Boiler Feed Water Pumps

GBH Enterprises, Ltd.

Engineering Design Guide: GBHE-MAC-1504

Boiler Feedwater Pumps

Information contained in this publication or as otherwise supplied to Users is believed to be accurate and correct at time of going to press, and is given in good faith, but it is for the User to satisfy itself of the suitability of the information for its own particular purpose. GBHE gives no warranty as to the fitness of this information for any particular purpose and any implied warranty or condition (statutory or otherwise) is excluded except to the extent that exclusion is prevented by law. GBHE accepts no liability resulting from reliance on this information. Freedom under Patent, Copyright and Designs cannot be assumed.

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Engineering Design Guide:

Boiler Feedwater Pumps

CONTENTS

SECTION

0

INTRODUCTION/PURPOSE

1

SCOPE

5

SECTION ONE - INTEGRATING.THE PUMP INTO THE SYSTEM

2

AVAILABILITY CONSIDERATIONS

3

CHOICE OF NUMBER OF PUMPS

4

CHOICE OF PUMP TYPE 4.1 4.2 4.3 4.4

5

DRIVERS 5.1 5.2 5.3

6

Barske Type Peripheral Type Multistage Centrifugal Type Inlet Booster Pumps

Steam Turbines Electric Motors Dual Drivers

DUTY 6.1 6.2 6.3

Differential Head Capacity NPSH


7

MINIMUM FLOW THROUGH PUMP 7.1 7.2 7.3

8

Limit due to Instability in Head/Flow Characteristic Minimum Flow Arrangements Variable Speed Control

TRANSIENT EFFECTS 8.1 8.2

Reverse Flow upon Trip Flashing

9

TEMPERATURE GRADIENTS IN CASINGS

10

INLET STRAINERS AND CASING PIPING CONNECTIONS 10.1 10.2

11

Strainers Casing Piping Arrangements

SEAL COOLING 11.1 11.2

Packed Glands for Category 1 Pumps Mechanical Seals

SECTION TWO - PUMP CONSTRUCTION FEATURES

12

MINIMUM FLOW THROUGH PUMP 12.1 12.2

13

Limit Due to First Stage Cavitation Damage Limit Due to Instability in Head/Flow Characteristic

IMPELLER/DIFFUSER STAGE 13 .1 Local Cavitation 13.2 Labyrinth or Bushing Clearances 13.3 Clearances Affecting Hydraulic Forces


14

ROTOR 14. 1 14.2 14.3 14.4

15

GLAND ARRANGEMENTS 15.1 15.2 15.3 15.4

16

Rotor Dynamics Rotor Mechanical Balance Torsional Critical Speeds Rotor Assembly

Guarding Soft-Packed Glands Mechanical Seals for Category 1 Pumps Mechanical Seals for Category 2 Pumps

AXIAL HYDRAULIC THRUST BALANCE METHODS 16. 1 Opposed Impeller Configuration 16.2 Balance Disc 16.3 Balance Piston

17 17.1 17.2

CASING Casing Type Casing Connections

18

MATERIALS

19

DRIVER CONSIDERATIONS


BIBLIOGRAPHY APPENDICES: A B C D E F G H

NOTES ON BFW PUMP/DRIVER ARRANGEMENTS PROPERTIES OF WATER DEFINITION OF PUMP CATEGORY TRANSIENT PHENOMENA IN DEAERATORS NOTES ON PROPRIETARY LEAR-OFF VALVES NOTES ON CASTINGS FOR HIGH-DUTY IMPELLERS AREA RATIO METHOD TECHNICAL COMPARISON SHEETS

FIGURES 1

TYPICAL CROSS-SECTION OF PUMP IN CATEGORY 1

2A

TYPICAL CROSS-SECTION OF PUMP IN CATEGORY 2 SHOWING AXIALLY SPLIT CASING

2B

TYPICAL CROSS-SECTION OF PUMP IN CATEGORY 2 SHOWING BARREL CASING WITH RING TYPE CARTRIDGE

3

TYPICAL CROSS-SECTION OF PUMP IN CATEGORY 3 FOR POWER STATION APPLICATIONS WITH DRY RUNNING CAPABILITIES

DOCUMENTS REFERRED TO IN THIS ENGINEERING DESIGN GUIDE


1

SCOPE

This Engineering Design Guide covers the special requirements of pumps in Boiler Feedwater service. SECTION ONE - INTEGRATING THE PUMP INTO THE SYSTEM

2

AVAILABILITY CONSIDERATIONS

The general reliability classifications are given in Engineering Design Guide GBHE-MAC-5101. Pump reliability is more difficult to obtain as the pressure of the system increases. Current practice is to distinguish three categories of the system by reference to the nominal pressure of the steam boiler drum. Category Steam pressure, bar abs.

1 43

2 43 – 125

3 160

Category 3 covers a model thermal power station and is beyond the scope of this Design Guide. Consult Furnace Section about the consequences of BFW supply failure. For Category 2, BFW pumps are normally required to have 100% availability in order to avoid boiler 'dry-out'. Then: (a)

A standby pump on autostart for immediate readiness is required.

(b)

As a precaution against failure of the electricity supply at least one pump requires a steam turbine driver with the steam supply taken either from the boiler or from a secure source. See Appendix A.

Some boilers, particularly those within Category 1, are able to cope with BFW supply failure and consequently permit the use of electric motors for both main and standby pumps.


3

CHOICE OF NUMBER OF PUMPS

Arrangement 1 One 100% duty pump together with one standby pump. Current practice is to make the standby pump identical to the main pump on grounds of: (a)

Common spares holding

(b)

Common maintenance and operational methods

(e)

No process interruption upon short-term trip of main pump

(d)

Identical piping arrangement, minimizing design" effort

Arrangement 2 Two running 50% duty pumps with one identical standby pump. This arrangement may be selected in order to: (a)

Reduce the NPSH requirement to avoid the need for inlet booster pumps

(b)

Ensure continuity of BFW supplied for process use. This condition is important when delivery lines are long or heat exchangers are included which can quickly generate steam upon BFW flow stoppage.

(c)

Increase overall power efficiency when the steaming rate is expected to be less than 50% of plant rating for long periods, thus allowing operation with one pump.

(d)

Relax constraints on steam system control.


Special Arrangements The following arrangements should NOT be used without thorough systems analysis: (a)

Two running pumps each rated at 100% duty in order to guarantee BFW supply continuity at 100% of rated flow.

(b)

'Minimum duty' standby pump sets, either rated at a low pressure and so requiring rapid boiler depressuring before they can be started, or rated at a fraction of the main pump capacity.

4

CHOICE OF PUMP TYPE

4.1

Barske Type

The high-speed one or two stage Barske type pump requires a high NPSH and may have an unstable Q-H characteristic. Nevertheless, where the steam demand varies widely, with campaigns of low flow operation, consider the use of 3 or more such pumps in parallel so that the number of running pumps can be adjusted to meet the BFW demand.

4.2

Peripheral Type

The peripheral type pump is well suited to BFW service for small capacities because its Q-H characteristic is inherently stable and because BFW is sufficiently pure and free from suspended solids to make acceptable this pump's sensitivity to erosion. Specify this type of pump in preference to reciprocating pumps.

4.3

Multistage Centrifugal Type

Most BFW pumps are of the horizontal shaft multistage centrifugal type. The choice of the number of stages is not a precise determination. Some guidance is given in Clause Cl.4 of GBHE-EDG-MAC-1014.


For small pumps where Q.H1/2 < 200 or where the efficiency is not important:

As the number of stages increases the bearing span increases and the rotor becomes more sensitive to dynamic effects. As a first estimate take the limiting number of stages as:

Where

k=

34

for stiff rotors, capable of 'dry' running.

and

k=

186 for conventional rotors

A double-entry first stage is the equivalent of 2 single-entry stages for this calculation. Where the number of stages thus calculated is less than the number of stages required to obtain a reasonable efficiency, then consider a vertical shaft pump.

4.4

Inlet Booster Pumps

These may be needed when the NPSH available is insufficient to meet the NPSH required by the main pump. The conventional arrangement has the booster directly coupled to a double ended electric motor which also drives the geared main pump.


