials. In the following section, the possibilities of preventing corrosion by means of different coatings will be examined.
To protect the base steel, zinc coating sacrifices itself slowly by galvanic action.
Steel coated with a more noble metal, such as nickel, corrodes more rapidly if the coating is damaged.
Fig 1.6.24: Galvanic vs. barrier corrosion protection
Metallic coatings There are two types of metallic coatings. One is where the coating is less noble than the substrate, and the other, electroplating, is where a more noble metal is applied to the substrate as a barrier layer.
Metallic coatings less noble than the substrate Zinc coatings are commonly used for the protection of steel structures against atmospheric corrosion. Zinc has two functions. It acts as a barrier coating, and it provides galvanic protection. Should an exposed area of steel occur, the zinc surface preferentially corrodes at a slow rate and protects the steel. The preferential protection is referred to as cathodic protection. When damage is minimal, the protective corrosion products of zinc will fill the exposed area and stop the attack.
Metallic coatings nobler than the substrate Electroplating of nickel and chromium coatings on steel are nobler than the substrate. Unlike galvanic coatings where the coating corrodes near areas where the base metal is exposed, any void or damage in a barrier coating can lead to an immediate base metal attack.
73
1. Design Sectionof 1.6pumps and motors Materials 1.1 Pump construction, (10)
<
Non-metallic coatings (conversion coatings) Conversion coatings are included in non-metallic coatings, also known as inorganic coatings. Conversion coatings are formed by a controlled corrosion reaction of the substrate in an oxidized solution. Examples of conversion coatings are anodizing or chromating of aluminum and phosphate treatment of steel. Anodizing is mainly used for surface protection of aluminum, while chromating and phosphating are usually used for pre-treatment to improve paint adhesion and to help prevent the spreading of rust under layers of paint.
Paints As mentioned, paints are an important class of organic coating. Figure 1.6.25 shows several types of organic coatings. A typical paint formulation contains polymeric binders, solvents, pigments and additives. For environmental reasons, organic solvents are often replaced by water or simply eliminated, as in powder coating. Painted metal structures usually involve two or more layers of coating applied on a primary coating, which is in direct contact with the metal.
Organic coatings Organic coatings contain organic compounds and are available in a wide range of different types. Organic coatings are applied to the metal by methods of spraying, dipping, brushing, liningcurrent). or electro-coating (paint applied by means of electric They may or may not require heat-curing. Both thermoplastic coatings (i.e. polyamide, polypropylene, polyethylene, PVDF and PTFE) and elastomercoatings are applied to metal substrates to combine the mechanical properties of metal with the chemical resistance of plastics, but paints are by far the most widely used organic coating.
74
Physical states of common organic coatings Resin type Acrylic Alkyd Epoxy Polyester Polyurethane Vinyl
Solvent- Water- Powder based based coating X X X X X X
X X X X X
Two comp. liquid
X X X X X
X X X
Fig 1.6.25: Physical states of common organic coatings
Chapter 2. Installation and per formance reading
Section 2.1: Pump installation 2.1.1 2.1.2 2.1.3 2.1.4 2.1.5
New installation Existing installation-replacement Pipe flow for single-pump installation Limitation of noise and vibrations Sound level
Section 2.2: Pump performance 2.2.1 Hydraulic terms 2.2.2 Electrical terms 2.2.3 Liquid properties
Section 2.1 Pump installation
Accuracy of suited pump type for an installation has significant impact on optimum operation. The larger the pumps, the greater the costs with respect to investment, installation, commissioning, operation and maintenance – basically the life cycle costs (LCC). An extensive product portfolio combined with competent advice and after-sales service is the foundation of a proper selection. The following analysis, recommendations and pump tips are general for any installation but, to a greater extent, relevant for medium to large sized installations. Recommendations for new and existing installations follow.
2.1.1 New installation • If the pipework has not been planned, the selection of a pump type can be based on other primary criteria, such as efficiency, investment costs or lifecycle costs (LCC). This will be covered in a later
2.1.2 Existing installation–replacement Tips for optimum pump selection for existing installation follows.
Pre-investigation of the installation should include: • Basic pipe flow – pipes in and out of the building, e.g. from the ground, along the floor or from the ceiling
• Specific pipework at the point of installation, e.g. in-line or end-suction, dimensions, manifolds Space availability – width, depth and height Accessibility for maintenance, i.e., doorways Availability/accessibility of lift equipment Floor type, e.g. solid or suspended floor with basement • Existing foundation • Existing electrical installation
• • • •
Previous pump installation • Pump make, type, specifications including old duty point, shaft seal, materials, gaskets, controlling
• History, e.g. lifetime, maintenance
section.
• If the pipework has been planned, pump selection is equivalent to pump replacement in an existing installation.
Future requirements • Desired improvements and benefits • New selection criteria including duty points and operating times, temperature, pressure, liquid specs
• Supplier criteria, e.g. availability of spare parts Advisory • Major changes might be beneficial in long or short term and should be documented, e.g. installation savings, life cycle costs (LCC), reduced environmenta l impact (noise, vibration accessibility for maintenance)
Selection • Should be based on priorities agreed to by customer For the selection of correct pump type and installation advice, two main areas are important: Pipe flow and limitation of noise and vibrations. This will be dealt with on the following pages.
76
2.1.3 Pipe flow for single-pump installation Figure 2.1.1 is based on single-pump installation. In parallel installations, accessibility plays a major role for suitability of a pump choice. Simple pipework with few bends as possible is the criteria for pump choice in a single-pump installation.
Scores: Best choice Good choice Acceptable choice Not applicable
Pump type
Pipework To the pump:
Along floor
From ground
From ceiling
Wallmounted
A. In-lin e close-co upled (horizontal or vertical mounting)
B. End-suction close- couple d C. End -su cti on long- couple d (horizontal or vertical (only horizontal mounting) mounting)
From the pump: Along floor
Best choice
Good choice
Good choice
To ground
Best choice
Good choice
Good choice
To ceiling
Good choice
Bestchoice
Along floor
Good choice
Best choice
Acceptable choice
To ground
Good choice
Best choice
Acceptable choice
To ceiling
Good choice
Best choice
Best choice
Along floor
Best choice
Acceptable choice
Acceptable choice
To ground
Best choice
Good choice
Good choice
To ceiling
Good choice
Best choice
Best choice
Wallmounted
Best choice
Good choice
Not applicable
Bestchoice
Fig. 2.1.1 Pipework and pump type
77
Section 2.1 Pump installation
Accessibility plays a major role in how well a specific pump choice is suited to an installation of several pumps in parallel. In-line pumps installed in parallel do not always provide the best accessibility because of the pipwork, see figure 2.1.2. End-suction pumps installed in parallel provide better accessibility, see figure 2.1.3.
2.1.4 Limitation of noise and vibrations
Fig. 2.1.2: Three in-line pumps in parallel; limited maintenance access because of pipework
To achieve optimum operation and minimize noise and vibration, vibration dampening of the pump may be necessary. Generally, this should be considered for pumps with motors above 7.5 hp. Smaller motor sizes, however, may also cause noise and vibration due to rotation in the motor and pump and by the flow in pipes and fittings. The effect on the environment depends on correct installation and the condition of the entire system. Three ways to limit noise and vibration in a pump installation are: Foundation considerations, dampeners and expansion joints.
Fig. 2.1.3: Three end-suction pumps in parallel; easier maintenance access because of pipework Floor Solid ground
Foundation Floor constructions can be solid or suspended. Fig. 2.1.4: Solid floor construction
Solid – minimum risk of noise due to low transmission of vibrations, see figure 2.1.4. Suspended – risk of floor amplifying the noise. Basement can act as a resonance box, see figure 2.1.5.
Ground floor
Wall
The pump should be installed on a plane on a rigid surface. There are four basic installations for the two types of floor constructions: Floor, foundation, floating foundation and foundation suspended on vibration dampeners.
Basement
Floor Solid ground
Fig. 2.1.5: Suspended floor construction
78
Floor
Floor Direct mounting on floor, hence direct vibration transmission, see figure 2.1.6.
Fig. 2.1.6: Floor
Floor Base plate Pump unit
Foundation Poured directly on concrete floor, see figure 2.1.7.
Floating foundation Resting on a dead material, e.g. sand, hence reduced risk of transmitting vibration, see figure 2.1.8.
Foundation suspended on vibration dampeners Optimum solution with controlled vibration transmission, see figure 2.1.9. The weight of a concrete foundation should be 1.5 x the pump weight. This weight is needed to get the dampeners to work efficiently at low pump speed.
Fig. 2.1.7: Foundation
Floor
Foundation Base plate Pump unit
Floor Sand
Foundation Base plate Pump unit
Fig. 2.1.8: Floating foundation
Fig. 2.1.9: Foundation suspended on vibration dampeners
Floor Vibration dampeners Foundation Base plate Pump unit
Pump unit
Fig. 2.1.10: The same foundation rules apply to vertical in-line pumps
Foundation Vibration dampeners Floor
79
Section 2.1 Pump installation
Dampeners Vibration dampener selection requires the following data:
pressure loss on the pressure side. At high water velocities (16.4 ft/s or greater), it is best to install larger expansion joints, corresponding to the pipework.
• Forces acting on the dampener • Motor speed with consideration of speed control • Required dampening in % (suggested value is 70%) Dampener selection varies from installation to installation. An incorrect selection may increase the vibration level. The supplier should, therefore, size vibration dampeners. Pumps installed with vibration dampeners should always have expansion joints fitted at both the suction and the discharge side to prevent the pump from being supported by the flanges.
Expansion joints Expansion joints are installed to:
• Absorb expansions/contractions in the pipework caused by liquid temperature changes
• Reduce mechanical strain in connection with pressure waves in the pipework
• Isolate mechanical noise in the pipework (not for metal bellows expansion joints) Expansion joints should not be installed to compensate for inaccuracies in the pipework, such as center displacement or misalignment of flanges. Expansion joints are fitted at a minimum distance of 1 to 1.5 times the pipe diameter from the pump on the suction side as well as on the discharge side. This prevents the development of turbulence in the expansion joints, resulting in better suction conditions and a minimum
80
Expansion joint
Foundation Pump unit
Base plate Vibration dampeners Floor
Fig. 2.1.11: Installation with expansion joints, vibration dampeners and fixed pipework
Figures 2.1.12-2.1.14 show examples of rubber bellows expansion joints with or without tie bars. Fig. 2.1.12: Rubber bellows expansion joints with tie bars
Expansion joints with tie bars can be used to minimize the forces caused by the expansion joints and are recommended for sizes larger than four inches. An expansion joint without tie bars will exert force on
Fig. 2.1.13: Rubber bellows expansion joints without tie bars
the pump flanges, which in turn affects the pump and the pipework. The pipes must be fixed so that they do not stress the expansion joints and the pump, see figure 2.1.11. The fixed points should always be placed as close to the expansion joints as possible. Follow the expansion joint supplier’s instructions.
Fig. 2.1.14: Metal bellows expansion joints with tie bars
At temperatures above 212°F combined with a high pressure, metal bellows expansion joints are often preferred, due to the risk of rupture. Lp (dB)
Pain threshold
120 Threshold of hearing
2.1.5 Sound level The sound level (L) in a system is measured in decibel (dB). Noise is unwanted sound. The level of noise can be measured in the following three ways:
100 Music
80 60
Speech 40 20 0
1. Pressure – Lp : The pressure of the air waves 2. Power – L w : The power of the sound 3. Intensity - L l: The power per m2 (will not be covered in this book)
20
501 002 00 500Hz 1
2
5
102 0kHz Frequency kHz
Fig. 2.1.15: Threshold of hearing vs. frequency
It is not possible to compare the three values directly, but it is possible to calculate between them based on standards. A rule-of-thumb is:
Smaller pumps, e.g. 2 hp: Lw = LP + 11 dB Larger pumps, e.g. 150 hp: Lw = LP + 16 dB
81
Section 2.1 Pump installation
Sound levels are indicated as pressure when they are below 85 dB(A) and as power when exceeding 85 dB(A). Noise is subjective and depends on a person´s ability to hear. Therefore, the above mentioned measurements get weight according to the sensitivity of a standard ear, see figure 2.1.15. The weighting is known as A-weighting [dB(A)], expressed as: LpA, and the measurements are adjusted depending on frequency. In some cases the weighting increases and in other cases it decreases, see figure 2.1.16. Other weightings are known as B and C but they are used for other purposes not covered in this book.
dB (A)
10 0 -10 -20 -30 -40 -50 -60 -70 -80 10
100
1000
10000 Hz
Fig. 2.1.16 A-weighting curve
15
In the case of two or more pumps in operation, the sound level can be calculated. If the pumps have the same sound level, the total sound level can be calculated by adding the value, see figure 2.1.17. For example, two pumps is Lp + 3 dB, three pumps is Lp + 5 dB. If the pumps have different sound levels, values from figure 2.1.18 can be added. Indications of sound level should normally be stated as free field conditions overreflecting surface, meaning the sound level on a hard floor with no walls. Guaranteeing values in a specific room in a specific pipe system is difficult because these values are beyond the reach of the manufacturer. Certain conditions could have a negative impact (increased sound level) or a positive impact on the sound level. Recommendations to installation and foundation can be given to eliminate or reduce the negative impact of sound level.
10
5
4
8
12
16
20
24
Fig. 2.1.17 Increase of the total sound pressure level with equal sources
3
2.5
2
1.5
Experience values: 1
Rise of + 3 dB + 5 dB +10 dB
Perceived as: Slightly noticeable Clearly noticeable Twice as loud
0.5
2
4
6
8
10
Fig. 2.1.18 Increase of the total sound pressure level with different sources
82
Section 2.2 Pump performance
When examining a pump, several things should be evaluated. For example, if the pump is rusty or makes abnormal noise, a number of values must be identified in order to determine if the pump is performing properly. On the next pages, three values are presented for examining a pump’s performance: Hydraulic terms, electrical terms, mechanical terms and liquid properties.
2.2.1 Hydraulic terms Flow, pressure and head are the most important hydraulic terms pertinent to pump performance.
Flow Flow is the amount of liquid that passes through a pump within a certain period of time. Volume flow and mass flow are the two flow parameters considered for a performance reading.
Qm = ρ . Q
; Q =
Qm ρ
Volume flow Volume flow (Q) is read from a pump curve - or, put in another way, a pump can move a given volume per unit of time, measured in gallons per minute, no matter the density of the liquid. For water supply, for example, volume flow is the most important parameter because a certain volume of water is needed for drinking or irrigation. Throughout this book the term flow refers to volume flow.
Water
Examples
Unit
VolumeflowQ
GPM
Density
lb/ft3
62.30
58.86
lb/h
22000
20730
lb/s
6.1
5.7
Mass flow Qm
at 68°F
a2t 48 °F
44.02
Fig. 2.2.1: Calculation examples
Mass flow Mass flow (Qm) is the mass which a pump moves per unit of time and is measured in pounds per second. The liquid temperature has an influence on how big a mass flow can move per unit of time since the liquid density changes with the temperature. In heating, cooling and air-conditioning systems, the mass flow is essential to identify because the mass is the carrier of energy (see section on Heat Capacity).
83
Section 2.2 Pump performance
Pressure Pressure (p) is a measure of force per unit area. Total pressure is the sum of the static pressure and the dynamic pressure:
Later in this chapter, dynamic pressure in connection with determining the head of a pump will be discussed.
psta psta
Q
Static pressure Static pressure psta is the pressure measured with a pressure gauge placed perpendicular to the flow or in a non-moving liquid, see figure 2.2.2.
ptot
pdyn psta ptot
ptot
Fig. 2.2.2: How to determine the static pressure Psta, the dynamic pressure Pdyn and the total pressure Ptot
Dynamic pressure Dynamic pressure pdyn is caused by liquid velocity and is calculated by the following formula:
1 2 where: ρ is the density of the liquid in [lb/ft 3] v is the velocity of the liquid in [ft/s]
D1
v1
1 2 So, an increase in pipe diameter, as the one shown in figure 2.2.2 results in an increase in the static head which is measured with the pressure gauge p 2. In most pumping systems, the dynamic pressure pdyn has a minor impact on the total pressure. For example, if the velocity of a water flow is 14.7 ft/s, the dynamic pressure is around 1.45 psi, which is considered insignificant in many pumping systems.
84
v2
A P
Dynamic pressure can be converted into static pressure by reducing the liquid velocity and vice versa. Figure 2.2.3 shows a part of a system where the pipe diameter increases from D1 to D2 resulting in a decrease in liquid speed from v1 to v2. Assuming that there is no friction loss in the system, the sum of the static pressure and the dynamic pressure is constant throughout the horizontal pipe.
p2
p1
D2
B ptot psta
pdyn
Fig. 2.2.3: The static pressure increases if the liquid velocity is reduced. The figure applies for a system with insignificant friction loss
i
i
l i i i
i
Measuring pressure Pressure is measured in psi (Ib/in²), or bar (10 5 Pa). When dealing with pressure, it is important to know the point of reference for the pressure measurement. Two types of pressure are essential with pressure measurement: Absolute pressure and gauge pressure.
Conversion table for pressure units d e s ig n a t i o n
k P a ft of H2O m of H2O
psi
1 psi
1
6.895
1 kPa
0.145
1
1 feet of H2O
0.4335
2.969
1 m of H2O
1.422
9.806
1 m H2O
14.696
1 bar 14.504 * Physical atmosphere
Absolute pressure Absolute pressure (Pabs) is defined as the pressure above absolute vacuum, 0 atmospheres, that is the absolute zero for pressure. Usually, “absolute pressure” is used in cavitation calculations.
bar
atm
2.307
0.703
0.335
0.102
1
0.305
3.281
1
0.097
0.098
101.325 33.9
10.333
1
1.013
33.5
10.197
0.987
1
100
0.068
0.069
0.0097
0.01
0.0295
0.03
Fig. 2.2.4: Conversion table for pressure units
Gauge pressure Gauge pressure (Pg), often referred to as overpressure, is higher than normal atmospheric pressure (1 atm). Normally, pressure p is stated as gauge pressure because most sensor and pressure gauge measurements account for the pressure difference between the system and the atmosphere. Throughout this book the term pressure refers to gauge pressure.
Head The head (H) of a pump is an expression of how high the pump can raise a liquid. Head is measured in feet (ft) and is independent of the liquid density. The following formula shows the relationship between pressure (p) and head (H):
t f 1 . 6 2
t f 1 . 4 3
t f 5 . 2 4
t f 4 . 5 3
Brine at 68°F
Water at 68°F
Water at 203°F
Diesel oil at 68°F
SG = 1.3
SG = 0.997
SG = 0.96
SG = 0.80
14.7 psi = 26.1 ft
14.7 psi = 34.1ft
14.7 psi = 35.4 ft
14.7 psi = 42.5 ft
1 4p.7s i
1 4p.7s i
1 4p.7s i
1 4p.7si
l Fig. 2.2.5: Pumping four different liquidsi ati 14.7 psi at the discharge side of the pump results in four different heads i (ft), hence four different duty points i i
i
2.307 SG where : H is the head in [ft] p is the pressure in psi SG is the specific gravity of the liquid Pressure p is measured in [psi].
Other pressure units are used as well, see figure 2.2.4. The relationship between pressure and head is shown in figure 2.2.5, where a pump handles four different liquids.