5

DRIVERS

5.1

Steam Turbines

The advantages of using different designs of main and standby turbines are sufficiently large to outweigh the reasons given in Clause 2 for selecting identical main and standby pumps. In normal operation the failure rate of small high efficiency turbines is significantly higher than that for the conventional small turbine which has only 1 or 2 rows of impulse blading with large non-critical clearances. Consequently the preferred driver arrangement is to have main pumps driven by electric motors or high efficiency turbines and the standby pump driven by a conventional turbine provided with slowroll and quickstart facilities. Standby turbines of normal construction are held on slowroll to ensure instant readiness. For this purpose current practice is to leave the emergency steam stop valve open and the inlet autostart valve closed but bypassed by a small restrictor whose orifice size is empirically adjusted to give the required slowroll speed. The autostart valve actuator is damped to give a valve stroking time of 10 seconds (for a linear valve characteristic) in order to avoid speed overshoot. Check that the slowroll speed is above the hydrodynamic limit for journal bearings. Small standby turbines having single wheels with integrally cast blades, exemplified by t he Terry turbine, need not slowroll. A standby pump with steam turbine driver imposes a sudden high demand on the steam supply system upon starting. Such a steam demand can be supplied direct from the boiler. However, in process plant these turbines are normally supplied from an intermediate pressure header. Conventional control systems cope better with reductions than with increases in steam demand; for sensitive systems the preferred driver for the standby pump is an electric motor. Then, current practice is to specify Arrangement 2 with each main pump having a turbine driver. For the emergency condition of electrical supply failure; specify the turbine driver to be capable of developing rated power when exhausting either to a secure steam system or to atmosphere via a pressure relief valve.


5.2

Electric Motors

Current practice is to use only induction type motors arranged for direct-on-line start. Consult Electrical Section on the advisability of connecting motors to different unit supply boards in order to reduce upsets due to supply faults. It is expensive to obtain slowroll operation on an electric motor driven pumpset. For standby duty, first consider the set as stationary and review the consequent effects on both pump design and layout. Standby pumps are preferably directdrive. Gears that remain stationary need special consideration; the oil console should then be provided with an electric motor driven auxiliary oil pump continuously running to maintain an oil film over the gear teeth.

5.3

Dual Drivers

One pump may be coupled to two drivers, each capable of driving the whole set. Such an arrangement gives a cost saving for large pumps when the reliability of the normal driver is much lower than that of the pump. This case arises when the normal driver is an electric motor connected to an unreliable electric supply system and the alternative driver is a steam turbine. Now the failure rate of Category 2 pumps is of the order of 0.5/year, consequently this arrangement is not justifiable for most sites with access to the CEGB grid, when the supply failure rate is of the order of 0.1/year. The dual drive arrangement may be used for power recovery, where the steam turbine is in continuous operation and the motor can act as an induction generator exporting surplus energy as electrical power.

6

DUTY

This is calculated using Engineering Design Guide GBHE-EDG-MAC-1014 with amendments given in the following notes.


6.1

Differential Head

When calculating the required differential head at rated flow, with throttle control valves WIDE-OPEN, the delivery vessel pressure is taken as the set pressure of the boiler drum relief valves. This step does not guarantee that rated water flow can be then delivered because relief valves require a further increase in pressure when passing their rated steam flow. The steep pump characteristic can accommodate this pressure increment (of 5 - 10%) at some marginally lower flow. Pumps may require a correction to the simple head calculation; refer to Appendix B.

6.2

Capacity

Allow for the minimum flow requirement as given in Clause 7.2. Early Category I installations had duties incorporating an allowance for damaged boiler tubes but this practice has been discontinued.

6.3

NPSH

When calculating available ·NPSH note that some boilers have only single element level control: this demands either a generous NPSH allowance for acceleration head or a heavily damped throttle control valve, to cope with transient boiler upsets. BFW pumps may require unusually large values of NPSH. As speed or capacity increase, keeping S n constant maintains hydraulic performance but the intensity of local cavitation increases. There is a limiting value of NPSH for a given material and impeller construction. For typical castings in 13% Cr steels take this limit as defined by:

S n (NPSH p)1/4 < 1.2 Where the values are taken at the pump best efficiency point and apply over the operating range 85 - 105% of the capacity at BEP; S n should be further reduced if a wider range of operational capacity is required.


Sudden failure of the steam supply to the de-aerator gives a transient reduction in the available NPSH. If the residence time for the de-aerator is less than 600 seconds, use the method given in Appendix D to calculate the correction; otherwise apply a correction of 0.1 m.

7

MINIMUM FLOW THROUGH PUMP

7.1

Limit Due to Instability in Head/Flow Characteristic

Typical pumps have Ns ~ 0.06 for which the minimum flow to avoid instability is 20% of the pump capacity at BEP. As Ns increases, this minimum flow limit also increases but no generalized quantitative guidance can be given. Small high-speed Barske pumps require a minimum flow of the order of 60% BEP flow.

7.2

Minimum Flow Arrangements

Proprietary leak-off valves which combine the functions of a non-return valve and a bypass valve are currently limited to Category 1 installations. Current practice to ensure the minimum flow limit is to employ a simple continuous bypass through a let-down MULTIPLE restrictor back to the inlet vessel. This continuous bypass bas been used for process purposes. The capacity of the pump is then increased to supply both the bypass and the delivery flows. Large category 2 pumps may merit a dedicated control system to maintain minimum flow by opening the bypass valve when the delivery flow falls below the specified minimum flow. Such a system should be of high integrity. 7.3

Variable Speed Control

Such experience as is available is discouraging. The chief considerations are: (a)

Commercial speed governors for small steam turbines have deadband zones of about 0.25% when specified to NEMA Class 'D'. For typical pump and system characteristics, this limits the stable range of control on a single pump to about 70% of BEP capacity; below this capacity ordinary throttle valve control is used with the speed held constant.


Arrangements where two pumps run in parallel exacerbate the stability problem and the only successful method used to date is to run the turbines at constant speed and rely on throttle valve regulation.

(b)

Fluid couplings provide significant excitation torques at discrete frequencies and have been identified as the cause of impeller failures. Variable speed electric motors of the Schrage or commutator type similarly provide excitation torques.

8

TRANSIENT EFFECTS

8.1

Reverse Flow upon Trip

Process heat exchangers in the delivery of BFW pumps constitute accumulators capable of supplying steam which may run the pumps in reverse rotation. Current practice is to provide a reliable non-return valve in each pump discharge line and NOT install reverse rotation locks. For' steaming' exchangers the risk is greater and two non-return valves in series are required at each pump discharge.

8.2

Flashing

Sudden reductions in the speed of turbine driven booster pumps cause steam flashing in the main pump where booster pumps supply BFW through feedheaters giving water temperatures> 200C at the main pump inlet. The rate of deliberate speed adjustment should be limited to give a pressure reduction rate not exceeding 3 bar/min in the feed heater. . 9

TEMPERATURE GRADIENTS IN CASINGS

Non-return valves leak a small amount. Such leakage can lead to thermal stratification in the casing and thence to casing distortion. Preventive measures are: (a)

The elimination of temperature gradients by slowrolling the standby pump, to induce a forward flow of feedwater through the pump and bypass return system, and to guarantee water mixing within the pump.


(b)

The provision of a bypass round the discharge non-return valve in order to ensure a sufficiently large reverse flow. The required flowrate depends on pump size and layout but lies in the range 2% to 10% of the pump rated capacity. It is ESSENTIAL to maintain adequate thermal insulation around the pump casing. A simple hole drilled through the non-return valve element suffers erosion and generates noise: the recommended arrangement is an external piped bypass embodying a multiple restrictor which can be empirically adjusted during commissioning. This arrangement re-introduces the risk of reverse steam flow from steaming process heat exchangers; consequently a reverse rotation detector should be provided to trip shut the discharge block valve. The distribution of water flows through the stationary pump is indeterminate unless the natural downward drift of cooled water is encouraged. Study the detail arrangement of the pump in conjunction with the manufacturer to identify possible stagnant zones. Extraction of cool water from such zones may entail the provision of an ancillary pump for return to the de-aerator. Select a glandless canned motor pump for this duty to avoid problems of air ingress.

(c)

Temperature gradients can be prevented by positively isolating the pump to create a truly stagnant water condition. The pump then has to withstand 'cold start' conditions of thermal shock. Both these requirements are difficult to realize; consequently this scheme should be considered only as the last resort.

10

INLET STRAINERS AND CASING PIPING CONNECTIONS

10.1

Strainers

Permanent strainers prevent serious damage caused by ingress of welding rod stubs. Specify an aperture of 1.0 to 1.5 mm in the strainer element. Fine mesh strainers are NOT recommended because they clog easily and bring pump operation to a stop.


Pre-commissioning should include acid cleans and wash-through procedures which adequately remove scale from the -pump inlet system before the pump is finally connected and started. . 10.2

Casing Piping Arrangements

Separate inlet, bypass and balance water lines should be provided for each pump. Manifolding each functional group of lines close to the de-aerator to avoid multiple vessel branches is permissible, with the following provisos: (a)

Fault rates more correctly describe the system, not the pump alone, consequently the expected reliability may not be achieved when main and standby systems share common elements.

(b)

Pressure drop variations may be excessive; this is particularly true of balance water lines, which should therefore be generously sized.

(c)

NEVER permit bypass lines to be returned to inlet lines or their manifold.

When reviewing layout arrangements remember that incorrect operation of valves accounts for a large percentage of cases of damage to BFW pumps. It is important to avoid accumulation of steam bubbles and to ensure pump priming: accordingly the rise in the discharge piping adjacent to the pump should be continuous and exceed a slope of 1 in 40. The line connecting the booster pump to the main pump inlet should be provided with a bypass line and orifice restrictor directly returning to the supply vessel in order to guarantee continuous venting. The flow should not be less than 10% of the booster pump BEP capacity.


11

SEAL COOLING

11.1

Packed Glands for Category 1 Pumps

The gland enclosure drain should be piped to a safe disposal point where efflux of flash steam will not obscure operationally important items. When the pump inlet water temperature exceeds 70C, cool flush water should be injected through a lantern ring. The design leakage rate from each gland should be taken as: .

where

D is the sleeve diameter,

mm

N is the shaft rotational speed,

r/s

Heat exchangers should be rated for 300% of this flow for each gland and shall cool the flush water to a temperature not exceeding 65C with cooling water or ambient air temperature at summer values. The flush should be BFW. Normally the source is a tapping from the main pump 1st or 2nd stage discharge. When the source pressure exceeds 25 bar g, a pressure limiting system should be provided, for example:


11.2

Mechanical Seals

(a)

For some Category 1 pumps a single mechanical seal may be used, having a water-cooled seat. Normally, the cooling water for such seats is run to drain in order to avoid seat distortion due to high or fluctuating cooling water pressure.

(b)

Specify mechanical seals for Category 2 pumps. The seal circulates BFW through an external loop which includes a heat exchanger with cooling water as the cooling medium. This cooler is mounted above the seal in order to promote thermosyphon action during pump standstill. The standby BFW pump is normally required to run for a short period after failure of the cooling water supply. For an orderly shutdown it is sufficient to have an overhead reservoir capable of providing the required flow for about 10 minutes.

SECTION TWO PUMP CONSTRUCTION FEATURES

12

MINIMUM FLOW THROUGH PUMP

12.1

Limit Due to First Stage Cavitation Damage

The usable capacity range of a pump narrows as S n increases because flow recirculation in the larger impeller eye results in increased cavitation damage potential a£ flows away from the stage best efficiency point. 1/4

When S n (NPSH p) > 0.8 check manufacturer's offer by comparing available NPSH against curve of 'onset cavitation NPSH' for the recommended minimum flow point.

12.2

Limit Due to Instability in Head/Flow Characteristic

Manufacturers find it difficult to shape the pump characteristics: experience indicates that all tenders should be thoroughly checked to verify that the stability requirement has been satisfied. Refinery Process Stream Purification Refinery Process Catalysts Troubleshooting Refinery Process Catalyst Start-Up / Shutdown Activation Reduction In-situ Ex-situ Sulfiding Specializing in Refinery Process Catalyst Performance Evaluation Heat & Mass Balance Analysis Catalyst Remaining Life Determination Catalyst Deactivation Assessment Catalyst Performance Characterization Refining & Gas Processing & Petrochemical Industries Catalysts / Process Technology - Hydrogen Catalysts / Process Technology – Ammonia Catalyst Process Technology - Methanol Catalysts / process Technology – Petrochemicals Specializing in the Development & Commercialization of New Technology in the Refining & Petrochemical Industries Web Site: www.GBHEnterprises.com

Note that the classic design methods based on impeller vane angle as the fundamental parameter do not afford the insight given by the area ratio design method - see Appendix G.

13

IMPELLER/DIFFUSER STAGE

13.1

Local Cavitation

This occurs when the water velocity is high, typically 60 m/ s relative to the metal surface. It is not related to the inlet NPSH. Peripheral and Barske type stages have open impellers which can be fully machined and which can be used at heads up to 1500 m when manufactured in 18/8 austenitic steels. Centrifugal pump stages with closed impellers and bladed diffusers of optimum design for high efficiency are manufactured as castings. Errors occur when using traditional pattern-making and casting techniques. A convenient classification is by differential head h across one stage, as follows: (a)

h < 40 m No special care needed.

(b)

40 m <. h < 100 m

Good quality commercial castings acceptable against visual examination by the inspecting engineer for gross defects. (c)

100 m < h < 250 m The castings should be subject to inspection for: (1)

Core shifts. Variations in vane peripheral thickness are unimportant but shifts altering vane angles should not be accepted.

(2)

Surface finish. All protuberances should be smoothly leveled and the casting techniques adequate for the production of a surface finish better than 25 µm Ra when viewed against a comparator tablet for a ground finish.


(d)

250 m < h < 660 m Material should be 13/4 Cr Ni steel. Performance adjustment by filing the rear surface of a vane should be forbidden. After adjustment by machining the vane tips (but not the shrouds) to smaller diameter, the upper surface of the vane should be smoothly faired. The trailing edge angle is limited to 0.2 radians.

Impellers should not be balanced by local removal of material from the disc/shroud surfaces. Such surfaces should run true. (e)

h > 660 m

Stages above 660 m head need thorough investigation of the design and manufacturing techniques. Special impeller materials will be required, exemplified by 17-4 PH stainless steel.

13.2

Labyrinth or Bushing Clearances

Small clearances improve the hydraulic efficiency only as measured during the works performance test. Such small clearances quickly wear to some equilibrium value; the consequent rubbing increases the risks of rotor seizure or induced whirl. (a) Clearances should not be less than the values given in the following table based on BS 4500.


(b)

Flexible rotors sag appreciably at rest but approach the nominal geometric centerline during operation. Eccentric clearance settings are NOT acceptable; bearing and casing centerlines should coincide. Bushing wear-rate can be dramatically increased when the pump is run at much reduced speed, eg during slowroll. Then either: (1)

the shaft sag should be less than the minimum radial clearance obtained after allowing the assembly tolerances.

(2)

the minimum speed should exceed the speed at which hydrodynamic forces become effective in supporting the shaft (commonly ~ 30% normal speed).

OR


13.3

Clearances Affecting Hydraulic Forces

Most multistage centrifugal BFW pumps have vaned diffusers. The radial gap between the impeller vane tips and the diffuser vane leading edges affects the pressure pulsations at the blade passing frequency, the radial hydraulic forces and t he broadband noise level. Let C be the radial gap, R the impeller vane tip radius and E the radial eccentricity of the shaft centre to the diffuser geometric centre. Then

C – E > 0.02 R Irregularities in castings may cause variations in the measured gap C. The inspecting engineer should verify that these measurements fall within the band of 100% to 140% of the least value found for C. Any machining of the leading edges of the diffuser vanes should be followed by hand finishing to produce the rounded entry profile. Pumps with double volutes should be subject to inspection to verify that both the radius to the tip and the leading edge profile of each volute cutwater is identical. Then, using the same nomenclature C - E > 0.04 R Impellers should have an ODD number of vanes to reduce perturbing torques.


14

ROTOR

14.1

Rotor Dynamics

Multistage centrifugal pumps in Categories 1 and 2 normally have operating speeds above the first lateral critical speed obtained by assuming the pump run on air; they rely on the damping provided by the pumped water to suppress dynamic effects. Unwanted perturbations chiefly arise from operating at part capacity, producing large low-frequency radial hydraulic forces. Category I pumps should have rolling element bearings, within the limits set by Specification GBHE-MAC-18-06A. Future systems using feed heaters may require the main BFW pump to have a stiff shaft capable of running dry for short periods.


Current practice is to supply the BFW pump directly from the de-aerator and permit the use of flexible shaft pumps which cannot withstand running dry. The pump shaft diameter and bearing span may be varied by pump designers for a given hydraulic duty, rotor speed and number of stages. Guidance on acceptable relationships of these parameters is given in Fig 14.1 which is based upon avoidance of excessive shaft bush wear. Note that such wear has led to shaft fracture in Zone A. Category 2 pumps should have the second critical speed of the rotor, as the natural frequency in air, not less than 130% of the maximum continuous speed. This frequency is easily measured and is a useful inspection check on manufacture.