85
Section 2.2 Pump performance
How to determine the head The pump head is determined by reading the pressure on the flanges of the pump 2p, p1 and then converting the values into head, see figure 2.2.6. However, if a static difference in head is present between the two measuring points, as it is in the case in figure 2.2.6, it is necessary to compensate for the difference. And if the port dimensions
The correction due to the difference in port diameter is caused by the difference in the dynamic pressure. Instead of calculating the correction from the formula, the contribution can be read in a nomogram, see appendix F. v2 D2
of the two measuring points differ from one another, the actual headhas to be corrected for this as well.
p2
The actual pump head H is calculated by the following formula:
v1 h2
2.31
D1 p1
h1
SG where : H is the actual pump head in [ft] p is the pressure atthe flanges in [ft] SG is the specific gravity ofthe liquid g is the acceleration of gravity in [ft/s2] h is the staticheight in [ft] v is the liquidvelocity in [ft/s] The liquid velocity v is calculated by the following formula:
0.4085
where: v is the velocity in [ft/s] Q is the volume flow in [GPM] D is the port diameter in [in] A is the area Combining these two formulas, head, H,depends on the following factors: The pressure measurements 1pand p2, the difference in static height between the measuring points h2-h1, the flow through the pump Q, and the diameter of the twoports D1 and D2 .
2.31
0.4085 Q SG
86
Fig. 2.2.6: Standard end-suction pump with dimension difference on suction and discharge ports
Calculation example
v2 = 17.8 ft/s2
A pump of the same type as the one shown in figure 2.2.7 is installed in a system with the following data: Q = 1057 GPM p1 = 7.25 psi p2 = 15.9 psi Liquid: Water at 680F
D2= 4.9 in
p2 = 15.9 psi h2 - h1 = 1 ft
D1 = 5.9 in 2
v1 = 12.3 ft/s
Suction port diameter D1 = 6 in Discharge port diameter D2 = 5 in The difference in height between the two ports where the pressure gauges are installed is h 2-h1 = 1 ft
p1 = 7.25 psi
Fig. 2.2.7: Standard end-suction pump with different dimensions of suction and discharge ports (Example)
We are now able to calculate the head of the pump:
2.31
0.4085 Q
SG
2.31 (15.9 - 7.25) 1.0
1
0.4085 1057
4.9
5.9
19.98 1 5.82 26.80 ft
As it appears from the calculation, the pressure difference measured by pressure gauges is about 1 ft lower than what the pump is actually performing. The deviation is caused by the difference in height between the pressure gauges (1 ft) and by the difference in port dimensions,which in this case is 1 inch.
87
Section 2.2 Pump performance
If the pressure gauges are placed at the same static height or if a differential pressure gauge is used for the measurement, it is not necessary to compensate for the difference in height (h2-h1). With in-line pumps, where inlet and outlet are placed at the same level, the two ports often have the same diameter. For these types of pumps a simplified formula is used to determine the head:
2.31 ( SG
p1
p2
h1
(
h2
Fig.2.2.7.a: Inline pump with same static height on inlet and outlet. h2 = h1
H = head in ft P = psi SG = specific gravity
Differential pressure The differentialpressure ( p) is the pressure difference between the pressures measured at two points, that is, the pressure drops across valves in a system. Differential pressure is measured in the same units as pressure.
Dry cooler
h
System pressure The system pressure is the static pressure, which refers to when the pumps are not running. System pressure is important to consider when dealing with a closed system. The system pressure, measured in feet in the lowest point, must always be higher than the height of the system to ensure that the system is filled with liquid and can be vented properly.
88
Chiller Hsyst Hsyst > h
Fig.2.2.8: The system pressure Hsta in a closed system has to be higher than the physical height of the installation
Cavitation Cavitation in a pump occurs when the suction pressure is lower than the vapor pressure of the liquid pumped, see figures 2.2.9 and 2.2.10. When the pressure on the suction side of the pump drops below the vapor pressure of the pumped liquid (figure 2.2.10 yellow dot), vapor bubbles form. As the pressure in the pump rises, the bubbles collapse releasing shock waves (figure 2.2.10 red dot) which can damage impellers. The rate of damage depends on the properties of the impeller material. Stainless steel is more resistent to cavitation than bronze, and bronze is more resistant than cast iron, see section 1.6.3. Additional damage to bearings, shaft seals and welds may occur due to increased noise and vibration caused by cavitation. This damage is often only detected when the pump is disassembled. Pump performance is harmed by cavitation due to decreases in both flow (Q) and head (H), see figure 2.2.11.
Net Positive Suction Head To calculate the risk of cavitation, the Net Positive Suction Head Required (NPSHr) for the pump is compared with the Net Positive Suction Head Available r, which is the amount of (NPSHa) of the system. NPSH suction head required to ensure the pump performs at full capacity, is determined by the manufacturer and typically included on the performance curve. NPSH a is a function of the system in which the pump will be applied and is calculated as follows:
hmax = Maximum suction head Hb = Atmospheric pressure at the pump site; this is the theoretical maximum suction lift, see figure 2.2.13 Hf = Friction loss in the suction pipe NPSHr = Net Positive Suction Head read at the NPSH curve at the highest operational flow, see figure 2.2.12.
a = Front of impeller vanes b = Back of impeller vanes
a b Imploding vapor bubbles Fig.: 2.2.9: Implosion of cavitation bubbles on the back of impeller vanes p
] a P [ e r u
a
a = Front of impeller vanes b = Back of impeller vanes
b
s s e r P
p1
p
Vapor pressure
Impeller inlet
Impeller outlet
Fig.: 2.2.10: Development of pressure through a centrifugal pump H
NPSHa = H b + H s — H f — Vp Hb = Barometric Pressure, in feet absolute Hs = Suction Head, infeet absolute (positive or negative) Hf = Friction loss in suction piping, in feet absolute Vp = Vapor pressure at the maximum operating temperature, in feet absolute
Curve when pump cavitates
Q
Fig.: 2.2.11: Pump curve when pump cavitates H
NPSHa must be greater than the NPSHr to avoid cavitation .
Calculation of the risk of cavitation To avoid cavitation, the following formula is used to calculate the maximum suction head:
NPSH
Q
hmax = H b — H f — NPSHr — H v — H s
Fig.: 2.2.12: NPSH - curve
89
Section 2.2 Pump performance
The NPSH value indicates to what extent the pump is unable to create absolute vacuum, that is to raise a full water column 33.89 ft above sea level, see figure 2.2.13.
Height above sea level (ft)
Barometric pressure p (psi)
Water column H (ft)
0
14.692
33.89
212
1640.4
13.567
31.92
210.2
3280.8
13.039
30.05
204.8
6561.6
11.531
26.57
199.4
NPSH can either be considered in terms of NPSH r (required) or NPSHa (available). NPSHrequired The required suction head for the pump NPSHavailable The available suctionhead in the system
b
Fig.: 2.2.13: Barometric pressure above sea level
The NPSH value of a pump is determined by Hydraulic Institute testing standards and is made as follows. The suction head is reduced while the flow is kept at a constant level. When the differential pressure has decreased by 3%, the pressure at the pump’s suction side is read and the NPSH value of the pump is defined. The testing is repeated at different flows, forming the basis of the NPSH curve. Hv – Vapor pressure of the liquid. For more information concerning vapor pressure of water, go to Appendix D.
b
h
Hv (ft) 413
360
3 28
320
259
300
270
148 131 115 98 82
250
66
230
49 39 33 26
212
NPSH
194
Hb
176 158 140 122 104
Fig.: 2.2.14: System with indication of the different values that are important in connection with suction calculations
86 68 50 32
2.2.2 Electrical terms To examine a pump’s performance, a range of values must be considered. In this section the most important electrical values are presented: Power consumption, voltage, current and power factor.
tm (°F ) 370
280
Hf
Hv
Hs – Safety factor. H s depends on the situation and normally varies between 1.5 ft and 3 ft. For typical curve for liquid containing gas see figure 2.2.15.
Boiling point of water (°f)
] t f [ H
NPSH Liquid with air
Vented liquid
Q [GPM]
Fig.: 2.2.15: Typical NPSH curve for liquid containing gas
90
20 16 13 10 6.6 4.9 3.3 2.6 2.0 1.3 0.9 0.7 0.3
Power consumption Pumps are made of several components, see figure 2.2.16. The power consumption (P) of the different components is designated as follows:
P1 P1
The power input from the mains, or the amount of power the consumer must purchase.
P2
The power input to the pump, or the power output from the motor, often referred to as shaft power or brake horsepower (Bhp).
PH
Hydraulic power; the power that the pump transfers to the liquid in the form of flow and head, also known as water hp (Whp).
P2
For the most common pump types, the term power consumption normally refers to P 2. Power is measured in horsepower (hp).
Efficiency
PH
Efficiency (η) normally only covers the efficiency of the pump part, ηP. A pump’s efficiency is determined by several factors, including the shape of the pump housing, the impeller and diffuser design and the surface roughness. For typical pump units consisting of both pump and electric motor, the total efficiency ηT also includes the efficiency of the motor: Fig. 2.2.16: Pump unit with indication of different power consumption levels
If a frequency converter is also included, the efficiency of the entire unit must include the efficiency of the frequency converter (ηfc ):
91
Section 2.2 Pump performance
Voltage Like pressure drives flow through a hydraulic system, voltage (v) drives a current (I) through an electrical circuit. Voltage is measured in volts (V) and can be direct current (DC), e.g. 1.5 V battery – or alternating current (AC), e.g. electricity supply for houses, etc. Normally, pumps are supplied with AC voltage supply.
L1 L2 L3 N
} }
480V Three-phase supply 230V Single-phase supply
Ground Fig. 2.2.17: Mains supply, e.g.3 x 480 V
The layout of an AC main supply differs from one country to another. The most common layout is four wires with three phases (L1, L2, L3) and a neutral (N). A ground connection is added to the system as well, see figure 2.2.17. For a 3x480 V/230 V main supply, the voltage between any two of the phases (L1, L2, L3) is 480 V. The voltage between one of the phases and neutral (N) is 230 V. The ratio between the phase-phase voltage and the phase-neutral voltage is determined by the formula at right.
The ratio between the phase-phase voltage and the phase-neutral voltage is:
Current Current (I) is the flow of electricity and is measured in ampere (A). The amount of current in an electrical circuit depends on the supplied voltage and the resistance/ impedance in the electrical circuit.
Power and power factor Power (P) consumption is of high importance when it comes to pumps. For pumps with standard AC motors, the power input is found by measuring the input voltage and input current and by reading the value cosj on the pump motor nameplate. The term cosj is the phase angle between voltage and current and is referred to as power factor (PF). The power consumption P1 can be calculated by the formulas shown at right for a single-phase or a three-phase motor.
92
AC single-phase motor, e.g. 1 x 230 V
AC three-phase motor, e.g. 3 x 480 V
2.2.3 Liquid properties
Viscosity
When making system calculations, the following liquid properties should be considered: Liquid temperature, specific gravity, heat capacity, and viscosity.
Kinematic viscosity is measured in centiStokes [cSt] (1 cSt = 10-6 m2/s). The unit [SSU] Saybolt Universal is
Liquid temperature
formula: [SSU] = 4.62 . [cSt]
The liquid temperature (t,T) is measured in °F (Fahrenheit), °C (Celcius), or K (Kelvin). Temperature units of °C and K are actually the same, but 0°C is the freezing point of water and 0°K is the absolute zero; that is -273.15°C, the lowest possible temperature. The calculation between Fahrenheit and Celcius is °F = .°C 1.8 + 32. Hence, the freezing point of water is 0°C and 32°F, and the boiling point is 100°C and 212°F.
also used in connection with kinematic viscosity. For kinematic viscosity above 60 cSt, the Saybolt Universal viscosity is calculated by the following
18.42
Btu/lbm °F 0% pure water
16.74
20%
15.07
34% 44%
13.39
52%
11.72
Specific Gravity The Specific Gravity (SG) is a dimensionless unit defined as the ratio of density of the material to the density of water at a specified temperature of 68°F. See appendix K.
10.04 8.37 -40
-4
32
68
104
140
176
212
248°F
Fig. 2.2.18: Heat capacity vs. temperature for ethylene glycol
Liquid heat capacity The heat capacity (Cp) shows how much additional energy a liquid can contain per mass when it is heated. Liquid heat capacity depends on temperature, see figure 2.2.18. Heat capacity is considered in systems for transporting energy, such as heating, air-conditioning and cooling. Mixed liquids, such as glycol and water for air-conditioning, have a lower heat capacity than pure water, so higher flow is required to transport the same amount of energy.
93
Chapter 3. System hydraulics
Section 3.1: System characteristics 3.1.1 Single resistances 3.1.2 Closed and open systems
Section 3.2: Pumps connected in parallel and series 3.2.1 Pumps in parallel 3.2.2 Pumps connected in series
Section 3.1 System characteristics
Previously, in section 1.1.2, the basic characteristics of pump performance curves were discussed. In this chapter the pump performance curve at different operating conditions as well as a typical system characteristic will be examined. Finally, the interaction between a pump and a system will be discussed. System characteristic describes the relation between flow (Q) and head (H). The system characteristic depends on the type of system in question, closed or open.
• Closed systems A closed system is a circulating system like heating or air-conditioning systems, where the pump has to overcome the friction losses in the pipes, fittings, valves, etc. in the system.
• Open systems An open system is a liquid transport system like a water supply system where the pump must address the static head as well as overcome the friction losses in the pipes and components. When the system characteristic is drawn in the same system of co-ordinates as the pump curve, the duty point of the pump can be determined as the point of intersection of the two curves, see figure 3.1.1. Open and closed systems consist of resistances (valves, pipes, heat exchanger, etc.) connected in series or parallel, which altogether affect the system characteristic. Following is a discussion on how these resistances affect the system characteristic.
96
Fig. 3.1.1: The point of intersection between the pump curve and the system characteristic is the duty point of the pump
3.1.1 Single resistances Every component in a system constitutes a resistance against the liquid flow which leads to a head loss. The following formula is used to calculate the head loss:
∆H = k . Q2 k is a constant, which depends on the component in question, and Q is the flow through the component. As it appears from the formula, the head loss is proportional to the flow to the second power. So, if it is possible to lower the flow in a system, a substantial reduction in the pressure loss occurs.
Resistances connected in series The total head loss in a system consisting of several components connected in series is the sum of head losses that each component represents. Figure 3.1.2 shows a system consisting of a valve and a heat exchanger. If we do not consider the head loss in the piping between the two components, the total head loss, Htot, is calculated by adding the two head losses:
Fig. 3.1.2: The head loss for two components connected in series is the sum of the two individual head losses
∆Htot = ∆H1 + ∆H2 Figure 3.1.2 shows how the resulting curve will look and what the duty point will be if the system is a closed system with only these two components. As it appears from the figure, the resulting characteristic is found by adding the individual head losses, H, at a given flow Q. The figure shows that the more resistance in the system, the steeper the resulting system curve will be.
97
Section 3.1 System characteristics
Resistances connected in parallel Contrary to connecting components in series, connecting components in parallel results in a more flat system characteristic. This is because the components installed in parallel reduce the total resistance in the system, and thereby the head loss. The differential pressure across the components connected in parallel is always the same. The resulting system characteristic is defined by adding all the components’ individual flow rates for a specific H. Figure 3.1.3 shows a system with a valve and a heat exchanger connected in parallel. The resulting flow can be calculated by the following formula for a head loss equivalent to H
Q tot = Q 1 + Q2
Fig. 3.1.3: Components connected in parallel reduce the resistance in the system and result in a more flat system characteristic
3.1.2 Closed and open systems As mentioned previously, pump systems are split into two types: Closed and open systems. This section will examine the basic characteristics of these systems.
Closed systems Typically, closed systems are systems which transport heat energy in heating systems, air-conditioning systems and process cooling systems. A common feature of these closed systems is that the liquid is circulated and is the carrier of heat energy. Heat energy is what the system must transport. Closed systems are characterized as systems with pumps that overcome the sum of friction losses which are generated by all the components. Figure 3.1.4 shows a schematic drawing of a closed system where a pump circulates water from a heater through a control valve to a heat exchanger.
98
Fig. 3.1.4: Schematic drawing of a closed system
All these components, along with the pipes and fittings, result in a system characteristic as shown in figure 3.1.5. The required pressure in a closed system (which the system curve illustrates) is a parabola starting at the point (Q,H) = (0,0) and is calculated by the following formula:
H = k . Q2 As the formula and curve indicate, the pressure loss is approaching zero when the flow drops.
Fig. 3.1.5: The system characteristic for a closed system is a parabola starting at point (0,0)
Open systems Open systems use the pump to transport liquid from one point to another, e.g. water supply irrigation and industrial process systems. In these systems, the pump deals with the static head of the liquid and must overcome the friction losses in the pipes and the system components. There are two types of open systems: • Open systems where the total required static head is positive. • Open systems where the total required static head is negative.
Fig. 3.1.6: Open system with positive static head
Open system with positive static head Figure 3.1.6 shows a typical open system with positive static head. A pump transports water from a break tank at ground level up to a roof tank on the top of a building. The pump must provide a head higher than the static head of the water (h), as well as the necessary head to overcome the total friction loss between the two tanks in piping, fittings, valves, etc. (Hf). The pressure loss depends on the rate of flow, see figure 3.1.7.
Q1
Q
Fig. 3.1.7: System characteristic together with the pump performance curve for the open system in figure 3.1.6
99
Section 3.1 System characteristics
Figure 3.1.7 shows that, in an open system, no water flows if the maximum head (Hmax) of the pump is lower than the static head (h). Only when H > h will water start to flow from the break tank to the roof tank. The system curve also shows that the lower the flow rate, the lower the friction loss (H f) and, consequently, the lower the power consumption of the pump. So, the flow (Q1) and the pump size have to match the need for the specific system. This is a general rule for liquid transport systems: A larger flow leads to a higher pressure loss, whereas a smaller flow leads to a smaller pressure loss and, consequently, a lower energy consumption.
Open system with negative static head A typical example of an open system with negative required head is a pressure booster system, as in a water supply system. The static head (h) from the water tank brings water to the consumer. The water flows, although the pump is not running. The difference in height between the liquid level in the tank and the altitude of the water outlet (h) results in a flow equivalent to Qo. However, the head is insufficient to ensure the required flow (Q1) to the consumer, so the pump has to boost the head to the level (H1) in order to compensate for the friction loss (Hf) in the system. The system is shown in figure 3.1.8, and the system characteristic and the pump performance curve are shown in figure 3.1.9.
Fig. 3.1.8: Schematic drawing of a open system
The resulting system characteristic is a parabolic curve starting at the H-axes in the point (0,-h). The flow in the system depends on the liquid level in the tank. If the water level in the tank is reduced, the height (h) is reduced. This results in a modified system characteristic and a reduced flow in the system, see figure 3.1.9.
100
Fig. 3.1.9: System characteristic andthe pump performance curve for the open system shown in figure 3.1.8
Section 3.2 Pumps connected in parallel and series
To increase total pump performance in a system, pumps are often connected in parallel or series. This section will focus on these two ways of connecting pumps.
3.2.1 Pumps in parallel Pumps connected in parallel are often used when: • The required flow is higher than one single pump
can supply • The system has variable flow requirements which
are met by switching parallel-connected pumps on and off Normally, pumps connected in parallel are of similar type and size. However, the pumps can be of different size, or one or several can bespeed-controlled, and thereby have different performance curves.
,
To avoid bypass circulation in pumps that are not running, a check valve is connected in series with each pump. The resulting performance curve for a system consisting of several pumps in parallel is determined by adding the flow, which the pumps deliver at a specific head. Figure 3.2.1 shows a system with two identical pumps connected in parallel. The system’s total performance curve is determined by adding Q 1 and Q2 for every value of head which is the same for both pumps, H1=H2 . Because the pumps are identical, the resulting pump curve has the same maximum head, H max, but the maximum flow, Qmax, is double. For each value of head, the flow is the double as for a single pump in operation:
Fig. 3.2.1: Two pumps connected in parallel with similar performance curves
Q = Q1 + Q 2 = 2 Q 1 = 2 Q 2
101
Section 3.2 Pumps connected in parallel and series
Figure 3.2.2 shows two different sized pumps connected in parallel. When adding Q1 and Q2 for a given head H1=H2, the resulting performance curve is defined. The hatched area in figure 3.2.2 shows that P1 is the only pump to supply in that specific area because it has a higher maximum head than P2.