FIG 14.1

HORIZONTAL MULTISTAGE PUMPS WITH RADIAL FACE MECHANICAL SEALS


NOTES ON FIG 14.1 1

At each impeller location the shaft may be grooved for a retaining ring or be stepped in diameter. Now the effective diameter should correspond to the shaft response acting as a beam in bending; for this it is sufficiently accurate to average all the diameters over the middle third of the span, ignoring stress concentration factors.

2

A useful retrospective check is to obtain the restoring force by measurements on the assembled rotor. The expression

Z X Z+2

d5 L3

Is then replaced by:

k. w. D δ as a close approximation where: W δ. D k

14.2

= = = =

weight of rotor (excluding coupling spacer) measured deflection of rotor at midspan diameter of shaft/casing bushes constant ~including g at 9.81 m/s2) of value 8.6 x 10-6

kg mm mm

Rotor Mechanical Balance

Each rotor element should be individually balanced before assembly on to the shaft. Dummy half-keys are necessary to ensure correct balancing. For these elements the balance quality according to BS 5265 should be better than G6.3 and better than G2.5 for pumps lying in zone B of Fig B3.1 above the 60% capacity line. Category 1 pump rotors may have the impellers located by free shaft sleeves locked by end shaft nuts. Such rotors should NOT be balanced after preassembly but checked as Clause 14.4.


Rotors for horizontal split casing pumps are fully assembled and have each impeller individually located. The check balance of the assembled rotor should show a quality better than G6.3. If it does not, the rotor should be disassembled, each element rebalanced and the shaft checked for straightness.

14.3

Torsional Critical Speeds

The exciting torques is normally small but so is the system damping. No torsional critical frequency of the pump/driver system, calculated for the rotor running in air shall be within 20% of the vane pass frequencies. This margin can be reduced to 6% when the critical frequencies have been measured when running in water. These margins include the normal variation in speed of induction electric motor drivers and an allowance of 0.6% for variation in to. frequency of the motor electric supply. The speed regulation performance of steam turbine governors should be individually assessed.

14.4

Rotor Assembly

Alignment of the hydraulic channels affects to performance. The centerlines of each impeller exit channel and the diffuser is normally aligned by adjusting the axial location of individual impellers. However, an over-riding requirement is that the impeller front shroud surface shall never overlap the surface of the diffuser channel, despite axial float in the thrust bearing and differential thermal expansion between rotor and casing.


CATEGORY 1

PUMPS WITH RING SECTION CASINGS

After pre-assembly outside the casing, the rotor should be checked by dial gauge measurements at stations along the rotor whilst it is supported in V-blocks. This test verifies that shaft sleeve and impeller mating ends are perpendicular to the axis. The criterion for rotor straightness is that first mode bends should have a maximum TIR (total indicator reading) limited to 1:10000 of the span between bearing midpoints.

CATEGORY 2

PUMPS WITH HORIZONTAL SPLIT CASINGS

The impellers are individually mounted on a stepped shaft so that the bores of the impellers progressively increase. This provision demands careful numbering of impellers. High speed rotors have the impeller bore machining limits adjusted to retain a transition fit when rotating at the maximum continuous speed. This requires an interference fit upon assembly. It is essential that impellers do not swash, i.e. when the rotor is assembled and supported on 'V'-blocks, a dial indicator bearing axially against the machined impeller shroud surface should show no variation in reading upon rotor rotation.

CATEGORY 2

PUMPS HAVING BARREL CASINGS WITH RING SECTION CARTRIDGES

Treat as Category 1 pump with ring section casing but add another criterion for straightness; viz. that second mode bend should be small, less than 10% of the first bend, assessed as the mean TIR at the 25% and 75% points along the span.

CATEGORY 2

PUMPS HAVING BARREL CASINGS WITH AXIALLY SPLIT CARTRIDGES

Treat as pumps having horizontal-split casings.


15

GLAND ARRANGEMENTS

15.1

Guarding

For all pumps specify that: (a)

Shaft guards at the gland will contain scalding spray leakage.

(b)

The gland enclosure drain size is adequate for gland failure conditions.

15.2

Soft-Packed Glands

Category 1 pumps normally have soft-packed glands to assist rotor damping and stability. Check that: (a)

Packing is one of the standard metric sizes of 6, 8, 10, 12.5, 15 mm square section. Current material used is Walkers Fortune 417. Shaft sleeves are provided, of 13% chromium steel preferably BS 970 420 S 45 hardened to 240 - 280 VH and ground to a surface finish better than 0.4 µmR a. It is essential that the sleeve surface is concentric with the shaft: this should be verified at the time when the rotor is inspected for straightness.

(b)

When the pump inlet water temperature exceeds 70C, cool flush water is injected through a lantern ring.

15.3

Mechanical Seals for Category 1 Pumps

Category 1 pumps may be fitted with mechanical seals having water cooled seats in order to avoid the need for a cooled BFW flush supply to a packed gland. The preferred seal is a Crane Type 502 to material code 468A for operation with pump inlet water temperatures up to a maximum of 105C. The sealbox should be 'dead-ended' but check the arrangement for vapor-locked zones. Refinery Process Stream Purification Refinery Process Catalysts Troubleshooting Refinery Process Catalyst Start-Up / Shutdown Activation Reduction In-situ Ex-situ Sulfiding Specializing in Refinery Process Catalyst Performance Evaluation Heat & Mass Balance Analysis Catalyst Remaining Life Determination Catalyst Deactivation Assessment Catalyst Performance Characterization Refining & Gas Processing & Petrochemical Industries Catalysts / Process Technology - Hydrogen Catalysts / Process Technology – Ammonia Catalyst Process Technology - Methanol Catalysts / process Technology – Petrochemicals Specializing in the Development & Commercialization of New Technology in the Refining & Petrochemical Industries Web Site: www.GBHEnterprises.com

Metric seals are required to one of the following series:

For pumps employing a balance disc check that: (a)

The seal springs are set at their correct compression with the pump balance disc gap brought to zero by jacking the rotor axially.

(b)

a rotor stop is provided to limit the axial displacement to 1.5 mm.

15.4

Mechanical Seals for Category 2 Pumps

Specify Borg Warner Mechanical Seal Type D or DRT to material code 5H4A (BW) with a seal water heat exchanger to Bulletin 1860-15 with Inconel tubing (Austenitic stainless steel tubing is NOT acceptable). Specify the Type DRT seal where the peripheral speed (based on the seal diameter) exceeds 20 m/s. The seal circulates BFW through an external loop which includes a separator and a heat exchanger with CW as the cooling medium. This cooler is mounted above the seal in order to promote thermosyphon action. Particles of iron oxide are prevented from reaching the seal faces by a magnetic particle separator. Refinery Process Stream Purification Refinery Process Catalysts Troubleshooting Refinery Process Catalyst Start-Up / Shutdown Activation Reduction In-situ Ex-situ Sulfiding Specializing in Refinery Process Catalyst Performance Evaluation Heat & Mass Balance Analysis Catalyst Remaining Life Determination Catalyst Deactivation Assessment Catalyst Performance Characterization Refining & Gas Processing & Petrochemical Industries Catalysts / Process Technology - Hydrogen Catalysts / Process Technology – Ammonia Catalyst Process Technology - Methanol Catalysts / process Technology – Petrochemicals Specializing in the Development & Commercialization of New Technology in the Refining & Petrochemical Industries Web Site: www.GBHEnterprises.com

16

AXIAL HYDRAULIC THRUST BALANCE METHODS

16.1

Opposed Impeller Configuration

This demands a complicated casing shape; a specific assurance on casting quality control should be obtained from the foundry nominated by the pump manufacturer. A thrust bearing is required to take the residual thrust: to ensure that this thrust is unidirectional the number of single-entry impellers should preferably be odd.

16.2

Balance Disc

Current practice is to restrict the use of balance discs to Category 1 pumps. The balancing thrust is automatically adjusted by the change of pressure in cavity 'X' consequent upon axial movement of the rotor changing gap “ “.

Balance is achieved when the gap is of the order of 60pm, depending on the clearance in the preceding piston section. Efforts to increase efficiency by decreasing the leakage flow through a reduction of ~ should be resisted, otherwise the pump becomes very sensitive to particulate contamination of the feedwater. Rubbing occurs during transients. SULFINUZ anti-galling surface treatment has been used successfully to lengthen disc life but the performance was erratic. Current practice is to use hardened chromium steels.


The dynamic compliance of the balance disc gives a natural frequency of axial vibration of the rotor. Interaction may occur with metal diaphragm type shaft couplings whose spacers have their own intrinsic natural frequency bands of axial vibration. Such an effect should be specifically assessed for pumps running above 65 rls. For truly parallel balance disc faces the restoring force rapidly decreases as the faces come nearly into contact. This effect is reduced by extending the face relief nearly to the disc periphery. It is essential that disc faces are NEVER machined convex to each other.

Because the action of balancing demands freedom of rotor axial displacement, the first choice of bearing type is either the journal or the cylindrical roller.

16.3

Balance Piston

This is required for Category 2 pumps. A piston does not provide exact balance: a thrust bearing is required to take the residual thrust. The piston should be proportioned to ensure that this residual thrust is unidirectional over the normal operating range. There is a hydraulic self-centering action, but this is reduced if the piston is serrated.


Current practice is to ensure that: (a)

the bush is truly cylindrical. At H > 1250 m the bore should be finished by a floating stone hone.

(b)

the assembly positioning aligns the bush along the geometric centerline between bearings.

17

CASING

17 .1 Casing Type CATEGORY 1

PUMPS

These are invariably of the ring section type having one cell per stage, the whole being held together by through bolts. Both rotor and casing are assembled progressively cell by cell; consequently: (a)

it is important that all parts are match marked

(b)

the fit of the impeller on the shaft should be a light driving fit.

Symmetrical loading by the through bolts is important. This implies that the extension of each through bolt is measured and brought to the same value. Torque-spanners are useful only for small pumps where the bolt size is less than M20. Check that the Manufacturer's Manual includes specific instructions on the sequence and method of tightening the through bolts. This type of casing is prone to thermal distortion because the through bolts are not in contact with the feedwater and the uniformity of bolt temperature depends on the quality of the lagging application. Consequently this type should be used only where the pump draws feedwater directly from the de-aerator vessel and the water temperature is less than 120C.

CATEGORY 2

PUMPS - NOMINAL SYSTEM PRESSURE 43 to 70 BAR

Ring section casings may be used subject to a limit of 7 stages. Such pumps have been used in double pump arrangements, both pumps being driven by a double-ended electric motor, but connected in series for BFW flow. Refinery Process Stream Purification Refinery Process Catalysts Troubleshooting Refinery Process Catalyst Start-Up / Shutdown Activation Reduction In-situ Ex-situ Sulfiding Specializing in Refinery Process Catalyst Performance Evaluation Heat & Mass Balance Analysis Catalyst Remaining Life Determination Catalyst Deactivation Assessment Catalyst Performance Characterization Refining & Gas Processing & Petrochemical Industries Catalysts / Process Technology - Hydrogen Catalysts / Process Technology – Ammonia Catalyst Process Technology - Methanol Catalysts / process Technology – Petrochemicals Specializing in the Development & Commercialization of New Technology in the Refining & Petrochemical Industries Web Site: www.GBHEnterprises.com

CATEGORY 2

PUMPS - NOMINAL SYSTEM PRESSURE 43 to 100 BAR

The horizontal-split casing type enables the rotor to be assembled as a unit prior to insertion into the casing. Volute stages are often used to simplify the casting; it is then important that the two half-casings match at the split-line. The inspecting engineer should visually examine the assembled casing to verify that the volute surfaces have been ground to obtain a smoothly continuous surface over the joint. Vaned diffuser rings may be inserted to form cells similar to ring section pumps. Check the method of ensuring simultaneous sealing of each ring to the casing and of the two casing halves: alloy surface weld deposit maybe needed.

CATEGORY 2 PUMPS - NOMINAL SYSTEM PRESSURE ABOVE 100 BAR The barrel type casing is required. The cartridge is usually assembled as a ringsection pump but when the limits for rotor dynamics are approached the cartridge assembly in an axial split arrangement should be considered.

17.2

Casing Connections

(a)

Configuration Horizontal-split casings have side connections in order to retain the advantage of removing the upper half casing without disturbing the inlet or discharge piping. For ring section and barrel casings, specify top discharge and top or side inlet connections. Maintenance of barrel casing pumps is carried out by withdrawing the pump cartridge, leaving the casing installed. Consequently the high pressure discharge branch may be butt-welded to the piping.

(b)

Rating For Category 1 pumps specify that the inlet and discharge connections are rated for the same pressure.


Category 2 pumps may have the inlet connection rated at the higher value of: (1)

Nominal 40 bar rating

(2)

120% of the pressure setting on the relief valve fitted between the pump and the inlet block valve.

(c)

Piping Loads

The connections to ring-section pump casings are broken for maintenance; consequently the piping should be designed to have NO forces, nor moments, and be separately supported so that the pump can be removed without disturbing the pipe system. Horizontal-split casings are torsionally very weak when the upper half-casing has been removed; consequently the applied moments should be zero when the associated piping is cold. Note that this condition occurs not only when all pumps are shut down but also when companion pump(s) remain in operation.

18

MATERIALS

CATEGORY 2

PUMPS

Current practice is to use the 13% chromium steels because of their good thermal and mechanical properties and immunity from corrosion by demineralized feedwater. Because the feedwater is de-aerated, castings for horizontal-split casings may have the Chromium content reduced to the lower limit of 5%. High quality complex castings have become progressively more difficult to obtain. Consequently for horizontal-split casings, the inspection plan and weld repair techniques should be agreed with the foundry at the time of placing the purchase order. Barrel pump casings may be in carbon steel. The preferred casting alloy for impellers and diffuser vane rings is 13/4 Cr Ni steel. Austenitic stainless steels should not be used; in particular they should NOT be used for shafts.


For rubbing pairs, viz. impeller/casing wear rings and shaft sleeves/bushings, it is advantageous for the stationary component to have the highest practicable hardness. Current practice is to specify casing wear rings in PH stainless steel aged to produce a hardness of 300-340 HV and ground to a surface finish better than 1.6 µm Ra upon visual inspection against a comparator tablet. Impeller wear rings are vulnerable to disturbance on heavy rubs; consequently the preferred arrangement for the highest intrinsic reliability is to integrate the wear ring into the impeller casting. This favors the use of 13/4 Cr Ni alloy heat-treated to obtain a hardness of 240-280 VH, also ground to a surface finish of 1.6 µm Ra. Such a rubbing pair is reported to be good for speeds up to 60 m/s. When the rubbing speed is less than 35 m/s, Ni-Resist casing wear rings may be used.

CATEGORY 1

PUMPS

The important feature is that steam systems in Category 1 normally do not employ demineralized feedwater and the pump inlet temperature is less than 120 C. Early practice using bronze impellers and fittings has been discontinued because of erratic deterioration of bronze components and the high thermal expansion of this material. Individual cells of ring section pumps may be in grey cast iron to BS 1452 Grade 14 or 17 but specify the end-sections in carbon steel or nodular cast iron for design pressures above 16 bar g. Impellers and diffuser rings may be in grey cast iron to BS 1452 Grade 14 or 17 provided that H < 40 m, except for the first stage which should be in 13/4 CrNi steel when S n >0.28. Casing wear rings are normally Ni-resist. For demineralized feedwater follow the practice for Category 2 pumps.


19

DRIVER CONSIDERATIONS

For Category 2 pumps, check that the torque/speed curve for the motor during run-up is shaped to give both reasonably uniform acceleration and a total run-up time of 2-6 seconds. Longer run-up times demand pumps whose rotor parameters are within Zone C as described in Clause 14.1.


BIBLIOGRAPHY 1

Convention by Inst Mech Eng on Advanced - Class Boiler Feed Pumps Sept 1970, (Proceedings 1969-70, Vol.184, Part 3N).

Paper 1

A feed pump design concept for 660-MW generating sets, by J Richardson, BSc., C.Eng., MI Mech.E., and J M Taylor.

Paper 2

The design principles for boiler feed pumps for CEGB. 660-MW units, by F 0 J Otway, MA, C.Eng., MII1ech.E. When the Central Electricity Generating Board increased the unit rating of their turbo generators from 500 MW to 660 MW, the design basis of the boiler feed pumps was completely reviewed. The prime requirement was that the pumps should be made less sensitive to mechanical fault and capable of surviving disturbed suction conditions without failure. At the same time they were to be capable of rapid replacement. This paper explains the decision that the pumps should have only two or three stages, with stiff shafts. Gland security and light-load protection are discussed.