Speed-controlled pumps connected in parallel For varying flow demand, speed-controlled pumps connected in parallel offer efficient pump performance. This method is common to water supply and pressure boosting systems. Later in chapter 4, speed-controlled pumps will be discussed in detail.
Fig 3.2.2: Two pumps connected in parallel with unequal performance curves
A pumping system with two speed-controlled pumps with the same performance curve covers a wide performance range, see figure 3.2.3. A single pump covers the required pump performance up to Q 1. Above Q1 both pumps must operate to meet the performance needed. If both pumps are running at the same speed, the resulting pump curves look like the orange curves shown in figure 3.2.3. Note that the duty point at Q 1 is reached with one pump running at full speed. The duty point can also be achieved when two pumps are running at reduced speed, see figure 3.2.4 (orange curves). The figure also compares efficiency. The duty point for one pump running at full speed results in low pump efficiency because the duty point is located far out on the pump curve. The total efficiency is much higher when two pumps run at reduced speeds, although the maximum efficiency of the pumps decreases slightly at reduced speeds.
Fig. 3.2.3: Two speed-controlled pumps connected in parallel (same size). The orange curve shows the performance at reduced speed
Even though one single pump is able to maintain the required flow and head, it is sometimes necessary due to efficiency and, thus, energy consumption to use both pumps at the same time. Whether to run one or two pumps depends on the actual system characteristic and pump type.
102
Fig. 3.2.4: One pump at full speed compared to two pumps at reduced speed. In this case the two pumps have the highest total efficiency
3.2.2. Pumps connected in series Normally, pumps connected in series are used in systems where high pressure is required. This is also the case for multistage pumps that are based on the series principle; that is, one stage equals one pump. Figure 3.2.5 shows the performance curve of two identical pumps connected in series. The resulting performance curve is made by marking the double head for each flow value in the system of co-ordinates. This results in a curve with the double maximum head (2⋅Hmax) and the same maximum flow (Qmax) as each of the single pumps.
Fig. 3.2.5: Two equal sized pumps connected in series
Figure 3.2.6 shows two different sized pumps connected in series. The resulting performance curve is determined by adding H1 and H2 at a given common flow Q1=Q2. The hatched area in figure 3.2.6 shows that P2 is the only pump to supply in that area because it has a higher maximum flow than P 1.
Q
Fig. 3.2.6: Two different sized pumps connected in series
As discussed in section 3.2.1, unequal pumps can be a combination of different sized pumps or of one or several speed-controlled pumps. The combination of a fixed-speed pump and a speed-controlled pump connected in series is often used in systems where a high and constant pressure is required. The fixed-speed pump supplies the liquid to the speedcontrolled pump whose output is controlled by a pressure transmitter, (PT), see figure 3.2.7. Fig. 3.2.7: Equal sized fixed-speed pump and speed-controlled pump connected in series. A pressure transmitter PT together with a speed controller is making sure that the pressure is constant at the outlet of P2.
103
Chapter 4. Performance adjustment of pumps
Section 4.1: Adjusting pump performance 4.1.1 Throttle control 4.1.2 Bypass control 4.1.3 Modifying impeller diameter 4.1.4 Speed control 4.1.5 Comparison of adjustment methods 4.1.6 Overall efficiency of the pump system 4.1.7 Example: Relative power consumption when the flow is reduced by 20%
Section 4.2: Speed-controlled pump solutions 4.2.1 Constant pressure control 4.2.2 Constant temperature control 4.2.3 Constant differential pressure in a circulating system 4.2.4 Flow-compensated differential pressure control
Section 4.3: Advantages of speed control
Section 4.4: Advantages of pumps with integrated frequency converter 4.4.1 Performance curves of speed-controlled pumps 4.4.2 Speed-controlled pumps in different systems
Section 4.5: Frequency converter 4.5.1 Basic function and characteristics 4.5.2 Components of the frequency converter 4.5.3 Special conditions regarding frequency converters
Section 4.1 Adjusting pump performance
When selecting a pump for a given application, it is important to choose one where the duty point is in the high-efficiency area of the pump. Otherwise, the power consumption of the pump is unnecessarily high, see figure 4.1.1. However, sometimes it is not possible to select a
η
H
[ft]
[%]
60 50 40 30
70 60
20
50 40
10
20
30
pump that fits the optimum duty point because the requirements of the system change or the system curve changes over time. Therefore, it may be necessary to adjust the pump performance so that it meets the changed requirements. The most common methods of changing pump performance are: • Throttle control • Bypass control • Modifying impeller diameter • Speed control
Choosing a method of pump performance adjustment is based on an evaluation of the initial investment along with the operating costs of the pump. All methods can be carried out continuously during operation apart from the modifying impeller diameter–method. Often, oversized pumps are selected for the system. It is then necessary to limit the performance – primarily the flow rate, and in some applications, the maximum head. The four adjusting methods are discussed on the following pages.
106
10 0 0
5
10
15
20
25
30
35
40 Q [GPM]0
Fig.: 4.1.1: When selecting a pump it is important to choose one where the duty point is within the high efficiency area.
4.1.1 Throttle control A throttle valve may be placed in series with the pump, permitting the duty point to be adjusted. Throttling results in a flow reduction, see figure 4.1.2. The throttle valve adds resistance to the system, raising the system curve. Without the throttle valve, the flow is Q2. With the throttle valve connected in series with the pump, the flow is reduced to Q 1.
Hp Throttle valve
System Hv
Hs
H Pump
Throttle valves can be used to limit the maximum flow. In the example, the flow will never be higher than Q 3 even if the srcinal system curve is completely flat, meaning there is no resistance in the system. When the pump performance is adjusted with this method, the pump will deliver a higher head than necessary for that particular system.
Resulting characteristic Smaller pump System
Hv
Throttle valve
Hs Q Q1 Q2 Q3 Fig.: 4.1.2: The throttle valve increases the resistance in the system, consequently reducing the flow.
If the pump and the throttle valve are replaced by a smaller pump, the pump will be able to meet the desired flow Q1 at a lower pump head, resulting in less power consumption, see figure 4.1.2. Bypass valve
4.1.2 Bypass control Compared to the throttle valve, installing a bypass valve will result in a certain minimum flow, Q , in the BP pump independent of the system characteristics, see figure 4.1.3. The flow, Q P, is the sum of the flow in the system, QS, and the flow in the bypass valve, QBP..
QBP QP
QS
System
HP
H Bypass valve
Hmax Smaller pump
The bypass valve will introduce a maximum limit of head to the system, Hmax , see figure 4.1.3. Even when the required flow in the system is zero, the pump will never run against a closed valve. Like the throttling valve method, the required flow, QS, can be met by a smaller pump and no bypass valve, resulting in a lower flow and less energy consumption.
System
Qs
QBP Resulting characteristic
HP Pump
QBP
QS
QP
Q
Fig.: 4.1.3: The bypass valve diverts part of the flow from the pump, reducing the flow in the system
107
Section 4.1 Adjusting pump performance
4.1.3 Modifying impeller diameter Another way to adjust the performance of a centrifugal pump is to modify the impeller diameter, reducing the diameter which, consequently, reduces pump performance. Compared to the throttling and bypass methods, which can be carried out during operation, the impeller trimming has to be done in advance before the pump is installed or in connection with service; it cannot be done while the pump is in operation. The following formula shows the relationship between the impeller diameter and the pump performance:
Note that the formulas are an expression of an ideal pump. In practice, the pump efficiency decreases when the impeller diameter is reduced. For minor changes of the impeller diameter, Dx > 0.8 . Dn, the efficiency is only reduced by a few percentage points. The degree of efficiency reduction depends on pump type and duty point.
D
H
Hn Hx
Dn Dx
As it appears from the formulas, the flow and the head change with the same ratio: that is, the ratio change of the impeller diameter to the second power. The duty points following the formulas are placed on a straight line starting in (0,0). The change in power consumption is following the diameter change to the fourth power.
4.1.4 Speed control The last method of controlling the pump performance to be covered in this section is the variable speed control method. Speed control by means of a frequency converter is the most efficient way of adjusting pump performance exposed to variable flow requirements.
108
Qx
Qn
Q
Fig. 4.1.4: Change in pump performance when the impeller diameter is reduced
The following equation applies with close approximation to how the change in speed of a centrifugal pump influences the performance of the pump:
The affinity laws apply when the system characteristic remains unchanged for nn and nx and forms a parabola through (0,0) – see section 3.1.2 (p 99). The power equation implies that the pump efficiency is unchanged at the two speeds. The formulas in figure 4.1.5 show that the pump flow (Q) is proportional to the pump speed (n). The head (H) is proportional to the second power of the speed (n) whereas the power (P) is proportional to the third power of the speed. In practice, a reduction of the speed will result in a slight fall in efficiency. The efficiency at reduced speed (nx) can be estimated by the following formula, which is valid for speed reduction down to 50% of the maximum speed:
If the need for precise power saved is desired, frequency converter and motor efficiencies must be taken into account.
Fig. 4.1.5: Pump parameters for different affinity equations
109
Section 4.1 Adjusting pump performance
4.1.5 Comparison of adjustment methods When the pump and its performance-changing device is considered as one unit, the resulting QHcharacteristic of this device can be compared to different systems.
Throttle control The throttling method implies a valve connected in series with a pump, see figure 4.1.6a. This connection acts as a new pump at unchanged maximum head but reduced flow performance. For an illustration of the pump curve, Hn, the valve curve, and the curve for the complete system, - Hx, see figure 4.1.6b.
Bypass control When connecting a valve across the pump, the connection acts as a new pump at reduced maximum head and a QH curve with a changed characteristic, see figure 4.1.7a. The curve will be more linear than quadratic, see figure 4.1.7b.
a
b Throttle valve
Hn
a
Valve
b Bypass valve
Modifying impeller diameter This method does not imply extra components. Figure 4.1.8 shows the reduced QH curve (H x) and the srcinal curve characteristics (H n).
Hx
Fig. 4.1.6: Throttle valve connected in series with a pump
Hn
Hx
Valve
Fig. 4.1.7: Bypass valve connected across the pump
Speed control The speed control method results in a new QH curve at reduced head and flow, see figure 1.4.9. The characteristics of the curves remain the same. However, when speed is reduced the curves become more flat as the head is reduced to a higher degree than the flow. In comparison, the speed control method also makes it possible to extend the performance range of the pump above the nominal QH curve by increasing the speed above nominal speed level of the pump; see the Hy curve in figure 4.1.9. If this over-synchronous operation is used, the size of the motor has to be taken into account.
110
D
Hn
Hx
Hn
Hx
Fig. 4.1.8: Impeller diameter adjustment
Speed controller
Fig. 4.1.9: Speed controller connected to a pump
Hy
4.1.6 Overall efficiency of the pump system Both the throttling and the bypass method introduce some hydraulic power losses in the valves (Ploss = k Q H), therefore reducing efficiency of the pumping system. Reducing the impeller size in the range of Dx/Dn>0.8 does not have a significant impact on pump efficiency and does not have a negative influence on the overall efficiency of the system. The efficiency of speed-controlled pumps is only affected to a limited extent if the speed reduction does not drop below 50% of the nominal speed. The efficiency is only reduced by a few percentage-points, and it does not have an impact on the overall running economy of speed-controlled solutions, see figure 1.4.16 in section 1.4.5.
4.1.7 Example: Relative power consumption when the flow is reduced by 20 % In a given installation the flow has to be reduced from Q = 260 GPM to 220 GPM. In the srcinal starting point (Q = 260 GPM and H = 230 ft) the power input to the pump is set relatively to 100%. Depending on the method of performance adjustment, the power consumption reduction will vary. This is further discussed on the following pages.
111
Section 4.1 Adjusting pump performance
Throttle control The power consumption is reduced to about 94% when the flow drops from 264 to 220 GPM. The throttling results in an increased head, see figure 4.1.10. The maximum power consumption for some pumps is at a lower flow than the maximum flow. If this is the case, the power consumption increases
= Modified duty point = Original duty point
H [ft]
249 229 180
Q
P2 100% 94%
because of the throttle. 220
Bypass control To reduce the flow in the system, the valve has to reduce the head of the pump to 180 ft. This can only be done by increasing the flow in the pump. As it appears from figure 4.1.11, the flow is consequently increased to 356 GPM, which results in an increased power consumption of up to 10% above srcinal consumption. The degree of increase depends on the pump type and the duty point. Therefore, in some cases, the increase in P2 is equal to zero and in rare cases, P2 might decrease slightly.
Modifying impeller diameter
Q [GPM]
264
Fig. 4.1.10: Relative power consumption - throttle control H [ft]
= Modified duty point = Original duty point 229 180
Q
P2 110% 100%
220
264
356
Q [GPM]
Fig. 4.1.11: Relative power consumption - bypass control H [ft]
When the impeller diameter is reduced, both the flow
= Modified duty point = Original duty point
and the head of the pump drop. By a flow reduction of 20%, the power consumption is reduced to around 67% of its srcinal consumption, see figure 4.1.12.
Speed control When the speed of the pump is controlled, both the flow and the head are reduced, see figure 4.1.13. Consequently, the power consumption has reduced to around 65% of the srcinal consumption.
229 180
Q
P2 100% 67%
220
264
Q [GPM]
Fig. 4.1.12: Relative power consumption - modifying impeller diameter H [ft]
To obtain the best possible efficiency, the impeller diameter adjustment method or the speed-controlled method of the pump are the best options for reducing the flow in the installation.When the pump has to operate in a fixed, modified duty point, the impeller diameter adjustment method is the best solution. However, in installations where the flow demand varies, the speed-controlled pump is the best solution.
= Modified duty point = Original duty point 70 55
Q Q
P2 100% 65%
50
60
Q [GPM]
Fig. 4.1.13: Relative power consumption - speed control
112
Summary Figure 4.1.14 gives an overview of the different adjustment methods that are presented in the previous section. Each method has its pros and cons which should be considered when choosing an adjustment method for a system.
Continuous The resulting performance adjustment curve will have possible?
Method
Throttle control
Yes
Reduced Q
Overall efficiency Relative power consumption by 20% of the pump reduction in flow system Considerably reduced
94%
Considerably reduced
110%
Slightly reduced
67%
Slightly reduced
65%
Throttle valve Hn Hx Valve
Bypass control
Yes
Reduced H and changed curve
Bypass valve
Hn Hx Valve
Modifying impeller diameter
No
Reduced Q and H
D Hn Hx
Speed control
Yes
Reduced Q and H
Speed controller
Hn Hx Hy
Fig. 4.1.14: Characteristics of adjustment methods.
113
Section 4.2 Speed-controlled pump solutions
As discussed in the previous section, speed control of pumps is an efficient way of adjusting pump performance to the system. In this section the possibilities of combining speed-controlled pumps with PI-controllers and sensors measuring system parameters, such as pressure, differential pressure and temperature, are discussed. The different options
Setpoint pset
PIcontroller
Actual value p1 Pressure transmitter
Break tank
Speed controller
PT h
will be presented by examples.
Q1
Taps
p1 H1
H
4.2.1 Constant pressure control A pump has to supply tap water from a break tank to different taps in a building. The demand for tap water is varying, so the system characteristic varies according to the required flow. Due to comfort and energy savings, a constant supply pressure is recommended. As it appears from figure 4.2.1, the solution is a speed-controlled pump with a PI-controller. The PI-controller compares the needed pressure, pset, with the actual supply pressure, p1, measured by a pressure transmitter, PT. If the actual pressure is higher than the setpoint, the PI-controller reduces the speed and, consequently, the performance of the pump until p1 is equal to pset. Figure 4.2.1 shows what happens when the flow is reduced from Qmax to Q 1 . The controller reduces the speed of the pump from nn to nx to ensure that the required discharge pressure is p1 = pset. The pump installation ensures that the supply pressure is constant in the flow range of 0 to Qmax. The supply pressure is independent of the level, (h), in the break tank. If h changes, the PIcontroller adjusts the speed of the pump so that p 1 always corresponds to the setpoint.
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nn nx pset
h
Q1
Qmax
Q
Fig. 4.2.1: Water supply system with speed-controlled pump delivering constant pressure to the system
4.2.2 Constant temperature control Performance adjustment through speed control is suitable for a number of industrial applications. Figure 4.2.2 shows a system with a water-cooled injection molding machine for high quality production. The machine is cooled with water at 59oF from a cooling plant. To ensure that the molding machine runs properly and is cooled sufficiently, the return pipe temperature has to be kept at a constant level; tr = 68oF. The solution is a speed-controlled pump controlled by a PI-controller. The PI-controller compares the needed temperature, tset, with the actual return pipe temperature, tr, which is measured by a temperature transmitter, TT. This system has a fixed system characteristic, and, therefore, the duty point of the pump is located on the curve between Qmin and Qmax. The higher the heat loss in the machine, the higher the flow of cooling water is needed to ensure that the return pipe temperature is kept at a constant level of 68 oF.
Fig. 4.2.2: System with injection molding machine and temperature- controlled circulator pump ensuring a constant return pipe temperature
4.2.3 Constant differential pressure in a circulating system Circulating systems, typically closed systems, are well suited for speed-controlled pumps, see Chapter 3. A differential pressure controlled circulator pump is recommended for circulating systems with variable system characteristic, see figure 4.2.3. This figure shows a heating system with a heat exchanger where the circulated water is heated up and delivered to three consumers, such as radiators, by a speed-controlled pump. A control valve is connected in series at each consumer to control the flow according to the heat requirement. The pump is controlled according to a constant differential pressure measured across the pump. As depicted by the horizontal line in figure 4.2.3, the pump system offers constant differential pressure in the Q-range of 0 to Qmax.
Fig. 4.2.3: Heating system with speed-controlled circulator pump delivering constant differential pressure to the system
115
Section 4.2 Speed-controlled pump solutions
4.2.4 Flow-compensated differential pressure control
Setpoint Hset
The main function of the pumping system in figure 4.2.4 is to maintain a constant differential pressure across the control valves at the consumers, such as radiators. In order to do so, the pump must overcome friction losses in pipes, heat exchangers, fittings, etc.
PIcontroller
Actual value H1
Speed controller Q1
DPT1 DPT2
As discussed in Chapter 3, the pressure loss in a system is proportional to the flow in second power. The best way to control a circulator pump in a system like the one shown in the figure at right is to allow the pump to deliver a pressure that increases when the flow increases.
H
nn nx
When the demand of flow is low, the pressure losses in the pipes, heat exchangers, fittings, etc. are low as well, and the pump supplies only a pressure equivalent to what the control valve requires, H setHf. When flow demand increases, pressure losses increase to the second power, and the pump has to increase the delivered pressure, depicted as the blue curve in figure 4.2.4.
Hset Hf
H1
Q1
Qmax
Q
Fig. 4.2.4: Heating system with speed-controlled peed-controlled circulator pump pump delivering flow-compensated differential pressure to the system
Such a pumping system can be designed as follows:
• The differential pressure transmitter is placed across the pump and the system is running with flow-compensated differential pressure control – DPT1, see figure 4.2.4.
pump curve data has to be stored in the controller. This data is used to calculate the flow as well as how much the setpoint Hset must be reduced at a given flow to ensure that the pump performance meets the required blue curve in figure 4.2.4.
• The differential pressure transmitter is placed close to the consumers and the system is running with differential pressure control – DPT 2, see fig. 4.2.4.
116
The first solution places the pump, PI-controller,
The second solution requires more installation costs because the transmitter has to be installed at the installation site, and the necessary cabling has to be added. Both systems are equal in performance. The transmitter measures the differential pressure at the
speed control andeasy the installation transmitter and close to one another providing making it possible to get the entire system as one single unit, see section 4.4. To get the system up and running,
consumer and compensates automatically for the increase in required pressure in order to overcome the increase in pressure losses in the supply pipes, etc.