Paper 3

Boiler feed pump design for maximum availability, by R Weldon, BSc., C.Eng., MI Mech.E Describes Sulzer design for 660 MW stations.

Paper 4

Advanced-class boiler feed pumps for 660 MW generators, by TO Leith, BSc., C.Eng., MI Mech.E., J R McColl, BSc.C.Eng, MI Mech E, and M L Ryall, BSc, C.Eng, MI Mech E Describes Weir design for 660 MW station.

Paper 5

Development of a single-stage boiler feed pump for nuclear power stations, by A A Gasiunas, Dipl.Ing


The paper discusses the pressure pulses that are generated when impeller vanes travel past stationary vanes. The influence of geometrical and hydraulic non-symmetry on radial forces in a double volute casing is dealt with. Test results of a double suction model boiler feed pump are presented, including descriptions of durability runs with paint-coated hydraulic surfaces. Paper 6

Development of boiler feed pumps in Czechoslovakia, by V Hladis, B.Sa, and J Kupa Gives useful description of ring section pumps operating with BFW at 160C including measurements on through bolt temperature differences and casing distortion.

Paper 7

The single-stage high-speed nuclear feed pump, by A R Bush, B Mech.Eng, and J E C Valentin Feed pumps applied to early large-size light water moderated nuclear reactors in the United States were modifications of existing split-casing volute-type pipeline pumps. Reliability was the key consideration. The final selection was a single-stage solid casing with an integral diffuser and condensate injection seals. The manufacturing problems are discussed, together with performance tests run on the first machines of this type.

Paper 8

Application of the thermodynamic method of measurement for the determination of the boiler feed pump efficiency in large electrical power units of Electricite de France, by J S A Guitton, Ingenieur Ecole Centrale, and H Procaccia, Ingenieur EEIP

Paper 9

Dynamic hybrid bearing characteristics of annular controlled leakage seals, by H F Black, MSc., C Eng, MI Mech E, and D N Jenssen, BSc. The dynamic bearing characteristics of plain seals having appreciable length-to-diameter ratio are analyzed. Some experimental results are given and compared with theoretical predictions. It is shown that length-to diameter ratio has a considerable effect on seals of appreciable length, e.g. balance pistons, and that dynamic components of bearing action due to shaft rotation are comparable with the purely hydrostatic centering forces.


Paper 10

Radial forces in centrifugal pumps with guide vanes, by P Hergt, Dipl-Ing., and P Krieger At partial and overload conditions, radial decentralizing forces act upon the rotor of a centrifugal pump with diffuser guide vanes if the impeller is out of centre. The magnitude of these forces increase with growing eccentricity. At very small flows, these forces become non-stationary. They rotate at a considerably lower speed than the rotor speed. The paper discusses the effects and resulting shaft deflections and vibrations, from measurements On two experimental test rigs.

Paper 11

Feed pumps for modern steam boiler applications, design, development, and operation, by P S Neporozhnii and A K Kirsh

2

Design of Modern Boiler Feed Pumps by H H Anderson

3

Development of high-pressure boiler feed pumps in Britain during the last decade by G F Arkless. COny on Steam Plant Ancill1ary Equipment Proc lnst Mech Eng 1966/7 181 (Pt.3N) 6.

4

Analysis of cavitation damage in commercial marine condensate pumps. PE Paashaus ASME - SNAME Meeting New York December 1964. This paper refers to the "bulging" of impeller shrouds upon "hammering" when operating on free suction control.

5

Horenburg, 0: Schaden an Kesselspeisepumpen. Hinweis zur Schadenverhutung durch Auswertung von Schadenstatisiken. Der Maschinenschaden 43 (1970), No 4, pp 135/147. Describes pump damage due to thermal distortion of casings and ingress of foreign bodies through strainers.

6A

Honold, E : Vergleich von Kesselspeisepumpen fur hohe Enddrucke in Glieder-und Topfbauweise im Hinblick auf die Entwicklung grober Kraftwerksblocke, Mitt der VGB (June 1966). No 102, pp 149/152.


6B

Honold, E : Rin kritischer Vergleich von Topfund Gliederpumpe am Beispiel der Entwicklung der Hochstdruckkesselspeisepumpen in Deutschland und den USA. KSB Technische Berichte, No 1 (October 1960). These 2 papers discuss the choice of barrel or ring-section pump casings.

7A

Strub, RA : Abfall des Saugdruckes von Speisewasserpumpen bei starken Lastschwankungen. Technische Rundschau Sulzer 42 (1960), No 3, pp 41/44.

7B

Wollschlaeger, K : MaBnalmen zur Verbesserung der zulassigen Druckabsenkungsgeschwindigkeit bei Kesselspeisepumpen. Energie 18 (1966), No I, pp 16/18.

7C

Rahlwes, H : Untersuchung zur Klarung Speispumpen bei Gleitdruckentgasung. No 106, pp 61/67. von Zulaufstorungen an Mitt VGB (Feb 1967),

7D

Stonner, A : Ein Beitrag zur Schadenverhutung an Hochdruckkesselspeisepumpen. Energie 18 (1966), p 360. These 4 papers discuss flashing within pump casings and inlet lines. Paper 7D describes effects of leaking stop valves on standby pumps.

8

Leakage and Hybrid Bearing Properties of Serrated Seals in Centrifugal Pumps. H F Black & E A Cochrane, Sixth International Conferennce on Fluid Sealing 1973, 65, pp 61 - 69.

9

W Schumacher Wear and Galling of Nitrogen-Strengthened Stainless Steels Report in Machine Design August 1983 on paper presented at Wear of Materials 1983 Conference, Reston, V.a. Apr 1983. Indicates that nitrogen addition improves strength and hardness of stainless steels but does not improve wear resistance. Increasing nickel content decreases the wear resistance. Best material NITRONIC 60.


APPENDIX A NOTES ON BFW PUMP/DRIVER ARRANGEMENTS I

HISTORICAL PRACTICE FOR POWER STATIONS

In the UK early power stations had both main and standby pumps driven by steam turbines. The 1920's saw the introduction of the arrangement with the main pump driven by an electric motor and the standby pump by a steam turbine. This has remained the preferred practice for petro-chemical plants. The steam turbine driven high-speed main pump, with a 50% capacity starting/standby pump having electric motor drive, appeared about 1955. A variation of this arrangement used a main pump driven through a hydraulic coupling and gearbox from the turbo alternator set.

2

EXTRACT FROM ANSI/ASME BPV - VII, 1977 EDITION

C5.401

Feedwater should be available at the boiler at flow rates and pressures which are adequate to take care of any emergency.

C5.403

A spare feedwater pump or injector, in addition to the feedwater equipment required by PG-61.I, Section I, is preferable. Where feedwater pumps are electrically driven and there is no fully independent auxiliary source of electric supply, there shall be maintained, ready for service, steam-driven feed pumps or injectors of sufficient capacity to safeguard stoker-fired boilers in case of failure of electric power. This recommendation also applies to boilers fired by other methods if furnaces contain large amounts of refractory or are arranged to accumulate slag in the bottom.

C6.803

Normal Operation A connection should be provided between the pump discharge line and a point on the suction system, preferably the deaerator, as a means of preventing overheating of the pump when it is required to operate at shut-off pressure or extremely low rates of delivery. This line, which is commonly known as the recirculating connection, should be connected into the discharge line between the pump and the check valve.


The recirculating line should be connected back into the suction system at some point where the heat added by the pump will be dissipated before the water can re-enter the pump suction. This line should be provided with a stop valve and an orifice of such size as to permit recirculation at the minimum flow rate stipulated by the manufacturer of the pump. The stop valve should be locked or sealed in the wide-open position whenever the pump is in operation or ready for operation. All valves in hydraulic-balancing drum-leak-off lines should be locked or sealed in the wide-open position whenever the pump is ready for or in operation.

3

EXTRACT FROM PG-61 (GENERAL REQUIREMENTS FEEDWATER SUPPLY)

61.1

Except as provided for in PG-61.2 and PG-61.4, boilers having more than 500 sq.ft. (47 m2) of water heating surface shall have at least two means of feeding water. Except as provided for in PG-61.3, PG-61.4 and PG61.5, each source of feeding shall be capable of supplying water to the boiler at a pressure of 3% higher than the highest setting of any safety valve on the boiler. For boilers that are fired with solid fuel not in suspension, and for boilers whose setting or heat source can continue to supply sufficient heat to cause damage to the boiler if the feed supply is interrupted, one such means of feeding shall be steam operated.