Section 4.3 Advantages of speed control
A large number of pump applications do not require full pump performance 24 hours a day. Therefore, it is an advantage to be able to adjust the pump’s performance in the system automatically. As seen in section 4.1, the best possible way of adapting the performance of a centrifugal pump is by means of speed control of the pump. Speed control of pumps is normally made by a frequency converter unit. On the following pages, speed-controlled pumps in closed and open systems will be examined. The advantages that speed control provides and the benefits that speed-controlled pumps with frequency converters offer are presented first.
Reduced energy consumption Speed-controlled pumps use only the amount of energy needed to address a specific pump installation. Compared to other control methods, frequencycontrolled speed control offers the highest efficiency and the most efficient utilization of the energy, see section 4.1.
Low life cycle costs As we will see in Chapter 5, the energy consumption of a pump is a very important factor when calculating a pump’s life cycle costs. Therefore, it is important to keep the operating costs of a pumping system at the lowest possible level. Efficient operation leads to lower energy consumption and results in lower operating costs. Compared to fixed-speed pumps, it is possible to reduce the energy consumption by up to 50% with a speed-controlled pump.
Environment protection Energy-efficient pumps cause less pollution and harm to the environment.
Increased comfort Speed control in different pumping systems provides increased comfort in water supply systems, automatic pressure control, and where the soft-start of pumps reduce water hammer and noise generated by too high pressure in the system. In circulating systems, speed-controlled pumps ensure that the differential pressure is kept at a level so that noise in the system is minimized. Reduced system costs Speed-controlled pumps can reduce the need for commissioning and control valves in the system, thus reducing the total system costs.
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Section 4.4 Advantages of pumps with integrated frequency converter
In many applications, pumps with integrated frequency converters are the optimum solution. These pumps combine the benefits of a speed-controlled pump solution with the benefits gained from combining a pump, a frequency converter, a PI-controller and sometimes a sensor/pressure transmitter in one single unit, see figure 4.4.1. A pump with an integrated frequency converter is not just a pump, it is a system that can solve application problems or save energy in a variety of pump installations. Pumps with integrated frequency converters are ideal because they can be used instead of xed-speed pumps in replacement installations at no extra installation cost. The only requirement is a power supply connection and a fitting of the pump with an integrated frequency converter in the pipe system, and then the pump is ready for operation. After adjusting the required setpoint pressure, the system is operational. What follows is a brief description of the advantages that pumps with integrated frequency converter have to offer.
Setpoint
PIcontroller Frequency converter
M
PT
Easy to install Pumps with integrated frequency converters are just as easy to install as fixed-speed pumps. The motor is connected to the electrical power supply, and the pump is in operation. The manufacturer has made all internal connections and adjustments.
Optimal energy savings Because the pump, the motor and the frequency converter are designed for compatibility, operation of the pump system reduces power consumption.
One supplier One supplier can provide the pump, frequency converter and sensor which naturally facilitate the sizing, selection, and ordering procedures, as well as maintenance and service procedures.
118
Fig. 4.4.1: Pump unit with integrated frequency converter and pressure transmitter
Wide performance range Pumps with integrated frequency converters have a broad performance range which enables efficient performance under widely varied conditions and meets a wide range of requirements. Fewer pumps can replace many fixed speed pump types with narrow performance capabilities.
performance curve and the system characteristic of a closed and an open system.
H [ft] 100% 320 280 90% 240
4.4.1. Performance curves of speedcontrolled pumps
80%
200 70%
160
60%
120
50%
80
The following is a discussion of how a speed-controlled pump’s performance curve is read.
40
25%
0 0
20
40
60
80
100
120
140 Q [GPM]
P [hp] 1
Figure 4.4.2 provides an example of the performance curves of a speed-controlled pump. The first curve shows the flow-head (QH) curve, and the second curve shows the corresponding power consumption curve.
10 8 6 4 2 0 Q [GPM]
Fig 4.4.2: Performance curve for a speed-controlled pump
The performance curves are plotted for every 10% decrease in speed from 100% down to 50%. Likewise, the minimum curve represented by 25% of the maximum speed is also shown. As indicated in the diagram, you can select a specific duty point, QH, and find out at which speed the duty point can be reached and what the power consumption, P 1, is.
4.4.2 Speed-controlled pumps in different systems Speed-controlled pumps are used in a wide range of systems. The change in pump performance and, consequently, the potential energy savings depend on the system in question.
H
H
Pump curve
System characteristic
Closedsystem
Q
Pump curve
HO
System characteristic
Opensystem
Q
Fig 4.4.3: System characteristic point of a closed and an open system
As discussed in Chapter 3, the characteristic of a system is an indication of the required head a pump has to deliver to transport a certain quantity of liquid through the system. Figure 4.4.3 shows the
119
Section 4.4 Advantages of pumps with integrated frequency converter
Speed-controlled pumps in closed systems In closed systems, like heating and air-conditioning, the pump has to overcome the friction losses in the pipes, valves, heat exchangers, etc.In this section, an example of a speed-controlled pump in a closed system will be presented. The total friction loss by a full flow of 66 GPM is 39.3 ft, see figure 4.4.4.
The system characteristic starts in the point (0,0), shown by the red line in figure 4.4.5. Control valves in the system always need a certain operating pressure, so the pump cannot work according to the system characteristic. That is why some speedcontrolled pumps offer the proportional pressure control function, which ensures that the pump will operate according to the orange line shown in the figure. As you can tell from figure 4.4.5, the minimum performance is around 57% of the full speed. In a circulating system, operating at the minimum curve (25% of the full speed) can be relevant in some situations, such as night-time duty in heating systems.
Q = 66 GPM
Boiler or like
H
Consumers Fig. 4.4.4: Closed system
H
[ft] 100% 320 99% 280 90% 240
80%
200 70%
160
60%
120
50%
80 40
25%
0 0
20
40
60
80
100
120
140 Q [GPM]
P [hp] 1
10 8 6 4 2 0 Q [GPM]
Fig. 4.4.5: A speed-controlled pump in a closed system
120
Speed-controlled pumps in open systems The system characteristic as well as the operating range of the pump depend on the type of system in question. Figure 4.4.6 shows a pump in a pressure boosting / water supply system. The pump has to supply Q = 29 GPM to the tap which is placed h = 65 ft above the pump. The inlet pressure to the pump, p s, is 14.5 psi, the pressure at the tap, pt, has to be 29 psi and the total friction loss in the system by full flow, pf, is 18.8 psi.
he = 65.6 ft SG
pf = 18.8 psi ps = 14.5 psi Q = 29 GPM
Figure 4.4.7 shows the QH curve of a pump which meets the requirements described. The required head can be calculated by using the equation at right.
H pt - Pressure at tapping point ps - Suction pressure pf - Friction loss Q - Flow rate h - Static lift
For maximum head at a flow, Q, of 29 GPM, the equation to use follows:
H = He +
2.31 (pt)
SG
—
2.31 (ps)
SG
+
pt = 29 psi
Fig. 4.4.6: Pump in a water supply system
2.31 (pf)
SG
H = 65.6 + 2.31 (pt — ps + pf)
p -p (2-1) . 105 H + ρt. s = 20 + = 99.08 ft g 998 . 9.81
SG H = 65.6 + 2.31 (29 — 14.5 + 18.8) 1.0 H = 65.6 ft + 76.9 ft
H
[ft] 100% 200 99% 175 90% 150
H = 142.5 ft
80%
125 70%
HO
60%
75
To address this application from zero to maximum flow Q = 29 GPM, the pump operates in a relatively narrow speed band, from about 65%-99% of the full speed. In systems with less friction loss, the variation in speed will be even smaller. If there is no friction loss, the minimum speed in the above case is about 79% speed. As seen in the previous two examples, the possible
50%
50 25
25%
0 0
5
10
15
20
25
30
35 Q [GPM]
P [hp] 1
2.5 2.0 1.5 1.0 0.5 0 Q [GPM]
Fig. 4.4.7: A speed-controlled pump in an open system
variation in speed and power consumption is highest in closed systems. Therefore, the closed system accounts for the highest energy saving potential.
121
Section 4.5 Frequency converter
As mentioned, speed control of pumps involves a frequency converter. This section will provide a closer look at frequency converters, how they operate, and related precautions associated with using them.
4.5.1 Basic function and characteristics The speed of an asynchronous motor depends primarily on the pole number (2-pole, 4-pole, etc.) of the motor and the frequency of the voltage supplied. The amplitude of the voltage supplied and the load on the motor shaft also influence the motor speed, however, not to the same degree. Changing the frequency of the supply voltage is ideal for achieving asynchronous motor speed control. To ensure correct motor magnetization, it is also necessary to change the amplitude of the voltage.
T
f1 > f2
f1 f2 n
Fig. 4.5.1: Displacement of motor torque characteristic
Use of frequency/voltage control results in a change in torque which, in turn, changes speed. Figure 4.5.1 shows the motor torque characteristic (T) as a function of the speed (n) at two different frequencies/voltages. The load characteristic of the pump is also shown. As it appears from the figure, the speed is changed by changing frequency/voltage of the motor. The frequency converter changes frequency and voltage, so it can be concluded that the task of a frequency converter is to change the fixed supply voltage/frequency; for example, 3x480v/60 Hz into a variable voltage/frequency.
122
4.5.2. Components of the frequency converter In principle, all frequency converters consist of the same functional blocks. The basic function, as mentioned, is to convert the main electric supply into a new AC voltage with another frequency and amplitude. The frequency converter rectifies the incoming main electric supply and stores the energy in an intermediate circuit consisting of a capacitor. The DC voltage is then converted into a new AC voltage with another frequency and amplitude. The rectifier can handle 50 Hz or 60 Hz frequencies. Additionally, the incoming frequency will not influence the output frequency, as this is defined by the voltage/frequency pattern which is defined in the inverter. Using a frequency converter in connection with asynchronous motors provides the following benefits: • The system can be used in both 50 and 60 cycleareas without modifications • The output frequency of the frequency converter is independent of the incoming frequency • The frequency converter can supply output frequencies higher than mains supply frequency – making over synchronous operation possible As seen in figure 4.5.2, the frequency converter consists of three other components: An EMC filter, a control circuit and an inverter.
Mains supply AC EMC filter
Rectifier
Intermediate circuit DC
Inverter
Control circuit
Fig. 4.5.2: Functional blocks of the frequency converter
The EMC filter This block is not part of the primary function of the frequency converter and, in principle, could be left out. However, in order to meet EMC requirements and local requirements, the filter is necessary. The EMC filter prevents high noise signals from going back to the main electric supply and disturbing other electronic equipment connected to it. It also ensures that noise signals in the main electric supply generated by other equipment do not enter the electronic devices of the frequency converter, and cause damage or disturbances.
The control circuit The control circuit block has two functions. It controls the frequency converter and provides communication between the product and the surroundings.
The inverter The output voltage from a frequency converter is not sinusoidal like the normal mains voltage. The voltage
Vmotor Mean value of voltage
supplied to the motor consists of a number of squarewave pulses, see figure 4.5.3. The mean value of these pulses forms a sinusoidal voltage of the desired frequency and amplitude. The switching frequency can range from a few kHz up to 20 kHz, depending on the brand. To avoid noise in the motor windings, a frequency converter with a switching frequency above the range of audibility (~16 kHz) is preferable. This principle of inverter operation is called Pulse Width Modulation control (PWM), and it is the control principle most often used in frequency converters today. The motor current itself is almost sinusoidal. This is shown in figure 4.5.4a, indicating motor current (top) and motor voltage. In figure 4.5.4b, a section of the motor voltage is shown, indicating how the pulse/pause ratio of the voltage changes.
t
T = 1/fm Fig 4.5.3: AC voltage with variable frequency (fm) and variable voltage (Vmotor)
0 a
0
b *
* Detail
Fig 4.5.4: a) Motor current (top) and motor voltage at Pulse Width Modulation control. b) Section of motor voltage
123
Section 4.5 Frequency converter
4.5.3 Special conditions regarding frequency converters There are some conditions which the installer and user should be aware of when installing and using frequency converters or pumps with integrated frequency converters. A frequency converter will behave differently than a standard asynchronous motor at the main electirc supply side.
a
b
Fig 4.5.5 a): Three-phase, twopole standard asynchronous motor
Fig 4.5.5 b): Three-phase, twopole standard asynchronous motor with frequency converter
Non-sinusoidal power input, three-phase supplied frequency converters This type of frequency converter will not receive sinusoidal current from the electrical supply. This influences the dimensioning of the main elecrical cable, electric switch, etc. Figure 4.5.5 shows how the current and voltage appear for a: a) Three-phase, two-pole standard asynchronous motor b) Three-phase, two-pole standard asynchronous motor with frequency converter. In both cases the motor supplies 4.08 hp to the shaft. A comparison of the current shows the following differences, see figure 4.5.6:
Stand ard mot or
Mo to r w it h f re quency converter 460 V
Mains voltage
460 V
Mains current RMS
6.4 A
Mains current, peak
9.1 A
13.8 A
3.68 KW
3.69 KW
Power input, P1 cos ϕ, power factor (PF)
6.36 A
cos =ϕ0.83
PF = 0.86
Fig. 4.5.6: Comparison of current of a standard motor and a frequency converter
• The current for the system with frequency converter is not sinusoidal
• The peak current is much higher (approx. 52%) for the frequency converter option This is due to the design of the frequency converter connecting the electric supply to a rectifier followed by a capacitor. The charging of the capacitor occurs during short time periods where the rectified voltage is higher than the voltage in the capacitor at that moment. As mentioned, the non-sinusoidal current causes other conditions at the electric supply side of the motor. For a standard motor without a frequency converter, the relation between voltage (V), current (I) and power (P) is shown in the formula at right. The same formula cannot be applied for calculating the power input for motors with frequency converters.
124
V
PF
V
V
PF
(
(
PF
4.93 hp
Because these are not sinusoidal, there is no accurate way of calculating the power input based on simple current and voltage measurements. Instead, the power must be calculated by means of instruments and on the basis of instantaneous measurements of current and voltage. If the power (P) and the RMS value of current and voltage are known, the power factor (PF) can be calculated by the formula at right. The power factor has no direct connection with the way in which current and voltage are displaced in time. When measuring the input current in connection with installation and service of a system with a frequency converter, it is necessary to use an instrument that is capable of measuring “non-sinusoidal” currents. In general, current measuring instruments for frequency converters must be able to measure “True RMS.”
Frequency converters and earth-leakage circuit breakers Earth-leakage circuit breakers (ELCB) are used as extra protection in electrical installations. If a frequency converter is to be connected, the ELCB installed must be able to brake. If failure occurs on the DC side of the frequency converter, the ELCB must be able to brake. To ensure that the ELCB will brake in case of earth-leakage current, it must be labeled as shown in figures 4.5.7 and 4.5.8 Both types of earth-leakage circuit breakers are available on the market today.
Fig 4.5.7: Labelling of the ELCBfor single-phase frequency converters
Fig 4.5.8: Labelling of the ELCB for three-phase frequency converters
125
Chapter 5. Life cycle costs calculation
Section 5.1: Life cycle costs equation 5.1.1 Initial cost, purchase price 5.1.2 Installation and commissioning costs 5.1.3 Energy costs 5.1.4 Operating costs 5.1.5 Environmental costs 5.1.6 Maintenance and repair costs 5.1.7 Downtime costs (loss of production) 5.1.8 Decommissioning or disposal costs
Section 5.2: Life cycle costs calculation – an example
Energy costs 90%
Maintenance costs 2-5% Initial costs 5-8%
Section 5.1 Life cycle costs equation
In this section the elements that make up a pump’s life cycle costs (LCC) as well as how to calculate LCC will be addressed. Finally, an example will be presented to demonstrate how the LCC formula is applied. The life cycle costs of a pump are an expression of how much it costs to purchase, install, operate, maintain and dispose of a pump during its lifetime.
The Hydraulic Institute, Europump and the US Department of Energy have developed the Pump Life Cycle Costs (LCC) guide, see figure 5.1.1., This tool was designed to help companies minimize waste and maximize energy efficiency in different systems including pumping systems. Life cycle cost calculations aid in decision making associated with design of new or repair of existing installations.
Fig. 5.1.1: A guide to life cycle costs analysis for pumping systems
Typical life cycle costs
Initial costs
The life cycle costs (LCC) consist of the following:
Maintenance costs Energy costs
Cic Cin
Initial cost, purchase price Installation and commissioning costs
Ce Energy costs Co Operating costs including labor Cenv Environmental costs Cm Maintenance and repair costs Cs Downtime costs (loss of production) Cd Decommissioning or disposal costs In the following paragraphs, each of these elements is described. As it appears from figure 5.1.2, energy costs, initial costs and maintenance costs are the most important.
Fig. 5.1.2: Typical life cycle costs of a circulating system in the industry
LCC is calculated by the following formula: LCC = Cic + C in + C e + C o + C m + C s + C env + C d
128
5.1.1 Initial cost, purchase price The initial cost (Cic) of a pump system includes all equipment and accessories necessary to operate the system, such as pumps, frequency converters, control panels and transmitters, see figure 5.1.3.
Initial costs
Often, there is a trade-off between the initial cost and the energy and maintenance costs. For example, expensive components often have a longer lifetime or a lower energy consumption than inexpensive components.
5.1.2 Installation and commissioning costs
Control panels
Pump
Frequency converter
Transmitter
Fig. 5.1.3: Equipment that makes up a pumping system
8000 7000 6000
The installation and commissioning costs (C in) include the following:
5000 4000 3000
• • • •
Installation of the pumps Foundation Connection of electrical wiring and instrumentation Installation, connection and set-up of transmitters and frequency converters, etc • Commissioning evaluation at start-up
2000 1000 0
Initial costs
System 1 5200
System 2 7300
Fig. 5.1.4: Initial costs of a constant speed pump system (System 1) and a controlled pump system (System 2)
As in the case for initial costs, it is important to consider the trade-off options.Pumps with integrated frequency converters often have components integrated in the product. This kind of pump often has a higher initial cost but lower installation and commissioning costs.
129
Section 5.1 Life cycle costs equation
5.1.3 Energy costs
Other use 80%
In most cases, energy consumption (C e) is the largest cost in the life cycle costs of a pump system, where pumps often run more than 2000 hours per year. In fact, around 20% of the world’s electrical energy consumption is used for pump systems, see figure 5.1.5. Some of the factors influencing the energy consumption of a pump system includes: • Load profile • Pump efficiency (calculation of the duty point, see figure 5.1.6) • Motor efficiency (the motor efficiency at partial load can vary significantly between high efficiency motors and normal efficiency motors) • Pump sizing (often margins and round-ups tend to suggest oversized pumps) • Other system components, such as pipes and valves • Use of speed-controlled solutions. By using speedcontrolled pumps in the industry, it is possible to
reduce the energy consumption by up to 50%
5.1.4 Operating costs including labor Operating costs (Co) cover labor costs related to the operation of a pumping system, and, in most cases, are modest. Today, different types of surveillance equipment allow connection of the pump system to a computer network, lowering operating costs.
5.1.5 Environmental costs The environmental costs (Cenv) include the disposal of parts and contamination from the pumped liquid. This contribution to the life cycle costs of a pumping system in the industry is modest.
130
Pump systems 20%
Fig. 5.1.5: Energy consumption worldwide
η
[%]
80
New Existing
60 40 20 0
0
22
44
66
88
11 0
13 2
1 54
1 76
19 8
22 0
242
Q [GPM]
Fig. 5.1.6: Efficiency comparison of a new and an existing existing pump pump
5.1.6 Maintenance and repair costs Maintenance and repair costs (C m) relate to maintenance and repair of the pump system and include: Labor, spare parts, transportation and cleaning. Preventive maintenance will extend pump life, optimize pump performance and prevent pump breakdowns.
5.1.7 Downtime costs (loss of production) Downtime costs (Cs) are extremely important to pump systems used in production processes. Production stoppage is costly, even for a short period of time. Though one pump may be enough for the application, it is a good idea to install a standby pump that can take over in the event of an unexpected failure, see figure 5.1.7.
5.1.8 Decommissioning or disposal costs Depending on the pump manufacturer, decommissioning or disposal costs (Cd ) of a pump system varies. This cost is seldom taken into consideration when calculating LCC.
Calculating the life cycle costs The life cycle costs of a pump system are made up of the summation of the aforementioned components over the system’s lifetime. Typically, the lifetime range is 10 to 20 years. In the pump business, life cycle costs are normally calculated by a simplified formula with fewer elements to consider. This formula is shown at right.
Fig. 5.1.7: A standby pump assures that production continues in case of pump breakdown
Simplified: LCC = Cic + C e + C m Cic Initial costs, purchase price Ce
Energy costs
Cm
Maintenance and repair costs
131
Section 5.2 Life cycle costs calculation – an example
The example using the LCC formula mentioned on the previous page follows: An industry needs a new water supply pump and two solutions are taken into consideration:
Pump types Average power consumption Operating hours per day Working days per year Calculation period Total energy consumption
kw hours days years
USD/kwh .07
• A variable speed multistage centrifugal pump
Pump price Maintenance costs Energy costs
USD USD USD
Life cycle costs calculations will help determine which pump to install in the system. The application has the following characteristics: • 12 operating hours per day • 220 operating days per year • Lifetime of 10 years (calculation period)
11.31 12 220 10 298,584
.07 3602 1417 33,284
USD
45,000 40,000
7204 1417 20,066
38,303
28,688
Pump price Maintenance costs Energy costs
35,000 30,000 D S U
18.76 12 220 10 495,264
Electrical power price
Total costs
Variable speed
kwh
• A fixed speed multistage centrifugal pump
According to the calculations, the variable speed pump consumes 40% less energy than the fixed speed pump. However, the initial cost, C ic, of the variable speed pump is twice that of the fixed speed pump.
Fixed speed
25,000 20,000 15,000 10,000 5,000 0
Fixed speed
Variable speed
Fig. 5.1.8: Life cycle costs of a fixed and a variable speed pump
Based on this data, it is possible to calculate the life cycle costs of the two solutions. Even though the initial cost of a variable speed pump is twice as high as a fixed speed pump, the total cost of the variable speed solution is 25% lower than the fixed speed pump solution after 10 years. Besides the lower life cycle costs the variable speed pump provides, as discussed in chapter 4, some operational benefits, such as constant pressure in the system.
45,000 40,000 35,000 30,000 D S U
25,000 20,000 15,000 Fixed speed Variable speed
10,000 5,000
The payback time of the variable speed pump solution is a bit longer because the pump is more expensive. As you can tell from figure 5.1.9, the payback time is around 2½ years, and in general industrial applications, this is considered to be a good investment.
132
0 0
2
4
Years
6
8
Fig. 5.1.9: Payback time for a fixed and a variable speed pump
10
Appendix A) Notations and units B) Unit conversion tables C) SI-prefixes and Greek alphabet D) Vapor pre pressure ssure and specific specific gra gravity vity of water water at different different temperatures temperatures E) Orifice F) Change in static pressure due to change in pipe diameter G) Nozzles H) Nomogram for head losses in bends, valves, etc. I) Periodical system J) Pump standards K) Viscosity for different liquids as a function of liquid temperature
Appendix A
Notations and units The table below provides an overview of the most commonly used notations and units for pumps and pump systems.
U.S.
SI
unit
unit
ft GPM psi psi
gph ft
psi
ft
ft lb ft
lb gal
ft lb ft
in in ft
g = 32.174 ft/s m
ft RPM hp
134
745.7 w = 1 hp
Appendix B
Unit conversion tables The conversion tables for pressure and flow show the most commonly used units for pumping systems
CONVERSION FACTORS - UNITS OF LENGTH Examples: 2 Yards x 3 = 6 Feet x 0.333 = 1 Yard
Unit Inch Foot Yard Inch 1 0.0833 0.0278 Foot 12 1 0.333 Yard 36 3 1 Centimeters 0.3937 0.0328 0.0109 Meter 39.37 3.281 1.094 1 Mile = 5280 ft. = 1760 yards = 1609.3 meters = 1.61 Kilometers 1 Kilometer = 1000 meters = 1093.6 yards = .62137 miles
Centimeter 2.54 30.48 91.44 1 100
Meter 0.0254 0.3048 0.9144 0.01 1
CONVERSION FACTORS - UNITS OF FLOW Examples: 500 U.S. Gpm x .00144 = .72 U.S. Mgd. x 694.5 = 6945 U.S. Gpm U.S. Im p. Unit Gpm Gpm U.S.Gal./Min. 1 0.833 Imp. Gal./Min. 1.2 1 U.S. Mgd (2) 694.4 578.7 Imperial Mgd (2) 833.4 694.5
U.S. Mgd (2) 0.00144 0.00173 1 1.2
Im peria l M gd(2) 0.0012 0.00144 0.833 1
Cu. Ft. Cu. Meters Lit ers Barr els/Min . Barr els/24 Hrs. /Sec. /Hr. /Sec. (3) (3) 0.00223 0.227 0.0631 0.0238 34.25 0.00268 0.272 0.0757 0.0286 41.09 1.547 157.73 43.8 16.53 23786.6 1.856 189.28 52.56 19.83 28544
Cu. Ft./Sec. 448.8 374 0.646 0.538 1 101.9 28.32 10.68 Cu. Meters/Hr. 4.403 3.67 0.00634 0.00528 0.00981 1 0.2778 0.1047 Liters/Sec. 15.85 13.21 0.0228 0.019 0.0353 3.6 1 0.377 Barrels/Min. (3) 4 2 34.99 0.0605 0.0504 0.0937 9.534 2.65 1 Barrels/24 Hrs.(3) 0.0292 0.0243 0.000042 0.000035 0.000065 0.00662 0.00184 0.000694 (1) US Mgd = Million U.S. gallons per 24 hr. day. Imp Mgd = Million Imperial gallons per 24 hr. day. (2) 42 gal. bbl.
15360.4 150.8 542.86 1440 1
CONVERSION FACTORS - UNITS OF PRESSURE Examples: 15 Ft. Water x .433 = 6.49 Psi 15 Psi x 2.31 = 34.65 Ft. Water
In.Water Ft. Water Psi. In. Hg. Mm.Hg. Bar atm Kilopascal
In. Water 1 12 27.72 13.596 0.5353 401.86
407.19 4.0186
Ft. Water 0.0833 1 2.31 1.133 0.0446 33.49
33.93 0.3349
Psi 0.0361 0.433 1 0.4906 0.0193 14.503
In. Hg. 0.0736 0.883 2.04 1 0.03937 29.54
Mm. Hg. 1.87 22.43 51.816 25.4 1 750.5
14.696 0.1451
29.92 0.2954
760 0.7505
Bar 2.538 30.45 70.31 34.49 1.357 1
1.0133 —
atm 0.0025 0.0304 0.0703 0.0345 0.0014 0.987
1
—
135
Appendix C
SI-prefixes and Greek alphabet Factor 109 106 103 102 10 10-1 10-2 10-3 10-6 10-9
Prefix 1,000,000,000 1,000,000 1,000 100 10 0.1 0.01 0.001 0.000.001 0.000.000.001
Symbol
giga mega kilo hekto deka deci centi milli mikro nano
G M k h da d c m µ
n
Greek alphabet
136
Alfa
Α
Beta
Β
β
Gamma
Γ
γ
Delta
∆
δ
Epsilon
Ε
ε
Zeta
Ζ
ζ
Eta
Η
η
Theta
Θ
θ
Jota
Ι
ι
Kappa
Κ
κ
Lambda
Λ
λ
My
Μ
µ
Ny
Ν
ν
Ksi
ΚΣ
κσ
Omikron
Ο
ο
Pi
Π
π
Rho
Ρ
ρ
Sigma
Σ
σ
Tau
Τ
τ
Ypsilon
Υ
υ
Fi Khi
Φ Χ
φ χ
Psi
Ψ
ψ
Omega
Ω
ω
α
Appendix D
Vapor pressure pressure an and d sp specific ecific gr gravity avity of ofwwater ater at atdidifferent fferent tetemperatures mperatures This table shows the specific gravity [sg], vapor pressure p [psi] and the density ρ [lb/ft3] of water at different temperatures t [oF].
Properties of Water at Various Temperatures WATER TEMPERATURE 0
F
SPECIFIC GRAVITY
0
C
VAPOR PRESSURE
PSIA
DENSITY
lb/ft
FEET
32
0
1.002
0.0886
0.204
40
4.4
1.001
0.1217
0.281
62.425
45
7.2
1.001
0.1474
0.340
62.420
50
10.0
1.001
0.1780
0.411
62.410
55
12.8
1.000
0.2139
0.494
62.390
60
15.6
1.000
0.2561
0.591
62.370
65
18.3
.999
0.3056
0.706
62.340
70
21.1
.999
0.3629
0.839
62.310
75
23.9
.998
0.4296
0.994
62.270
80
26.7
.998
0.5068
1.172
62.220
85
29.4
.997
0.5958
1.379
62.170
90
32.2
.996
0.6981
1.617
62.120
95
35.0
.995
0.8153
1.890
62.060
100
37.8
.994
0.9492
2.203
62.000
110
43.3
.992
1.2750
2.965
61.980
120
48.9
.990
1.6927
3.943
61.710
130
54.4
.987
2.2230
5.196
61.560
140
60.0
.985
2.8892
6.766
61.380
150
65.6
.982
3.7184
8.735
61.190
160
71.1
.979
4.7414
11.172
60.990
170
76.7
.975
5.9926
14.178
60.790
180
82.2
.972
7.5110
17.825
60.570
190
87.8
.968
9.3400
22.257
60.340
200
93.3
.964
11.5260
2 7.584
60.110
210
98.9
.960
14.1230
3 3.983
59.860
212
100.0
.959
14.6960
35.353
59.810
220
104.4
.956
17.1860
41.343
59.610
230
110.0
.952
20.7790
50.420
59.350
240
115.6
.948
24.9680
60.770
59.080
250
121.1
.943
29.8250
73.060
58.800
260
126.7
.939
35.4300
87.050
58.520
270
132.2
.933
41.8560
1 03.630
58.220
280
137.8
.929
49.2000
1 22.180
57.920
290
143.3
.924
57.5500
1 43.875
57.600
3
62.400
137
Appendix E
Orifice Nipple orifices are typically used in boiler feed applications when boiler feed pumps need to discharge built-up pressure. These boiler feed pumps operate continuosly in order to provide on-demand hot water; but when no hot water is needed, the valve to the boiler
allows water to flow back to the reservoir tank. During feed pump system design, nipple orifices are sized using performance charts, like the ones shown in the figure below, derived from an acceptable mathematical approach that assumes aconstant discharge coefficient
is closed and the pump ends up operating under a harmful shut-off condition during extended periods of time in which there will be a rise in liquid temperature in the pump because the input horsepower being converted to heat in the pump is not dissipated. For that reason, in order to increase the run life of the pump and control the temperature rise, the system is designed to allow the feed pump to discharge its build-up pressure through a bypass line in which a nipple orifice is installed. The orifice dissipates the high pressure and
(Cd) of 0.61 for all orifices in he t general equation Q = 19.636 Cd d2 H0.5, where Q is in gpm, d is the nipple orifice diameter in inches, and H is the differential head in ft. of water.
Orifice size
Approximate Discharge Through Bypass Nipple Orifice 1000 1/8"
3/16"
1/4"
5/16"
1/2"
1"
7/16"
7/8"
3/8"
13/16" 3/4" 11/16" 5/8" 9/16"
) t e e F ( 100 d a e H
10 1
10
100
Flow (GPM)
138
1000
Appendix F
Change in static pressure due to change in pipe diameter As described in Chapter 2.2, a change in pipe dimension results in a change in liquid velocity and consequently, a change in dynamic and static pressure. When head has to be determined (see page 86), the difference in the two port dimensions requires a correction of the measured head.
Approximate Sudden Expansion Head Loss 100 d
D
10 ] t f [ H
1
0.1
10
100
1000
10000
Q[GPM]
d/D=1/1.5
d/D=1.25/2
d/D=2/2.5
d/D=2/3
d/D=2.5/3
d/D=4/5
d/D=4/6
d/D=5/6
d/D=5/8
d/D=6/8
d/D=10/14
d/D=12/14
d/D=12/16
d/D=14/16
d/D=14/18
d/D=2.5/4 d/D=8/10 d/D=16/18
d/D=3/4
d/D=3/5
d/D=8/12
d/D=10/12
d/D=16/20
d/D=18/20
Approximate Sudden Contraction Head Loss
100 D
d
10 ] t f [ H
1
0.1
10
100
1000
10000
Q[GPM]
139
Appendix G
Nozzles The relationship between the nozzle diameter d [inches], the needed flow Q [GPM] and the required pressure before the nozzle p [psi] is found by the nomogram below. We assume that the nozzle has a quadratic behavior, and d / D is less than 1/3.
Q1 = Q2
Nozzle Diameter (inch) 1
3/8
( pp ) 1
1/8 1/4 1/2
n 5/8
2
7/8
where n = 0.5. Some nozzles have a lower n value (check with the supplier). Pressure p [psi]
Nozzle diameter d [inch]
Flow Q [GPM]
D
/16 3/16
3/4 1
1 1/8
1 1/4
1 3/8
1 1/2
1 3/4
2
2 1/4
2 1/2
2 3/4
3
3 1/2
4
4 1/2
5
5 1/2
6
Approximate discharge of a nozzle 400
I S P
100
10
5 0.1
1
10
100
Q [GPM]
140
1000
10000
100000
Appendix H
Friction Loss for Water in New Sch. 40 Steel Pipe at 60° F (Frict. loss in ft. per 100 ft. - Vel. in ft. per sec.) gpm
1/8” (0.26 ID) 1/4” (0.36 ID) Vel. Frict. Vel. Frict.
3/8” (0.49 ID) Vel. Frict.
1/2” (0.62 ID) Vel. Frict.
3/4” (0.82 ID) 1” (1.04 ID) Vel. Frict. Vel. gpm Frict. 0.1
0.1
0.57
1.36
0.31
0.41
0.2
1.13
2.72
0.62
0.81
0.3
1.69
9.70
***
***
0.4
2.26
16.2
1.23
3.70
0.5
2.82
24.2
***
***
0.84
1.26
0.6
3.39
33.8
1.85
7.60
1.01
1.74
0.7
3.95
44.8
***
***
***
***
0.8
4.52
57.4
2.47
12.7
1.34
2.89
0.9
5.08
71.6
***
***
***
***
1.0
5.65
87.0
3.08
19.1
1.68
4.30
1.06
1.86
0.60
0.26
1.2
6.77
122
3.70
26.7
***
***
***
***
***
***
1.4
***
***
4.32
35.3
***
***
***
***
***
***
1.5
***
***
***
***
2.52
8.93
1.58
2.85
0.90
0.73
1.6
***
***
4.93
45.2
***
***
***
***
***
***
1.8
***
***
5.55
56.4
***
***
***
***
***
***
2.0
***
***
6.17
69.0
3.36
15.0
2.11
4.78
1.20
1.21
2.5
***
***
7.71
105
4.20
22.6
2.64
7.16
***
***
3.0
***
***
9.25
148
5.04
31.8
3.17
10.0
1.81
2.50
3.5
***
***
10.8
200
5.88
42.6
3.70
13.3
***
***
4.0
***
***
12.3
259
6.72
54.9
4.22
17.1
2.41
4.21
4.5
***
***
13.9
326
7.56
68.4
4.75
21.3
***
***
5.0 5.5
*** ***
*** ***
15.4 ***
396 ***
8.40 9.24
83.5 100
5.28 5.81
29.8 30.9
3.01 ***
6.32 ***
6.0
***
***
***
***
10.1
118
6.34
36.5
3.61
8.87
6.5
***
***
***
***
***
***
6.86
42.4
***
***
7.0
***
***
***
***
11.8
158
7.39
48.7
4.21
11.8
7.5
***
***
***
***
***
***
7.92
55.5
***
***
8.0
***
***
***
***
13.4
205
8.45
62.7
4.81
15.0
8.5
***
***
***
***
***
***
8.98
70.3
***
***
9.0
***
***
***
***
15.1
258
9.50
78.3
5.42
18.8
9.5
***
***
***
***
***
***
10.0
86.9
***
***
10.0
***
***
***
***
16.8
316
10.6
95.9
6.02
27.0
12
***
***
***
***
***
***
12.7
136
7.22
32.6
14
***
***
***
***
***
***
14.8
183
8.42
43.5
15
***
***
***
***
***
***
***
***
9.02
50.0
16
***
***
***
***
***
***
***
***
9.63
56.3
18
***
***
***
***
***
***
***
***
10.8
70.3
20
***
***
***
***
***
***
***
***
12.0
86.1
25
***
***
***
***
***
***
***
***
15.1
134
30
***
***
***
***
***
***
***
***
18.1
187
35
***
***
***
***
***
***
***
***
***
***
40 45
*** ***
*** ***
*** ***
*** ***
*** ***
*** ***
*** ***
*** ***
*** ***
50
***
***
***
***
***
***
***
***
***
60
***
***
***
***
***
***
***
***
***
*** *** *** ***
0.37 *** *** *** *** *** 0.74 *** 1.11 *** 1.48 ***
0.11 *** *** *** *** *** 0.38 *** 0.78 *** 1.30 ***
0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1.0 1.2 1.4 1.5 1.6 1.8 2.0 2.5 3.0 3.5 4.0 4.5
1.86 *** 2.23 *** 2.60 *** 2.97 *** 3.34 *** 3.71 4.45 5.20 *** 5.94 6.68 7.42 9.27 11.1 13.0 14.8 16.7 18.6 22.3
1.93 *** 2.68 *** 3.56 *** 4.54 *** 5.65 *** 6.86 9.62 12.8 *** 16.5 20.6 25.1 38.7 54.6 73.3 95.0 119 146 209
5.0 5.5 6.0 6.5 7.0 7.5 8.0 8.5 9.0 9.5 10.0 12 14 15 16 18 20 25 30 35 40 45 50 60
141
Appendix H
Friction Loss for Water in New Sch. 40 Steel Pipe at 60° F (Frict. loss in ft. per 100 ft. - Vel. in ft. per sec.) gpm
142
1 1/4” (1.38 ID) Vel. Frict.
1 1/2” (1.61 ID) Vel. Frict.
2” (2.07 ID) Vel. Frict.
2 1/2” (2.47 ID) Vel. Frict.
3” (3.07 ID) 4” (4.07 ID) Vel. Frict.Vel. gpm Frict.