61.2

Except as provided for in PG-61, a boiler fired by gaseous, liquid or solid fuel in suspension may be equipped with a single means of feeding water provided means are furnished for the shutting off of its heat input prior to the water level reaching the lowest permissible level established by PG60.

61.5

A forced-flow steam generator with no fixed steam and water line shall be provided with a source of feeding capable of supplying water to the boiler at a pressure not less than the expected maximum sustained pressure at the boiler inlet, as determined by the boiler manufacturer, corresponding to operation at maximum designed steaming capacity with maximum allowable working pressure at the superheater outlet.


4

EXTRACT FROM GERMAN CODE

The main points of the German statutory code for feedwater installations as published in the Bundesanzeiger (Federal German Official Gazette), Vol.4, No.19, 29 January 1952 are as follows. (1)

Every steam-raising plant must be equipped with at least two feedwater systems.

(2)

If only two feedwater pumps are installed, each must be capable of supplying: (a)

1.6 times the maximum continuous output of all boilers fed by the system when the boilers are not equipped with automatic feedwater control or the total steam output of the boiler plant does not exceed 30 t/h.

(b)

1.25 times the maximum continuous output of all boilers fed by a common feedwater system if the boilers are equipped with automatic feedwater control and the total steam output of the plants exceeds 30 t/h.

(c)

1.25 times the maximum continuous output of all boilers fed by the system where the pumps are driven direct from the prime mover. The total quantity of feedwater required is the sum of the quantities required for the maximum continuous output of all boilers fed by the system. According to German Standard Specification DIN 2901, the maximum continuous output is 25% above the normal and 'boilers fed by the system' include not' only those in regular use but also stand-by boilers, unless they have been certified as permanently out of service. The factor of 1.25 includes a 5% allowance for blowdown losses. If the blowdown losses etc. are greater than those corresponding to this figure the factor must be increased accordingly.

(3)

If more than two feedwater pumps are installed and a pump or power failure causes the pump having the maximum capacity to cut out, the remaining pumps operating in parallel must be capable of delivering at least 25% more feedwater than is required by the boilers.


(4)

The feed pumps must be capable of delivering the required amount of feedwater against a pressure 10% higher than the rated (permissible) maximum steam pressure in the boiler with the safety valves blowing off plus the resistance (pressure losses between the feed pump and the boiler). Each pump must be so arranged that in the event of a sudden stoppage of the pump no reversal of flow is possible.

(5)

At least two independent sources of power must 'be available to drive the feedwater pumps. It is however permissible in the case of steam-driven pumps to supply them from a common steam system. Only one actual supply 11 ne to the prime mover is required irrespective of whether the drive is by steam or electricity.

The feed pumps must be so connected to the source of power (whether purely electric or dual steam and electric) that, in the event of failure of one of the sources, the pumps which can still be retained in service are capable of supplying 1.25 times the required feedwater quantity when operating in parallel. Feed pumps which have cut out because of some failure or other but which can be switched to an alternative power supply are included in those regarded as available for service. The table gives various arrangements which are permissible with steam and electric drivers or combinations of both.

S = Steam driven

E = Motor driven

Output of the individual pumps in % of the maximum continuous load



APPENDIX B PROPERTIES OF WATER B.1

DATA SOURCE 'STEAM TABLES IN SI UNITS' Published by CEGB, Sudbury House, London, ECIA, 7AU

B.2

CALCULATION OF RATED DIFFERENTIAL HEAD H The density of water depends on both pressure and temperature. Now the temperature rise ΔT through each stage of a pump is given by:

The density change due to this temperature rise, together with the change due to the compressibility of water, is then used to calculate the differential head across the stage. Integrating these stage heads over the whole pump and referring the head to the inlet conditions gives the approximate relation:


APPENDIX C DEFINITION OF PUMP CATEGORY The development of BFW pumps has been matched to advances in the size and steam conditions of power stations. Stages in this historical development have been distinguished by features conveniently grouped by reference to the nominal pressure of the steam system.

Note that the pump casing pressure rating is higher than this nominal pressure. These categories group together construction elements of current horizontal multistage centrifugal pumps whose duties are akin to the duties of former power station pumps at their appropriate stage of development.



FIGURE 1 - TYPICAL CROSS-SECTION OF PUMP IN CATEGORY 1


FIGURE 2A -

TYPICAL GROSS-SECTION OF PUMP IN CATEGORY 2 SHOWING AXIALLY SPLIT CASING


FIGURE 2B - TYPICAL CROSS-SECTION OF PUMP IN CATEGORY 2 SHOWING BARREL CASING WITH RING TYPE CARTRIDGE


FIGURE 3 - TYPICAL CROSS-SECTION OF PUMP IN CATEGORY 3 FOR POWER STATION APPLICATIONS WITH DRY RUNNING CAPABILITIES


APPENDIX D TRANSIENT PHENOMENA IN DE AERATORS Experience in power station pumps has shown that a main turbine trip can cause the boiler feedwater pump to stop delivery. The instant that the turbine trips, both the bleed heating of the demin water and the steam supply to the de-aerator are interrupted, leading to a drop in the pressure in the deaerator. This drop may be so great that the water in the line to the boiler feed water pump starts boiling. Nomenclature H M Q

Enthalpy kJ/Kg Mass Kg Flow Kg/sec

Suffixes c i o s d h p

cold well water into deaerator water out of deaerator steam into deaerator deaerator heater pump inlet

For normal running conditions a heat balance gives:

When Qs - 0 the cool entry water Q i is heated the steam in the vapor space of the vessel. This tends to reduce the pressure in the vapor space and cause the hot water already in the deaerator to flash. The pressure decays as the bulk water temperature reduces. Eventually equilibrium is regained at the temperature of the inlet water. Meantime a transient condition exists where the water entering the pump is hotter than the water at the deaerator water surface. Thus the available NPSH at the pump inlet is temporarily reduced. Refinery Process Stream Purification Refinery Process Catalysts Troubleshooting Refinery Process Catalyst Start-Up / Shutdown Activation Reduction In-situ Ex-situ Sulfiding Specializing in Refinery Process Catalyst Performance Evaluation Heat & Mass Balance Analysis Catalyst Remaining Life Determination Catalyst Deactivation Assessment Catalyst Performance Characterization Refining & Gas Processing & Petrochemical Industries Catalysts / Process Technology - Hydrogen Catalysts / Process Technology – Ammonia Catalyst Process Technology - Methanol Catalysts / process Technology – Petrochemicals Specializing in the Development & Commercialization of New Technology in the Refining & Petrochemical Industries Web Site: www.GBHEnterprises.com

Two further phenomena complicate the real behavior: 1

The contents of the deaerator are not perfectly mixed - this causes the loss of NPSH to be underestimated.

2

The flash steam creates a band of bubble/water mixture extending down from the free surface by an amount depending on the severity of the transient. Note that this does not alter the effective head: the bubbles change both the apparent density am the surface level of the mixture in the vessel. However, if bubbles are formed in the inlet line, unstable two-phase flow can occur with complete failure to pump forward.

If the flow through the pump is maintained, Q0 is constant. When the steam to the deaerator is turned off, Qs - 0 at time t = 0 Then:

where H is the enthalpy in the deaerator at time t


Similarly if both steam an:! feed heating are stopped at the same time the enthalpy in the deaerator becomes:

The maximum loss in NPSH can be estimated from:


The enthalpy at the pump at time t is that which had occurred in the deaerator at a time:

The difference between the two enthalpies so obtained can be used to calculate the drop in NPSH provided that no vapor bubbles form in the pump inlet line.

On systems with a large transient decay it is necessary to optimize the available head during transient conditions by examining the inlet pipe to the pump Decreasing the diameter increases the velocity in the pipe and so increases the friction losses. On the other hand the water volume in the inlet pipe itself is reduced and as the transient decay is directly proportional to the inventory the transient loss is reduced. To optimize a system it is best to tabulate the friction loss and the transient loss for various pipe diameters. For the available NPSH to be a maximum during transients the velocity will be between 2 - 5 m/secs, the higher value applying to systems sensitive to transients. Reference: 'Analysis of Feedwater Pump Suction Pressure Decay under `Instant Turbine Load Rejection' - by G S Liao and P Leung (Journal of Engineering for Power, April 1972)

EXAMPLES To show the magnitude of the effect, two examples have been worked out. One applies to a power station with a small de aerator capacity, where steam heating to the deaerator and feed heat is lost at the same time. The second one applies to a deaerator of the type used on a typical chemical plant.