5
1.07
0.52
0.79
0.25
0.48
0.07
10
2.15
1.77
1.58
0.83
0.96
0.25
0.67
0.10
0.43
0.04
12 14
2.57 3.00
2.48 3.28
1.89 2.21
1.16 1.53
1.15 1.34
0.35 0.46
0.80 0.94
0.15 0.20
*** ***
*** ***
16
3.43
4.20
2.52
1.96
1.53
0.59
1.07
0.25
***
***
18
3.86
5.25
2.84
2.42
1.72
0.73
1.21
0.31
***
***
20
4.29
6.34
3.15
2.94
1.91
0.87
1.34
0.36
0.87
0.13
0.50
25
5.37
9.66
3.94
4.50
***
***
***
***
***
***
***
30
6.44
13.6
4.73
6.26
2.87
1.82
2.01
0.75
1.30
0.27
0.76
35
7.52
18.5
5.52
8.38
3.35
2.42
2.35
1.00
***
***
***
40
8.58
23.5
6.30
10.8
3.82
3.10
2.68
1.28
1.82
0.55
1.01
45
9.66
29.5
7.10
13.5
4.30
3.82
3.02
1.57
***
***
***
50
10.7
36.0
7.88
16.4
4.78
4.67
3.35
1.94
2.17
0.66
1.26
60
12.9
51.0
9.46
23.2
5.74
6.59
4.02
2.72
2.60
0.92
1.51
70
15.0
68.8
11.0
31.3
6.69
8.86
4.69
3.63
***
***
1.76
80
17.2
89.2
12.6
40.5
7.65
11.4
5.36
4.66
3.47
1.57
2.02
90
19.3
112
14.2
51.0
8.60
14.2
6.03
5.82
***
***
2.27
100
21.5
138
15.8
62.2
9.56
17.4
6.70
7.11
4.34
2.39
2.52
120
25.7
197
18.9
88.3
11.5
24.7
8.04
10.0
5.21
3.37
3.02
140
30.0
267
22.1
119
13.4
33.2
9.38
13.5
6.08
4.51
3.53
160
***
***
25.2
158
15.3
43.0
10.7
17.4
6.94
5.81
4.03
180
***
***
28.4
199
17.2
54.1
12.1
21.9
7.81
7.28
4.54
200 220
*** ***
*** ***
31.5 ***
241 ***
19.1 21.0
66.3 80.0
13.4 14.7
26.7 32.2
8.68 9.55
8.90 10.7
5.04 5.54
240
***
***
***
***
22.9
95.0
16.1
38.1
10.4
12.6
6.05
260
***
***
***
***
24.9
111
17.4
44.5
11.3
14.7
6.55
280
***
***
***
***
26.8
128
18.8
51.3
12.2
16.9
7.06
300
***
***
***
***
28.7
146
20.1
58.5
13.0
19.2
7.56
350
***
***
***
***
***
***
23.5
79.2
15.2
26.3
8.82
400
***
***
***
***
***
***
26.8
103
17.4
33.9
10.1
450
***
***
***
***
***
***
30.2
132
19.6
43.0
11.3
500
***
***
***
***
***
***
33.5
160
21.7
52.5
12.6
550
***
***
***
***
***
***
***
***
23.9
63.8
13.9
600
***
***
***
***
***
***
***
***
26.0
75.7
15.1
650
***
***
***
***
***
***
***
***
28.2
88.6
16.4
700
***
***
***
***
***
***
***
***
30.4
101
17.6
750
***
***
***
***
***
***
***
***
***
***
18.9
800
***
***
***
***
***
***
***
***
34.7
131
20.2
850
***
***
***
***
***
***
***
***
***
***
900
***
***
***
***
***
***
***
***
***
***
21.4 22.7 23.9 25.2 27.7 30.2 35.3
950
***
***
***
***
***
***
***
***
***
***
1000
***
***
***
***
***
***
***
***
***
***
1100
***
***
***
***
***
***
***
***
***
***
1200
***
***
***
***
***
***
***
***
***
***
1400
***
***
***
***
***
***
***
***
***
***
0.04 *** 0.07 *** 0.12 *** 0.18 0.25 0.33 0.42 0.52 0.61 0.86 1.16 1.49 1.89 2.27 2.70 3.19 3.72 4.28 4.89 6.55 8.47 10.5 13.0 15.7 18.6 21.7 25.3 28.9 32.8 37.0 41.4 46.0 50.9 61.4 72.0 97.6
5 10 12 14 16 18 20 25 30 35 40 45 50 60 70 80 90 100 120 140 160 180 200 220 240 260 280 300 350 400 450 500 550 600 650 700 750 800 850 900 950 1000 1100 1200 1400
Appendix H
Friction Loss for Water in New Sch. 40 Steel Pipe at 60° F (Frict. loss in ft. per 100 ft. - Vel. in ft. per sec.) gpm
5” (5.05 ID) Vel. Frict.
6” (6.07 ID) Vel. Frict.
8” (7.98 ID) Vel. Frict.
10” (10.02 ID) Vel. Frict.
12” (11.94 ID) 14” (13.12 ID) Vel. Frict. Vel. gpm Frict.
40
0.64
0.04
60
0.96
0.08
80 100
1.28 1.60
0.14 0.21
1.11
0.09
120
1.92
0.29
1.33
0.12
140
2.25
0.39
1.55
0.16
160
2.57
0.48
1.78
0.20
180
2.89
0.60
2.00
0.25
1.15
0.07
200
3.21
0.73
2.22
0.30
1.28
0.08
220
3.53
0.87
2.44
0.36
1.41
0.10
240
3.85
1.03
2.66
0.41
1.54
0.11
260
4.17
1.19
2.89
0.48
1.67
0.13
280
4.49
1.37
3.11
0.54
1.80
0.15
300
4.81
1.58
3.33
0.62
1.92
0.17
1.22
0.06
350
5.61
2.11
3.89
0.85
2.24
0.22
1.42
0.07
400
6.41
2.72
4.44
1.09
2.57
0.28
1.63
0.09
450
7.22
3.41
5.00
1.36
2.89
0.34
1.83
0.12
500
8.02
4.16
5.55
1.66
3.21
0.42
2.03
0.14
1.43
550
8.81
4.94
6.11
1.97
3.53
0.50
2.24
0.17
1.58
600
9.62
5.88
6.66
2.33
3.85
0.59
2.44
0.20
1.72
700
11.2
7.93
7.77
3.13
4.49
0.79
2.85
0.25
2.01
800
12.8
10.2
8.88
4.04
5.13
1.01
3.25
0.33
2.29
900
14.4
12.9
9.99
5.08
5.77
1.27
3.66
0.41
2.58
1000
16.0
15.8
11.1
6.23
6.41
1.55
4.07
0.49
2.87
40 60
1100
***
***
12.2
7.49
7.05
1.86
4.48
0.59
3.15
1200
19.2
22.5
13.3
8.87
7.70
2.20
4.88
0.70
3.44
1300
***
***
14.4
10.4
8.34
2.56
5.29
0.81
3.73
1400
22.5
30.4
15.5
12.0
8.98
2.96
5.70
0.94
4.01
1500
***
***
16.7
13.7
9.62
3.38
6.10
1.07
4.30
1600
25.7
39.5
17.8
15.6
10.3
3.83
6.51
1.21
4.59
1700
***
***
18.9
17.5
10.9
4.29
6.92
1.38
4.87
1800
28.8
49.7
20.0
19.6
11.5
4.81
7.32
1.52
1900
***
***
21.1
21.8
12.2
5.31
7.73
1.68
2000
32.1
61.0
22.2
24.1
12.8
5.91
8.14
1.86
2500
***
***
27.7
37.2
16.0
8.90
10.2
2.86
3000
***
***
***
***
19.2
12.8
12.2
4.06
3500
***
***
***
***
22.4
17.5
14.2
5.46
4000
***
***
***
***
25.7
22.0
16.3
7.07
4500
***
***
***
***
***
***
18.3
8.91
5000
***
***
***
***
***
***
20.3
11.0
6000
***
***
***
***
***
***
24.4
15.9
7000 8000
*** ***
*** ***
*** ***
*** ***
*** ***
*** ***
*** ***
*** ***
9000
***
***
***
***
***
***
***
***
10000
***
***
***
***
***
***
***
***
5.16 5.45 5.73 7.17 8.60 10.0 11.5 12.9 14.3 17.2 20.1 22.9 *** ***
80 100 120 140 160 180 200 220 240 260 280 300 350 400 450 500 550 600 700
0.06 0.07 0.08 0.11 0.14 0.18 0.21 0.25 0.29 0.34 0.39 0.44 0.50 0.57 0.64 0.70 0.78 1.19 1.68 2.25 2.92 3.65 4.47 6.39 8.63 11.2 *** ***
1.90 2.14 2.37 2.61 2.85 3.08 3.32 3.56 3.80 4.03 4.27 4.51 4.74 5.93 7.12 8.30 9.49 10.7 11.9 14.2 16.6 19.0 21.4 23.7
0.09 0.11 0.13 0.16 0.18 0.21 0.24 0.28 0.31 0.35 0.39 0.43 0.48 0.73 1.04 1.40 1.81 2.27 2.79 4.00 5.37 6.98 8.79 10.8
800 900 1000 1100 1200 1300 1400 1500 1600 1700 1800 1900 2000 2500 3000 3500 4000 4500 5000 6000 7000 8000 9000 10000
143
Appendix H
Friction Loss for Water in New Sch. 40 Steel Pipe at 60° F (Frict. loss in ft. per 100 ft. - Vel. in ft. per sec.) gpm
16” (15.00 ID) Vel. Frict.
1000 1500
1.82 2.72
0.07 0.14
2000 2500 3000 3500 4000 4500 5000 6000 7000 8000 9000 10,000 12,000 14,000 16,000 18,000 20,000 25,000 30,000 35,000 40,000 50,000
3.63 4.54 5.45 6.35 7.26 8.17 9.08 10.9 12.7 14.5 16.3 18.2 21.8 25.4 29.0 *** *** *** *** *** *** ***
0.25 0.38 0.54 0.72 0.92 1.15 1.41 2.01 2.69 3.49 4.38 5.38 7.69 10.4 13.5 *** *** *** *** *** *** ***
16” (15.00 ID) Vel. Frict.
20” (18.81) Vel. Frict.
24” (22.62 ID) Vel. Frict.
30” (29.00 ID)* 36” (35.00 ID)* Vel. Frict. Vel. Frict.
gpm 1000 1500
2.87 3.59 4.30 5.02 5.74 6.45 7.17 8.61 10.0 11.5 12.9 14.3 17.2 20.1 22.9 25.8 28.7 *** *** *** *** ***
0.14 0.21 0.30 0.40 0.51 0.64 0.78 1.11 1.49 1.93 2.42 2.97 4.21 5.69 7.41 9.33 11.5 *** *** *** *** ***
2.31 2.89 3.46 4.04 4.62 5.19 5.77 6.92 8.08 9.23 10.4 11.5 13.8 16.2 18.5 20.8 23.1 28.9 34.6 *** *** ***
0.08 0.12 0.17 0.23 0.30 0.37 0.46 0.65
2.39
0.07
2.79
0.09
3.19
0.12
3.59
0.15
***
***
3.99
0.18
2.43
0.05
4.79
0.26
2.91
0.86
5.59
0.34
3.40
1.11
6.38
0.44
3.89
1.39
7.18
0.55
4.37
1.70
7.98
0.67
4.86
2.44
9.58
0.96
5.83
3.29
11.2
1.29
6.80
4.26
12.8
1.67
7.77
5.35
14.4
2.10
8.74
6.56
16.0
2.58
9.71
10.2
20.0
4.04
12.1
14.6
23.9
5.68
14.6
***
27.9
7.73
17.0
***
***
***
19.4
***
***
***
***
0.08 0.10 0.13 0.16 0.20 0.28 0.37 0.48 0.60 0.73 1.13 1.61 2.17 2.83 ***
1.94
0.03
1.58 1.89 2.21 2.52 2.84 3.15 3.78 4.41 5.04 5.67 6.30 7.88 9.46 11.0 12.6 15.8
0.02 0.03 0.04 0.04 0.06 0.07 0.09 0.13 0.16 0.20 0.25 0.38 0.54 0.72 0.94 1.45
2000 2500 3000 3500 4000 4500 5000 6000 7000 8000 9000 10,000 12,000 14,000 16,000 18,000 20,000 25,000 30,000 35,000 40,000 50,000
Note: 1. Table based on Darcy-Weisback formula; with no allowance for age, differences in diameter, or any other abnormal condition of interior surface. Any Factor of Safety must be estimated from the local conditions and the requirements of each particular installation. For general purposes, 15% is a reasonable Factor of Safety. 2. The friction loss data is based on seamless Sch. 40 steel pipe. Cast iron (CI) pipe has a slightly larger ID than steel pipe in the 3” to 12” dia. range, which generally makes no practical difference with respect to water supply pumping problems. 3. Ductile Iron (DI) has a larger ID than both Sch. 40 steel and CI pipes for the same nominal diameter. Friction Losses in DI pipe can be approximated by multiplying the tabulated value by .75 in the 4” to 12” size range and .60 for 14” and larger sizes. 4. Velocity head values are not included in the table, as they are normally not considered as a component of Total Head (TH) calculations to solve water supply pumping problem. Velocity and Velocity head can be calculated using the following formulas: Vel. (fps) = gpm (.410)/(ID) 2 = gpm (.321)/Area (in. 2); where: Area (in 2) = Vel. Head (ft.) = (Vel.) 2 /2g = (Vel.) 2/64.4
π
(ID) 2/4
5. Velocity within column (vertical drop/riser pipe) shoul d be kept within the range of 415 - fps (5.0 fps is optimum). Velocity within horizontal distribution piping should be kept within the range of 1 - 6 fps (3.0 fps is optimum). 6. Tabulated friction loss values are calculated based on water at 60°F and a kinematic viscosity = 0.00001217 ft /sec. (31.2 SSU). Correct tabulated values for fluid temperatures other than 60°F as following: Temp(°F) 32 40 50 60 80 100 150 200 212 Correctionfactor 1.20 1.10 1.00 1.00 1.00 .95 .90 .85 .80 * The ID value specified for 30”and 36” sizes are forSch. 20 pipe. Sch. 40 pipe is not available indiameters greater than 24”
144
Appendix H
Friction Loss for Water in New Type L. Copper Tubing and Sch. 40 PVC Pipe (Frict. loss in ft. per 100 ft. - Vel. in ft. per sec.) 1/2” gpm
Tubing .545”ID Vel. Frict.
Pipe .622”ID Vel. Frict.
3/4” gpm
Tubing .785”ID Vel. Frict.
Pipe .824”ID Vel. Frict.
0.5 1.0 1.5
0.69 1.38 2.06
0.75 2.45 4.93
0.52 1.04 1.57
0.40 1.28 2.58
12 3
0.66 1.33 1.99
0.44 1.44 2.91
0.60 1.21 1.81
0.35 1.16 2.34
2.0 2.5 3.0
2.75 3.44 4.12
8.11 11.98 16.48
2.09 2.61 3.13
4.24 6.25 8.59
4 5 6
2.65 3.31 3.98
4.81 7.11 9.80
2.42 3.02 3.62
3.86 5.71 7.86
3.5 4.0 4.5
4.81 5.50 6.19
21.61 27.33 33.65
3.66 4.18 4.70
11.25 14.22 17.50
7 8 9
4.64 5.30 5.96
12.86 16.28 20.06
4.23 4.83 5.44
10.32 13.07 16.10
5.0 6.0 7.0
6.87 8.25 9.62
40.52 56.02 73.69
5.22 6.26 7.31
21.07 29.09 38.23
10 11 12
6.92 7.29 7.95
24.19 28.66 33.47
6.04 6.64 7.25
19.41 22.99 26.84
8.0 9.0 10.0
11.0 12.4 13.8
93.50 115.4 139.4
8.35 9.40 10.4
48.47 59.79 72.16
13 14 15
8.61 9.27 9.94
38.61 44.07 49.86
7.85 8.45 9.05
30.96 35.33 39.97
12.6 14.7
115.6 157.4
16 17 18
10.60 11.25 11.92
55.97 62.39 69.13
9.65 10.25 10.85
44.86 50.00 55.40
12.0 14.0 16.0 1” gpm
Tubing 1.03”ID Vel. Frict.
Pipe 1.05”ID Vel. Frict.
11/4” gpm
Tubing 1.27”ID Vel. Frict.
Pipe 1.38”ID Vel. Frict.
2 3 4
0.78 1.17 1.56
0.41 0.82 1.35
0.72 1.08 1.45
0.35 0.70 1.14
5 6 7
1.28 1.53 1.79
0.74 1.01 1.32
1.09 1.31 1.53
0.51 0.70 0.91
5 6 7
1.95 2.34 2.72
2.00 2.75 3.60
1.81 2.17 2.53
1.69 2.32 3.04
8 9 10
2.04 2.30 2.55
1.67 2.06 2.48
1.75 1.96 2.18
1.15 1.42 1.71
8 9 10
3.11 3.50 3.89
4.56 5.61 6.76
2.89 3.25 3.61
3.85 4.74 5.71
12 15 20
3.06 3.83 5.10
3.42 5.07 8.46
2.62 3.27 4.36
2.35 3.49 5.81
12 14 16
4.67 5.45 6.22
9.33 12.27 15.56
4.34 5.05 5.78
7.88 10.36 13.13
25 30 35
6.38 7.65 8.94
12.59 17.44 23.00
5.46 6.55 7.65
8.65 11.98 15.79
18 20
7.00 7.78
19.20 23.18
6.50 7.22
16.20 19.55
40 45
10.2 11.5
29.24 36.15
8.74 9.83
20.06 24.80
25 30 35 40
9.74 11.68 13.61 15.55
34.56 47.96 63.31 80.58
9.03 10.84 12.65 14.45
29.15 40.43 53.37 67.90
50 60 70 80
12.8 15.3 17.9 20.4
43.71 60.78 80.38 102.5
10.9 13.1 15.3 17.5
29.98 41.66 55.07 70.16
145
Appendix H
Friction Loss for Water in New Type L. Copper Tubing and Sch. 40 PVC Pipe (Frict. loss in ft. per 100 ft. - Vel. in ft. per sec.) 11/2” gpm
Pipe 1.61”ID Vel. Frict.
2” gpm
Tubing 1.98”ID Vel. Frict.
Pipe 2.07”ID Vel. Frict.
8
1.44
0.73
1.27
0.55
16
1.66
0.66
1.53
0.55
9 10
1.62 1.80
0.90 1.08
1.43 1.59
0.67 0.81
18 20
1.87 2.07
0.82 0.98
1.72 1.92
0.68 0.82
12 15 20
2.16 2.70 3.60
1.49 2.21 3.68
1.91 2.39 3.19
1.12 1.65 2.75
25 30 35
2.59 3.11 3.62
1.46 2.01 2.65
2.39 2.87 3.35
1.22 1.68 2.21
25 30 35
4.51 5.41 6.31
5.48 7.58 9.99
3.98 4.78 5.58
4.09 5.65 7.45
40 45 50
4.14 4.66 5.17
3.36 4.15 5.01
3.83 4.30 4.80
2.80 3.46 4.17
40 45 50
7.21 8.11 9.01
12.68 15.67 18.94
6.37 7.16 7.96
9.45 11.68 14.11
60 70 80
6.21 7.25 8.28
6.95 9.16 11.65
5.75 6.70 7.65
5.79 7.63 9.70
60 70 80
10.8 12.6 14.4
26.30 34.74 44.24
9.56 11.2 12.8
19.59 25.87 32.93
90 100 110
9.31 10.4 11.4
14.41 17.43 20.71
8.61 9.57 10.5
12.00 14.51 17.24
90 100 110
16.2 18.0 19.8
54.78 66.34 78.90
14.4 15.9 17.5
40.76 79.34 58.67
120 130 140
12.4 13.4 14.5
24.25 28.04 32.07
11.5 12.5 13.4
20.18 23.33 26.69
21/2”
146
Tubing 1.51”ID Vel. Frict.
Tubing 2.46”ID
Pipe 2.47”ID
gpm 20 25 30
Vel. 1.34 1.68 2.02
Frict. 0.35 0.52 0.72
Vel. 1.31 1.63 1.96
35 40 45
2.35 2.69 3.02
0.94 1.19 1.47
50 60 70
3.36 4.03 4.70
80 90 100
3”
Frict.
Tubing 2.95”ID
Pipe 3.07”ID
0.33 0.49 0.67
gpm 20 30 40
Vel. 0.94 1.41 1.88
Frict. 0.15 0.31 0.51
Vel. 0.87 1.30 1.74
Frict.
2.29 2.61 2.94
0.88 1.12 1.38
50 60 70
2.35 2.82 3.29
0.76 1.05 1.38
2.17 2.61 3.04
0.63 0.87 1.15
1.77 2.46 3.24
3.26 3.92 4.57
1.66 2.30 3.03
80 90 100
3.76 4.23 4.70
1.75 2.16 2.61
3.48 3.91 4.35
1.45 1.80 2.17
5.37 6.04 6.71
4.12 5.08 6.15
5.22 5.88 6.53
3.85 4.75 5.74
110 120 130
5.17 5.64 6.11
3.10 3.63 4.19
4.79 5.21 5.65
2.57 3.01 3.47
110 120 130
7.38 8.05 8.73
7.30 8.54 9.87
7.19 7.84 8.49
6.82 7.92 9.22
140 150 160
6.58 7.05 7.52
4.79 5.42 6.09
6.09 6.52 6.95
3.97 4.50 5.05
140 150 160
9.40 10.1 10.8
11.28 12.78 14.36
9.14 9.79 10.45
10.54 11.94 13.42
170 180 190
7.99 8.46 8.93
6.80 7.54 8.32
7.39 7.82 8.25
5.64 6.25 6.89
170 180 190
11.4 12.1 12.8
16.03 17.79 19.62
11.1 11.8 12.4
14.98 16.61 18.33
200 220 240
9.40 10.3 11.3
9.13 10.85 12.70
8.70 9.56 10.40
7.56 8.99 10.52
200 220 240
13.4 14.8 16.1
21.54 25.61 30.01
13.1 14.4 15.7
20.12 23.93 28.03
260 280 300
12.2 13.2 14.1
14.69 16.81 19.06
11.3 12.2 13.0
12.17 13.93 15.79
0.13 0.25 0.42
Appendix H
Friction Loss for Water in New Type L. Copper Tubing and Sch. 40 PVC Pipe (Frict. loss in ft. per 100 ft. - Vel. in ft. per sec.) 31/2” gpm
Tubing 3.43”ID Vel. Frict.
Pipe 3.55”ID Vel. Frict.
4” gpm
Tubing 3.91”ID Vel. Frict.
Pipe 4.63”ID Vel. Frict.
60 70 80
2.09 2.44 2.78
0.51 0.67 0.85
2.00 2.33 2.66
0.46 0.60 0.77
100 110 120
2.68 2.94 3.21
0.68 0.80 0.94
2.55 2.81 3.06
0.60 0.71 0.83
90 100 110
3.13 3.48 3.82
1.05 1.27 1.50
3.00 3.33 3.67
0.95 1.14 1.35
130 140 150
3.48 3.74 4.01
1.08 1.23 1.40
3.31 3.57 3.83
0.96 1.10 1.25
120 130 140
4.18 4.52 4.87
1.76 2.03 2.32
4.00 4.33 4.66
1.58 1.83 2.09
160 170 180
4.28 4.55 4.81
1.57 1.75 1.94
4.08 4.33 4.58
1.39 1.56 1.73
150 160 170
5.21 5.56 5.91
2.62 2.95 3.29
5.00 5.33 5.66
2.36 2.66 2.96
190 200 220
5.08 5.35 5.89
2.14 2.35 2.79
4.84 5.10 5.61
1.91 2.09 2.48
180 190 200
6.26 6.60 6.95
3.64 4.02 4.41
6.00 6.33 6.66
3.28 3.62 3.97
240 260 280
6.42 6.95 7.49
3.26 3.77 4.31
6.12 6.63 7.14
2.90 3.36 3.84
220
7.65
5.24
7.33
4.72
300
8.02
4.88
7.65
4.35
240 260
8.35 9.05
6.13 7.09
8.00 8.66
5.52 6.39
350 400
9.36 10.7
6.46 8.23
8.92 10.2
5.75 7.33
280 300 350
9.74 10.4 12.2
8.11 9.19 12.16
9.33 10.0 11.7
7.30 8.28 10.95
450 500 550
12.0 13.4 14.7
10.20 12.36 14.71
11.5 12.8 14.1
9.08 11.00 13.09
400 450 500
13.9 15.6 17.4
15.51 19.23 23.32
13.3 15.0 16.7
13.97 17.32 20.99
600 650 700
16.0 17.4 18.7
17.24 19.96 22.86
15.3 16.6 17.9
15.35 17.77 20.35
Note: 1. The friction losses listed under the pipe heading is approximately valid for Regular Weight Copper and Brass Pipe, in addition to Sch. 40 PVC Pipe 2. Table based on Darcy - Weisback formula 3. No allowance has been made for age, difference in diameter, or any abnormal condition of interior surface. Any factor of safety must be estimated from thelocal conditions and the requirements of each particular installation. It is recommended that for mostcommercial design purposes asafety factor of 15 to 20% be added to the values in the tables.
147
Appendix H
Friction Losses Through Pipe Valves and Fittings (Straight Pipe in Feet - Equivalent Length)
3/4 CLOSED
GLOBE VALVEWIDE OPEN
ANGLE VALVEWIDE OPEN
CHECK VALVEWIDE OPEN
5.0 7.0 9.0
19 26 36
9 12 16
5 6 8
2.0 3.0 4.0
.46 .60 .75
2.10 2.9 3.4
12.0 14.0 18.0
44 59 70
18 23 29
9 12 15
5.0 6.0 7.0
.79 .93 1.21
4.8 5.6 7.0
24.0 28.0 36.0
96 116 146
38 46 58
20 23 29
2 1/2” 3” 4”
1.39 1.69 2.40
8.4 10.0 14.0
41.0 52.0 70.0
172 213 285
69 86 116
6” 8” 10”
3.40 4.40 5.70
20.0 26.5 33.5
105 136 172
425 555 703
12” 14” 16”
6.80 8.20 9.10
40.6 48.5 53.0
196 233 274
815 978 1110
GATE VALVE 1/4 1/2 CLOSED CLOSED
SIZE OF PIPE (inches)
WIDE OPEN
1/8” 1/4” 3/8”
.14 .21 .27
.85 1.25 1.80
1/2” 3/4” 1”
.33 .46 .61
1 1/4” 1 1/2” 2”
ORDINARY ENTRANCE STD. TO PIPE 90° LINES ELBOW
MEDIUM SWEEP 90° ELBOW
LONG SWEEP 90° ELBOW
.74 1.0 1.4
.65 .86 1.15
.50 .70 .90
.90 1.4 1.6
1.6 2.3 2.7
1.50 2.0 2.5
1.10 1.5 2.0
9.0 11.0 15.0
2.5 3.0 3.5
3.6 4.5 5.4
3.5 4.0 5.0
2.5 2.9 3.6
35 43 57
17.0 21.0 27.0
4.0 5.0 6.5
6.5 8.5 12.0
6.0 7.0 9.5
4.4 5.5 7.2
175 225 285
86 115 141
39. 53. 65.
9.5 14. 16.
17. 22. 27.
15. 19. 23.
11.2 15.3 18.2
336 395 435
166 195 220
78. 92. 106.
18. 21. 26.
33. 37. 43.
27. 31. 36.
20.2 23.3 27.5
Use the smaller diameter in the column for pipe size. d Smaller diameter = D Largerdiameter ABRUPT CONTRACTION d d d D D D 1/4 1/2 3/4
ABRUPT ENLARGEMENT d d d D D D 1/4 1/2 3/4
SIZE OF PIPE (inches)
45° ELBOW
SQUARE 90° ELBOW
CLOSED RETURN BENDS
STD. TEE
STD. TEE
1/8” 1/4” 3/8”
.40 .50 .65
1.6 2.3 3.0
2.0 3.0 4.0
.50 .70 .90
1.6 2.3 3.0
.40 .50 .65
.30 .40 .50
.16 .22 .29
.74 1.0 1.4
.46 .62 .83
.16 .22 .29
1/2” 3/4” 1”
.80 1.0 1.5
4.0 5.0 6.0
5.0 6.0 7.0
1.10 1.5 2.0
4.0 5.0 6.0
.80 1.0 1.5
.60 .80 1.0
.36 .48 .62
1.6 2.3 2.7
1.2 1.4 1.6
.36 .48 .62
1 1/4” 1 1/2” 2”
1.7 2.0 2.5
8.0 9.5 13.0
9.0 11.0 14.0
2.5 2.9 3.6
8.0 9.5 13.0
1.7 2.0 2.5
1.4 1.6 2.0
.83 .97 1.30
3.6 4.5 5.4
2.3 2.7 3.5
.83 .97 1.30
2 1/2” 3” 4”
3.0 4.0 5.0
15.0 18.0 23.0
16.0 19.0 25.0
4.4 5.5 7.2
6” 8” 10” 12” 14” 16”
8.0 11.0 14.0 16.0 18.0 20.0
34.0 44.0 57.0 66.0 79.0 88.0
40.0 50.0 60.0 72.0 84.0 99.0
11.2 15.3 18.2 20.2 23.3 27.5
15.0 18.0 23.0 34.0 44.0 57.0 66.0 79.0 88.0
3.0 4.0 5.0 8.0 11.0 14.0 16.0 18.0 20.0
2.5 2.9 4.0 5.9 7.6 10.2 12.3 14.3 15.4
1.50 1.80 2.40 3.60 4.50 5.70 6.70 8.20 9.30
6.5 8.0 12.0 17.0 22.0 27.0 33.0 37.0 43.0
4.0 4.8 6.4 10.5 14.2 16.5 18.4 22.3 25.5
1.50 1.80 2.40 3.60 4.50 6.80 7.50 9.00 10.20
Note: 1. 1/8” to 12” nominal sizes are based on standard steel pipe, 14” to 24” sizes are ID pipe. 2. Friction losses are based on screwed connection from 1/8” to 4” sizes and flanged connections from 6” to 24”
148
Appendix H
Typical Check Valve Friction Loss Chart
Typical Surface Plate / 90° Discharge Elbow Friction Loss Chart SURFACE PLATE / 90°DISCHARGE FRICTION LOSS CHART
149
Appendix H
Steel Pipe Friction Loss & Velocity Chart
Note: Above chart indicates average values for standard weight steel pipe. Hazen - Williams roughness constant (C) = 140.
Equivalent Pipe Capacity Comparison Main Size 2” 3” 4” 6” 8” 10” 12” 14” 16” 18” 20”
Smaller Pipe Size (Number of smaller pipes required to provide carrying capacity equal to a larger pipe) 3/4” 1” 2” 3” 4” 6” 8” 13 39 84 247 530 957
6 18 39 115 247 447 724 1,090
1 2 6 18 39 71 115 174 247 338 447
1 2 6 13 24 39 59 84 115 153
1 2 6 11 18 27 39 53 71
1 2 3 6 9 13 18 24
1 1 2 4 6 8 11
10”
1 1 2 3 4 6
NOTE: Comparing the ratio of the square of diameters will provide the capacity equivalent relationship (ie. how many 12” lines will be required to equal the capacity of a 16” line? - (16 ) / (12 ) = 1.77 or 2 - 12” lines
150
Appendix I
Periodic system
1 H
2 He
Hydrogen
Helium
3 Li
4 Be
5 B
6 C
7 N
8 O
9 F
10 Ne
Lithium
Beryllium
Boron
Carbon
Nitrogen
Oxygen
Fluorine
Neon
11 Na
12 Mg
13 Al
14 Si
15 P
16 S
17 Cl
18 Ar
Sodium
Magnesium
Aluminium
Silicon
Phosphorus
Sulphur
Chlorine
Argon
19 K
20 Ca
21 Sc
22 Ti
23 V
24 Cr
25 Mn
26 Fe
27 Co
28 Ni
29 Cu
30 Zn
31 Ga
32 Ge
33 As
34 Se
35 Br
36 Kr
Potassium
Calcium
Scandium
Titanium
Vanadium
Chromium
Manganese
Iron
Cobalt
Nickel
Copper
Zinc
Gallium
Germanium
Arsenic
Selenium
Bromine
Krypton
42 Mo
43 Tc
37 Rb
Rubidium
38 Sr
39 Y
40 Zr
41 Nb
Strontium
Yttrium
Zirconium
Niobium
Molybdenum Technetium
44 Ru
45 Rh
46 Pd
47 Ag
48 Cd
49 In
50 Sn
51 Sb
52 Te
53 I
54 Xe
Ruthenium
Rhodium
Palladium
Silver
Cadmium
Indium
Tin
Antimony
Tellurium
Iodine
Xenon
55 Cs
56 Ba
57 La
72 Hf
73 Ta
74 W
75 Re
76 Os
77 Ir
78 Pt
79 Au
80 Hg
81 Tl
82 Pb
83 Bi
84 Po
85 At
86 Rn
Caesium
Barium
Lutetium
Hafnium
Tantalum
Tungsten
Rhenium
Osmium
Iridium
Platinum
Gold
Mercury
Thallium
Lead
Bismuth
Polonium
Astatine
Radon
109 Mt
110 Ds
111 Rg
112 Uub
113 Uut
114 UUq
115 U Up
116 U Uh
117 U Us
118 U Ud
87 Fr
88 Ra
89 Ac
104 Rf
105 Db
106 Sg
107 Bh
108 Hs
Francium
Radium
Actinium
Rutherfordium
Dubnium
Seaborgium
Bohrium
Hassium
59
60 Nd
61 Pm
58 Ce Cerium
Pr
Praseodymium Neodymium Promethium
Meitnerium Damstadtium Roentgenium Ununbium
Ununtrium Ununquadium
Sm
63 Eu
64 Gd
65 Tb
66 Dy
67 Ho
68 Er
69 Tm
70 Yb
71 Lu
Samarium
Europium
Gadolinium
Terbium
Dysprosium
Holmium
Erbium
Thulium
Ytterbium
Lutetium
62
90 Th
91 Pa
92 U
93 Np
94 Pu
95 Am
96 Cm
97 Bk
98 Cf
99 Es
100 Fm
101 Md
102 No
103 Lr
Thorium
Protactinium
Uranium
Neptunium
Plutonium
Americium
Curium
Berkelium
Californium
Einsteinium
Fernium
Mendelevium
Nobelium
Lawrencium
151
Appendix J
Pump standards Pump standards: ASME B73.1-2001 ASME B73.2-2003 EN 733 EN 22858
Specifications for horizontal end suction centrifugal pumps for chemical process Specifications for vertical in-line centrifugal pumps for chemical process End-suction centrifugal pumps, rating with 145.03 psi with bearing bracket End-suction centrifugal pumps (rating 232.06 psi) - Designation, nominal duty point and dimensions
Pump-related standards: ANSI/HI 1.6 ANSI/HI 1.3
ISO 3661 EN 12756 EN 1092
ISO 7005 DIN 24296
Centrifugal tests; detailed procedures on the setup and conduction of hydrostatic and performance tests Rotodynamic (centrifugal) pump applications; the standard cover the design and application of centrifugal pumps, pump classifications, impeller types, casing configurations, mechanical features, performance, selection criteria, and noise levels End-suction centrifugal pumps - Base plate and installation dimensions Mechanical seals - Principal dimensions, designation and material codes Flanges and their joints - Circular flanges for pipes, valves, fittings and accessories, PN-designated Metallicand flanges Pumps, pump units for liquids: Spare parts
Specifications, etc: ASME/ANSI B16.5-1996 ISO 9905 ISO 5199 ISO 9908 ISO 9906 EN 10204 ISO/FDIS 10816
Pipe flanges and flanged fittings Technical specifications for centrifugal pumps - Class 1 Technical specifications for centrifugal pumps - Class 2 Technical specifications for centrifugal pumps - Class 3 Rotodynamic pumps - Hydraulic performance tests -Grades 1 and 2 Metallic products - Types of inspection documents Mechanical vibration - Evaluation of machine vibration by measurements on non-rotating parts
Motor standards: Nema MG 1-2007 EN 60034/IEC 34
152
Information guide for general purpose industrial AC small and medium squirrel-cage induction motor standards Rotating electrical machines
Appendix K
Viscosity
Viscosity of typical liquids as a function of liquid temperature The graph shows the viscosity of typical liquids at different temperatures. As it appears from the graph, the viscosity decreases when the temperature increases.
Kinematic viscosity is measured in centiStokes [cSt] (1 cSt = 10-6 m2/s). The unit [SSU] Saybolt Universal is also used in connection with kinematic viscosity. The graph below shows the relationship between kinematic viscosity in [cSt] and viscosity in [SSU]. The SAE-number is also indicated in the graph. For kinematic viscosity above 60 cSt, the Saybolt Universal viscosity is calculated by the following formula: [SSU] = 4.62 . [cSt]
cSt 10000
The densities shown in the graph are for 68° F
8 V 6 4
Glycerol ρ: 1260
Kinematic viscosity centiStokes cSt
2
1000 8
3
6 4
Silicone oil
4 5
Fuel oil 2
8
20
Fruit juice ρ:
1000
Heavy 980
ρ:
4
Mean
Spindle oil ρ: 850
2
100
30 40 50
200
100
400 500
200
1000
300
ρ: 955
Gas and diesel oil ρ: 880
10
35 40
10
ρ: 900
6
32
50
Olive oil
Cottonseed oil ρ: 900
100
Sekunder Saybolt Universal SSU
1
2
SAE 10
Light 930
8
ρ:
6
Silicone oilρ: 1000
Milk ρ: 1030
4
300
Petroleum ρ: 800
Aniline ρ: 1030
2
ρ:
770
2000
1000
4000 5000
SAE 40
10000
SAE 60
8
Water ρ: 1000 4
Acetone ρ: 790
Acetic acid ρ: 1050
Ether ρ: 700 2
10
20
30
40
50
2000
SAE 50
3000
SAE 70
4000 5000
20000
10000
40000 50000
20000
100000
30000
Mercury ρ: 13570
t
0.1 0
SAE 30
Silicone oil
Petrol ρ: 750
6
SAE 20
400 500
3000 Ethyl Alkohol
1.0
-10
SAE no. ( at 68o F)
60
70
80
90
100°C
30000 40000 50000
200000
100000
153
Appendix K
Ethylene glycol Density of Aqueous Solutions of Ethylene Glycol Concentrations in Volume Percent Ethylene Glycol Temp., °F -3 0 -2 0 -1 0
10%
20%
30%
40%
-
50%
10 20 30 40
63.69 63.61
64.75 64.66
65.93 65.85 65.76 65.66
50
63.52
64.56
65.55
66.47
67.34
60 70 80 90 100 110
63.42 63.31 63.19 63.07 62.93 62.97
64.45 64.33 64.21 64.07 63.93 63.77
65.43 65.30 65.17 65.02 64.86 64.70
66.34 66.20 66.05 65.90 65.73 65.56
67.20 67.05 66.90 66.73 66.55 66.37
120
62.63
63.61
64.52
65.37
66.17
130 140 150 160 170
62.47 62.30 62.11 61.92 61.72
63.43 63.25 63.06 62.86 62.64
64.34 64.15 63.95 63.73 63.51
65.18 64.98 64.76 64.54 64.31
65.97 65.75 65.53 65.30 65.05
180
61.51
62.42
63.28
64.07
64.80
190 200
61.29 61.06
62.19 61.95
63.04 62.79
63.82 63.56
64.54 64.27
210 220
60.82 60.57
61.71 61.45
62.53 62.27
63.29 63.01
63.99 63.70
230 240
60.31 60.05
61.18 60.90
61.99 61.70
62.72 62.43
63.40 63.10
250
59.77
60.62
61.40
62.12
62.78
-
-
68.12
0
-
-
64.83
-
68.05 64.98
67.04 66.97 66.89
66.80 66.70 66.59
67.90 67.70 67.59 67.47
67.80
Note: Density in lb/ft3.