EXAMPLE

I - POWER STATION

Conditions: Capacity: (a) Deaerator

- 180 seconds

(b) feed section

- 60 seconds

(c) pump inlet lines

- 10 seconds

Temperatures and total heat s:

On loss of all heating steam, the deaerator pressure will decay as follows (using Formula 5):


TABLE I

This shows that the total loss in available NPSH is 9.7 m am occurs about 70 seconds after the steam has turned off. This compares to the value of 10 meters estimated by using Formula 7.


EXAMPLE

2

A fault only affects the steam heat to the deaerator. It does not affect the feed heat. The enthalpy in the de aerator will change slowly and the transient head loss is negligible.

APPENDIX E NOTES ON PROPRIETARY LEAK-OFF VALVES This type of valve combines the functions of a non-return valve and a minimum flow bypass valve. Typical arrangements of the Schroeder and Yarway leak off valves are shown in Figures E.l and E.2 respectively. The method of operation is ON-OFF. On low flow, the non-return valve element activates the kickback valve, allowing bypass flow of the desired rate. The kickback valve is not closed until the flow through the non-return valve element increases considerably.


Important features are as follows:(a) Such valves are subject to considerable internal wear, and typically have a life expectancy of 1000 hours on continuous kickback duty. (b) No alternative materials are commercially available for high pressure duties. Manufacturers allow for high pressure drops by increasing the number of let down stages within the kickback valve. The internals can be replaced without disturbing the main body of the valve from the pipework. (c) High noise levels are generated by leak off valves when on continuous kickback duty. Current practice is to employ these valves only for Category 1 installations. For Category 2 installations, adequate 'demand reliability' assessments are not available; consequently current practice is to use a permanent bypass or a modulated bypass valve.


FIGURE E1 -

SPECIAL COMPENSATING BY-PASS CHECK VALVE (DIAGRAM) SCHROEDER


FIGURE E2 -

SPECIAL COMPENSATING BY-PASS CHECK VALVE (DIAGRAM) YAAWAY


APPENDIX F NOTES ON CASTINGS FOR HIGH-DUTY IMPELLERS In order to obtain satisfactory castings it is essential to discuss the quality requirements with the foundry. The following notes are intended for use in drafting the specification for a particular pump in conjunction with the pump manufacturer. NOTES: 1

Material Specification High duty impellers are normally cast in 13/4 Cr Ni steel, typically ASTM A 487 CA6 NM.

. 2

Dimensional Tolerances and Surface Finish


2.1

Surfaces Left As-cast

Tolerances important for the pump Q-H characteristic (from the area ratio theory) are: - axial widths b to be within + 3% of design value - vane widths t to be within + 12% of design value The surface finish should be better than 25 µm Ra and the waviness better than 1 mm in 50 Mm.

2.2

Setting-up locations for machining

(a)

Swash The front surface of each intervane passage shall lie within + 0.01 b of the plane B which passes through the mean position of these surfaces and is perpendicular to the rotation axis.

(b)

Concentricity The datum shall be the circle A through the transition from front shroud to vane root. The asymmetrical deviation of the surface from the datum circle shall not exceed 0.1% of the circle diameter.

2.3

Uniformity relations

(a)

Vane spacing The peripheral spacing of the vanes shall be determined by measuring the chord length between vane trailing edges after proof machining the impeller outer diameter. All values shall lie between + 1.5% of the mean of the values.


(b)

Vane fillets The fillet radius shall be 40% to 60% of vane width t, subject to a limit of 40% of passage width b.

(c)

Inlet surface concentricity If S n > 0.55 the inner surface C shall be fully machined to give a surface finish better than 3.2 µm Ra . The resulting boundary between machined and as-cast surfaces shall be smoothly faired by hand finishing.

(d)

Balance Notwithstanding that individual tolerances have been satisfied, the distribution of variations shall be determined by check-balancing the impeller after preliminary machining. The balance quality shall not be worse than G.20 grade of BS 5265 before material removal for balancing.

2.4

Machining allowance and finish

(a)

Preliminary machining shall leave all dimensions oversize by 1 to 4 mm and give a surface finish not worse than 16/ µm Ra in order to obtain satisfactory balance and dye penetrant tests.

(b)

Any final machining shall give a surface finish better than 3.2 µm Ra on wetted surfaces.

3

Quality Control Points

3.1

Test samples shall be cut both tangentially and radially from material added to the hub ring and shall be identified by an Inspector's stamp.

3.2

All surf aces shall be subject to a dye-penetrant crack detection check immediately before final machining.

3.3

Shot blasting is not accepted.

3.4

Weld repairs to castings shall precede final heat-treatment.

3.5

Certification shall be to DIN 50049/3.1B.


SPECIAL PRECISION CASTINGS BY THE LOST-WAX PROCESS

Dimensional tolerance - +0.5% to -0.5% Surface Finish - 3.2 mR a Finished Weight - 10 kg Ruling Section - 5 - 500 mm 1.0 to 50 mm wall thickness

Material: Air cast Air cast Air cast

- ANC.6A - PE.IO - ANC.3B

APPENDIX G AREA RATIO METHOD Anderson conceived the parameter of the area ratio Y to relate the flow conditions at the pump impeller outlet to those at the casing volute throat, where:

Worster showed that the Q-H-E performance characteristics of a centrifugal pump are chiefly determined by Y and not by the impeller blade angle. Following Thorne, the impeller outlet area is defined as:

Where

Z is the number of impeller vanes b is the impeller passage width at the periphery


The throat area A t is the sum of the two corresponding areas for a double volute or the sum of the outlet areas between diffuser vanes defined as for the impeller. Intrinsic assumptions are that: (a)

High efficiency pumps are being sought for clean liquids of low viscosity, so that vane incidence angles and shapes are properly matched to the flow conditions.

(b)

Impellers have the optimum number of vanes, given approximately by 15. β where β is the vane angle to the tangent, in radians.

(c)

Impeller inlet dimensions are not distorted to obtain the exceptionally low NPSH capability corresponding to pump operation at suction specific speed (S n) values above 0.4.

Commercial and manufacturing considerations result in low specific speed pumps having large area ratios, where

However, when pumps having a stable Q-H characteristic or a non-overloading E-Q characteristic are required, then smaller values of Yare needed where:


The efficiency does not vary strongly with Y but there is an advantage for values of Y near unity.

REFERENCES I

Centrifugal Pumps - An alternative theory H H Anderson. Proc.l.Mech.E 1947 vol 157

2

The interaction of impeller and volute in determining the performance of a centrifugal pump. R C Worster. BHRA Report RR.679. 1960

3

Design by the Area Ratio Method E W Thorne. BPMA Sixth Tech Conf. Paper C2 1979


DOCUMENTS REFERRED TO IN THIS ENGINEERING DEPARTMENT DESIGN GUIDE This Engineering Design Guide makes reference to the following documents. AMERICAN SOCIETY FOR TESTING AND MATERIALS ASTM A 487

Steel Castings Suitable for Pressure Service (referred to in Appendix F Note I).

DEUTSCHES INSTITUT FUR NORMUNG DIN 50049

Materials Testing (referred to in Appendix F Note 3.5).

BRITISH STANDARDS BS 970

Specification for Wrought Steels for Mechanical Engineering Purposes (referred to in Clause 15.2).

BS 1452

Specification for Grey Cast Iron (referred to in Clause 18).

BS 3468

Austenitic Cast Iron (referred to in Clause 15.3).

BS 4500

ISO Limits and Fits (referred to in Clause 13.2).

BS 5265

Mechanical Balancing of Rotating Bodies (referred to in Clause 14.2 and Appendix F Note 2.3).

GBHE ENGINEERING SPECIFICATION GBHE-MAC-18-06A

Bearing Arrangements for Machines (referred to in Clause 14.1).

ENGINEERING DESIGN GUIDES GBHE-MAC-1014 Integration of Special Purpose Centrifugal Pumps into a Process (referred to in Clauses 4.3 and 6). GBHE-MAC-5101 Reliability Analysis - the Weibull Method (referred to In Clause 2).



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