Viscosity of Aqueous Solutions of E thylene Glycol Concentrations in Volume Percent Ethylene Glycol Temp., °F -3 0 -2 0 -1 0 0 10 20
154
10%
20% -
30% 0.0026
40% 0. 004 6 0.0036
50% 0 . 01 32 0 . 00 92 0 . 00 68 0.0052
0.0428 0.0271 0. 0 183 0. 0 130 0. 0 096 0.0073
30 40 50 60 70 80 90 100
0.0015 0.0012 0.0010 0.0009 0.0008 0.0007 0.0006 0.0006
0.0021 0.0017 0.0015 0.0012 0.0011 0.0009 0.0008 0.0007
0.0029 0.0024 0.0020 0.0017 0.0014 0.0012 0.0011 0.0009
0.0041 0.0033 0.0027 0.0023 0.0019 0.0017 0.0014 0.0013
0.0057 0.0045 0.0037 0.0031 0.0026 0.0022 0.0019 0.0016
110 120 130 140 150 160 170
0.0005 0.0005 0.0004 0.0004 0.0004 0.0003 0.0003
0.0007 0.0006 0.0005 0.0005 0.0005 0.0004 0.0004
0.0008 0.0007 0.0007 0.0006 0.0006 0.0005 0.0005
0.0011 0.0010 0.0009 0.0008 0.0007 0.0006 0.0006
0.0014 0.0012 0.0011 0.0010 0.0009 0.0008 0.0007
180 190 200 210 220 230 240
0.0003 0.0003 0.0002 0.0002 0.0002 0.0002 0.0002
0.0004 0.0003 0.0003 0.0003 0.0003 0.0003 0.0002
0.0004 0.0004 0.0004 0.0003 0.0003 0.0003 0.0003
0.0005 0.0005 0.0005 0.0004 0.0004 0.0004 0.0003
0.0006 0.0006 0.0005 0.0005 0.0004 0.0004 0.0004
250
0.0002
0.0002
0.0003
0.0003
0.0003
Appendix K
Propylene glycol Density of Aqueous Solutions of Propylene Glycol Concentrations in Volume Percent Propylene Glycol Temp., °F - 30 - 20 - 10
10%
20%
30% -
40%
50%
0
-
10 20 30 40
63.38 63.30
64.23 64.14 64.03
50 60 70 80 90 100 110 120 130 140 150
63.20 63.10 62.98 62.86 62.73 62.59 62.44 62.28 62.11 61.93 61.74
63.92 63.79 63.66 63.52 63.37 63.20 63.03 62.85 62.66 62.46 62.25
64.53 64.39 64.24 64.08 63.91 63.73 63.54 63.33 63.12 62.90 62.67
65.06 64.90 64.73 64.55 64.36 64.16 63.95 63.74 63.51 63.27 63.02
65.50 65.33 65.14 64.95 64.74 64.53 64.30 64.06 63.82 63.57 63.30
160 170 180 190
61.54 61.33 61.11 60.89
62.03 61.80 61.56 61.31
62.43 62.18 61.92 61.65
62.76 62.49 62.22 61.93
63.03 62.74 62.45 62.14
200 210
60.65 60.41
61.05 60.78
61.37 61.08
61.63 61.32
61.83 61.50
220 230
60.15 59.89
60.50 60.21
60.78 60.47
61.00 60.68
61.17 60.83
240
59.61
59.91
60.15
60.34
60.47
250
59.33
59.60
59.82
59.99
60.11
-
-
-
65.00 64.90 64.79 64.67
-
66.46 66.35 66.23
65.71 65.48 65.35 65.21
65.60
65.97 65.82 65.67
66.11
Note: Density in lb/ft 3.
Viscosity of Aqueous Solutions of Propylene Glycol Concentrations in Volume Percent Propylene Glycol Temp., °F - 30 -2 0 -1 0
10%
20% -
30% -
40% -
50% -
0.1049 0.0645
0 10 20 30 40 50 60 70 80 90 100 110 120 130 140 150 160 170
0.0019 0.0015 0.0013 0.0011 0.0009 0.0008 0.0007 0.0006 0.0006 0.0005 0.0005 0.0004 0.0004 0.0003 0.0003
0.0036 0.0028 0.0023 0.0019 0.0016 0.0013 0.0011 0.0010 0.0008 0.0007 0.0007 0.0006 0.0005 0.0005 0.0004 0.0004
0.0090 0.0067 0.0050 0.0039 0.0030 0.0024 0.0020 0.0016 0.0014 0.0012 0.0010 0.0009 0.0008 0.0007 0.0006 0.0006 0.0005
0.0275 0.0183 0.0124 0.0089 0.0065 0.0049 0.0037 0.0029 0.0024 0.0019 0.0016 0.0013 0.0011 0.0010 0.0009 0.0008 0.0007 0.0006
0.0412 0.0273 0.0187 0.0132 0.0096 0.0072 0.0055 0.0043 0.0034 0.0027 0.0023 0.0019 0.0016 0.0014 0.0012 0.0010 0.0009 0.0008
180 190 200 210 220 230 240 250
0.0003 0.0003 0.0003 0.0002 0.0002 0.0002 0.0002 0.0002
0.0004 0.0003 0.0003 0.0003 0.0003 0.0003 0.0002 0.0002
0.0005 0.0004 0.0004 0.0004 0.0003 0.0003 0.0003 0.0003
0.0006 0.0005 0.0005 0.0005 0.0004 0.0004 0.0004 0.0003
0.0007 0.0007 0.0006 0.0005 0.0005 0.0005 0.0004 0.0004
155
Appendix K
Sodium hydroxide ρ
ν
[lb/ft3] Concentration wt % = Temperature
[cSt]
% 5
ρ
[lb/ft3]
ν
[cSt]
10%
ρ
[lb/ft3]
ν
ρ
ν
ρ
ν
ρ
ν
ρ
ν
ρ
ν
ρ
ν
ρ
ν
ρ
ν
[cSt]
[lb/ft3]
[cSt]
[lb/ft3]
[cSt]
[lb/ft3]
[cSt]
[lb/ft3]
[cSt]
[lb/ft3]
[cSt]
[lb/ft3]
[cSt]
[lb/ft3]
[cSt]
[lb/ft3]
[cSt]
15 %
2 0%
25 %
3 0%
35%
40%
45%
5 0%
55%
32
66.17
69.73
73.29
76.78
80.21
83.27
86.40
89.58
92.58
95.51
41
66.04
69.60
73.16
76.59
80.09
83.15
86.21
89.20
92.39
95.38
97.13
50
65.98
69.48
73.04
76.41
79.90
83.02
85.96
88.83
92.26
95.20
96.95
59
65.92
68
65.79
1.3
69.23
1.7
72.66
2.5
76.09
3.6
79.53
77
65.67
1.1
69.10
1.5
72.54
2.1
75.97
3.1
79.34
5.1
82.34
72.85
69.35
79.72
76.28
82.77 6.2
82.52
85.65
91.83
88.64
97.32
94.76
96.51
85.33
16.8
8.39 8
25.4
91.39
38.2
94.32
51.8
8.3
85.15
13.3
88 .21
19.9
91.20
29.0
94.14
39.0
10.1
86
65.54
1.0
68.92
1.3
72.35
1.8
75.78
2.7
79.15
4.0
82.09
6.5
84.90
9.9
8.02 8
14.4
90.95
19.9
93.89
26.2
95
65.42
0.9
68.79
1.2
72.22
1.6
75.60
2.3
78.97
3.4
81.90
5.5
84.71
8.2
7.83 8
11.6
90.77
15.9
93.70
20.5
104
65.29
0.8
68.67
1.1
72.04
1.4
75.41
2.0
78.78
2.8
81.71
4.5
84.46
6.6
113
65.17
0.7
68.48
1.0
71.85
1.3
75.22
1.8
78.59
2.6
81.53
3.9
84.09
5.6
122
65.04
0.7
68.29
0.9
71.66
1.2
75.03
1.6
78.40
2.3
81.28
3.3
83.65
4.6
131
64.86
0.6
68.17
0.8
71.48
1.0
74.85
1.5
78.22
2.0
81.09
2.9
140
64.67
0.6
67.98
0.7
71.35
0.9
74.66
1.3
78.03
1.8
80.84
2.4
149
64.48
0.5
67.79
0.7
71.16
0.9
74.47
1.2
77.78
1.6
158
64.30
0.5
67.60
0.6
70.98
0.8
74.28
1.1
77.59
1.5
167
64.11
67.42
70.79
74.03
77.41
176
63.98
67.23
70.60
73.85
77.22
77
86
lb/ft3
12.0
93.45
14.7
9.9
93.01
12.1
6.0
89.64
7.8
92.58
9.4
104
113
50% 45% 40%
45%
35%
40% 10
30%
35%
25%
30% 81.15
90.52 90.08
100 55% 50%
87.39
6 8.71
8.9 7.5
cSt
99.88
93.64
7.58 8 87.14
96.13
20% 15% 10% 5%
25% 20%
1
74.91 15%
68.67
10% 5%
0 68
62.42 32
156
95
122
131
140
149
158 °F
50
68
86
104
122
140
158
176
°F
Appendix K
Calcium chloride
Sodium chloride
ρ
ν
ρ
ν
ρ
ν
ρ
ν
ρ
ν
ρ
ν
ρ
ν
ρ
ν
[lb/ft3]
[cSt]
[lb/ft3]
[cSt]
[lb/ft3]
[cSt]
[lb/ft3]
[cSt]
[lb/ft3]
[cSt]
[lb/ft3]
[cSt]
[lb/ft3]
[cSt]
[lb/ft3]
[cSt]
Concentration wt % = Temperature
10%
15%
20%
2 5%
Concentration wt % = Temperature
-13
77.72
7.7
5
-4
77.66
6.3
14
5.2
23
5 14
71.04
74.22
4.3
77.53
3.0
74.16
3.6
77.47
4.4
5%
10 %
15 %
20%
72.54
4.0
69.91
2.9
72.41
3.2
67.54
2.2
69.79
2.4
72.29
2.7 2.3
32
65.11
1.8
67.42
1.8
69.66
2.0
72.10
23
68.04
2.3
70.98
2.6
74.10
3.1
77.34
3.8
41
65.04
1.5
67.35
1.6
69.54
1.7
71.97
1.9
32
67.92
2.0
70.85
2.2
74.03
2.6
77.22
3.3
50
64.98
1.3
67.23
1.4
69.41
1.5
71.85
1.7
41
67.79
1.7
7 0.79
1.9
73.91
2.3
77.09
2.9
59
64.92
1.1
67.11
1.2
69.29
1.3
71.66
1.5
50
67.73
1.5
70.66
1.7
73.78
2.0
7 6.97
2.5
68
64.86
1.0
67.04
1.1
69.17
1.2
71.54
1.3
59
67.60
1.3
70.60
1.5
73.66
1.8
76.78
2.2
77
64.73
0.9
66.92
0.9
69.04
1.0
71.41
1.2
68
67.54
1.1
70.48
1.3
73.54
1.6
76.66
2.0
86
64.67
0.8
66.79
0.9
68.85
0.9
71.23
1.1
77
67.54
1.0
70.35
1.2
73.41
1.4
76.53
1.8
86
67.48
0.9
70.23
1.0
73.22
1.3
7 6.34
1.6
157
Index
A Absolute pressure Adjusting pump performance Aluminum ATEX (ATmosphère EXplosible) Austenitic (non-magnetic) Autotransformer starting
85 106 70 41 68 46
Axial flow pumps Axial forces
8 14
B Balanced shaft seal Basic coupling Bearing Insulated bearing Bellows seal Groundwater pump Bypass control
31 16 51 48 30 23 106
Constant differential pressure control 115 Constant pressure control 114 Constant temperature control 115 Copper alloys 69 Corrosion 60 Cavitation corrosion 63 Corrosion fatigue 64 Crevice corrosion Erosion corrosion Galvanic corrosion Intergranular corrosion Pitting corrosion Selective corrosion Stress corrosion cracking (SCC) Uniform corrosion Corrosion fatigue Coupling Basic coupling Flexible coupling Spacer coupling Crevice corrosion
62 63 64 62 61 62 63 61 64 16 16 16 16 62
C Canned motor pump 18 Cartridge seal 32 Casing 15 Double-volute 15 Single-volute 15 Return channel 15 Cast iron 66 Cavitation 10, 89 Cavitation corrosion 63 Centrifugal pump 8 Ceramics 71 Close-coupled pump 12, 13, 16 Closed system 96, 98 Coatings 73 Metallic coatings 73 Non-metallic coatings 74 Organic coatings 74 Computer-aided pump selection 58 Control 106 Throttle control 107 Bypass control 107 Speed control 108
D Decommissioning and disposal costs Deep well pump Density Density of water Density of brine Diaphragm pump Differential pressure Differential pressure control Dilatant liquid Direct-on-line starting (DOL) Dosing pump Double mechanical shaft seal Double seal in tandem Double seal in back-to-back Double-channel impeller Double-inlet Double-suction impeller Double-volute casing Downtime costs
131 23 10, 93 Appendix D Appendix K 25 88 116 55 46 25 33 33 34 21 17 11, 17 15 131
Index
Duty point Dynamic pressure Dynamic viscosity
Index
96 84 54
Frequency converter
47, 108, 118
G E Earth-leakage circuit breaker (ELCB)
125
Galvanic corrosion Gauge pressure Grey iron
Efficiency 10 Efficiency at reduced speed 109 Efficiency curve 10 H Electric motor 40 Head Flameproof motor 42 Heat capacity Increased safety motor 42 Hermetically sealed pump Non-sparking motor 42 Horizontal pump EMC directive 123 Hydraulic power EMC filter 123 Enclosure class (IP), motor 43 End-suction pump 12 I Energy costs 130 IEC, motor Energy savings 111, 114, 117 Immersible pump Environmental costs 130 Impeller Erosion corrosion 63 Double-channel Ethylene propylelediene rubber (EPDM) 72 Single-channel Expansion joints
80
F Ferritic (magnetic) Ferritic-austenitic or duplex (magnetic) Ferrous alloys Flameproof motor Flexible coupling Floating foundation Flow Mass flow 83 Volume flow 83 Units Appendix B Fluoroelastomers (FKM) Flushing Foundation Floating foundation 79 Floor 79 foundation 79 Vibration dampeners 79 Frame size
68 68 65 42 16 79 83
72 32 78
44
64 85 66
9, 85 93 18 12, 13 10, 91
40 22 14, 21 21 21
Vortex impeller 21 Increased safety motor Initial costs In-line pump Installation and commissioning costs Insulation class Intergranular corrosion
42 129 12, 13 129 44 62
K Kinematic viscosity
54, Appendix K
L
N
Life cycle costs Example
117, 128 132
Liquid Dilatant Newtonian Non-Newtonian Plastic fluid Thixotrophic Viscous Long-coupled pump Loss of production costs
54 55 55 55 55
55
54 12, 13, 16 131
Non-metallic coatings Non-Newtonian liquid Non-sinusoidal current Non-sparking motor NPSH (Net Positive Suction Head)
40 55 69 72 66 78 74 55 124 42 10, 89
O
M Magnetic drive Maintenance and repair costs Martensitic (magnetic) Mass flow Measuring pressure Mechanical shaft seal Bellows seal Cartridge seal Metal bellows seal Rubber bellows seal Function 29 Flushing 32 Metal alloys Ferrous alloys Metal bellows seal Metallic coatings Mixed flow pumps Modifying impeller diameter Motors Motor efficiency Motor insulation Motor protection Motor start-up Direct-on-line starting (DOL) Star/delta starting Autotransformer starting Frequency converter Soft starter Mounting of motor (IM) Multistage pump
NEMA, motor standard Newtonian fluid Nickel alloys Nitrile rubber Nodular iron Noise (vibration)
19 131 68 83 85 18, 28 30 32
Open system Operating costs Organic coatings O-ring seal Oversized pumps
96, 99 106, 130 74 30 106
P 32 31
65 65 32 73 8 108, 110 40 49 48 49 46 46 46 46 46, 47 46 43 11, 12, 13, 16
Paints 74 Perfluoroelastomers (FFKM) 72 Phase insulation 48 PI-controller 114 Pitting corrosion 61 Plastic fluid 55 Plastics 71 Positive displacement pump 24 Power consumption 10, 91 Hydraulic power 10, 91 Shaft power 91 Pressure 84 Absolute pressure 85 Differential pressure 88 Dynamic pressure 84 Gauge pressure 85 Measuring pressure 85 Static pressure 84 System pressure 88 Units 85, Appendix A Vapor pressure 90, Appendix D
Index
Index
Pressure control Constant differential pressure control 115 Constant pressure 114 Constant pressure control 114 Constant supply pressure 114 Pressure transmitter (PT) 114 Proportional pressure control 120 PTC thermistors Pulse Width Modulation (PWM) Pump Axial flow pump Borehole pump Canned motor pump Centrifugal pump Close-coupled pump Diaphragm pump Dosing pump Hermetically sealed pump Horizontal pump Immersible pump Long-coupled pump Magnetic-driven pump Mixed flow pump Multistage pump Positive displacement pump Radial flow pump Sanitary pump Single-stage pump Split-case pump Standard pump Vertical pump Wastewater pump Pump casing Pump characteristic Pump curve Pump installation Pump performance curve Pumps connected in series Pumps in parallel
50 123 8 23 18 8 12, 13, 16 25 25 18 12, 13 22 12, 13, 16 19 8
QH curve
9
R Radial flow pump Radial forces
8 15
Reinforced insulation Resistances connected in parallel Resistances connected in series Return channel casing Rubber Ethylene propylelediene rubber (EPDM) Fluoroelastomers (FKM) 72 Nitrile rubber (NBK) 72 Perfluoroelastomers (FFKM) 72 Silicone rubber (Q) 72 Rubber bellows seal 30
48 98 97 15 72 72
S
11, 12, 13, 16 24 8 20 15 12, 13, 17 17 12, 13 21
Pumps with integrated frequency converter Purchase costs PWM (Pulse Width Modulation)
Q
15 9, 96 9 77 9, 96 103 101 118 129 123
Sanitary pump Seal face Seal gap Selective corrosion Setpoint Shaft Shaft power Shaft seal Balanced shaft seal Unbalanced shaft seal Silicone rubber (Q) Single resistances Resistances connected in series Single-channel impeller Single-stage pump Single-suction impeller Single-volute casing Soft starter Sound level Sound pressure level Spacer coupling Static head Static lift
20 28 29 62 114 11 91 28 31 31 72 97 97 21 11, 12, 13, 15 11 15 46 81 82 16 99 99
Speed control 106, 108, 110 Variable speed control 108 Speed-controlled pumps in parallel 102 Split-case pump 12, 13, 17 Stainless steel 66 Standard pump 17 Standards 40 IEC, motor
NEMA, motor Sanitary standards Standstill heating of motor Star/delta starting Static pressure Steel Stress corrosion cracking (SCC) Stuffing box Submersible pump System characteristic Closed system Open system System costs System pressure
40
40 20 51 46 84 65 63 28 23 96
96, 98 96, 99 117 88
T Temperature Units
93 Appendix B
Thermoplastics Thermosets Thixotrophic liquid Throttle control Throttle valve Titanium Twin pump
71 71 55 106, 110-113 107 70 11
U Unbalanced shaft seal Uniform corrosion
31 61
V Vapor pressure 90, Appendix D Variable speed control 108 Vertical pump 12, 13 Vibration dampeners 79 Vibrations 78 Viscosity 54, Appendix K Dynamic viscosity 54 Viscous liquid 54 Viscous liquid pump curve 55 Voltage supply 47 Volume flow 83 Units Appendix A Volute casing 11 Vortex impeller 21 Wastewater pump 21
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U.S.A.
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GRUNDFOS Pumps Corporation 17100 West 118th Terrace Olathe, Kansas 66061 Phone: (913) 227-3400 Telefax: (913) 227-3500
GRUNDFOS Canada Inc. 2941 Brighton Road Oakville, Ontario L6H 6C9 Phone: (905) 829-9533 Telefax: (905) 829-9512
Bombas GRUNDFOS de Mexico S.A. de C.V. Boulevard TLC No. 15 Parque Industrial Stiva Aeropuerto C.P. 66600 Apodaca, N.L. Mexico Phone: 011-52-81-8144 4000 Telefax: 011-52-81-8144 4010
www.grundfos.com
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