Boiler Feed Water Pumps

7/26/2019 Boiler Feed Water Pumps

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Engineering Design Guide:GBHE-MAC-1504

Boiler Feedwater Pumps

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


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Engineering Design Guide: Boiler Feedwater Pumps

CONTENTS SECTION

0 INTRODUCTION/PURPOSE 5

1 SCOPE

SECTION ONE - INTEGRATING.THE PUMP INTO THE SYSTEM

2 AVAILABILITY CONSIDERATIONS

3 CHOICE OF NUMBER OF PUMPS

4 CHOICE OF PUMP TYPE

4.1 Barske Type4.2 Peripheral Type4.3 Multistage Centrifugal Type4.4 Inlet Booster Pumps

5 DRIVERS

5.1 Steam Turbines5.2 Electric Motors5.3 Dual Drivers

6 DUTY

6.1 Differential Head6.2 Capacity6.3 NPSH


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7 MINIMUM FLOW THROUGH PUMP

7.1 Limit due to Instability in Head/Flow Characteristic7.2 Minimum Flow Arrangements7.3 Variable Speed Control

8 TRANSIENT EFFECTS

8.1 Reverse Flow upon Trip8.2 Flashing

9 TEMPERATURE GRADIENTS IN CASINGS

10 INLET STRAINERS AND CASING PIPING CONNECTIONS

10.1 Strainers10.2 Casing Piping Arrangements

11 SEAL COOLING

11.1 Packed Glands for Category 1 Pumps11.2 Mechanical Seals

SECTION TWO - PUMP CONSTRUCTION FEATURES


12.1 Limit Due to First Stage Cavitation Damage12.2 Limit Due to Instability in Head/Flow Characteristic

13 IMPELLER/DIFFUSER STAGE

13 .1 Local Cavitation13.2 Labyrinth or Bushing Clearances13.3 Clearances Affecting Hydraulic Forces


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14 ROTOR

14. 1 Rotor Dynamics14.2 Rotor Mechanical Balance14.3 Torsional Critical Speeds14.4 Rotor Assembly

15 GLAND ARRANGEMENTS

15.1 Guarding15.2 Soft-Packed Glands15.3 Mechanical Seals for Category 1 Pumps15.4 Mechanical Seals for Category 2 Pumps

16 AXIAL HYDRAULIC THRUST BALANCE METHODS

16. 1 Opposed Impeller Configuration16.2 Balance Disc16.3 Balance Piston

17 CASING17.1 Casing Type17.2 Casing Connections

18 MATERIALS

19 DRIVER CONSIDERATIONS


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BIBLIOGRAPHY

APPENDICES:

A NOTES ON BFW PUMP/DRIVER ARRANGEMENTSB PROPERTIES OF WATERC DEFINITION OF PUMP CATEGORYD TRANSIENT PHENOMENA IN DEAERATORSE NOTES ON PROPRIETARY LEAR-OFF VALVESF NOTES ON CASTINGS FOR HIGH-DUTY IMPELLERSG AREA RATIO METHODH TECHNICAL COMPARISON SHEETS

FIGURES

1 TYPICAL CROSS-SECTION OF PUMP IN CATEGORY 1

2A TYPICAL CROSS-SECTION OF PUMP IN CATEGORY 2 SHOWINGAXIALLY SPLIT CASING

2B TYPICAL CROSS-SECTION OF PUMP IN CATEGORY 2 SHOWINGBARREL CASING WITH RING TYPE CARTRIDGE

3 TYPICAL CROSS-SECTION OF PUMP IN CATEGORY 3 FORPOWER STATION APPLICATIONS WITH DRY RUNNINGCAPABILITIES

DOCUMENTS REFERRED TO IN THIS ENGINEERING DESIGN GUIDE


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1 SCOPE

This Engineering Design Guide covers the special requirements of pumps inBoiler Feedwater service.

SECTION ONE - INTEGRATING THE PUMP INTO THE SYSTEM

2 AVAILABILITY CONSIDERATIONS

The general reliability classifications are given in Engineering Design GuideGBHE-MAC-5101. Pump reliability is more difficult to obtain as the pressure ofthe system increases. Current practice is to distinguish three categories of thesystem by reference to the nominal pressure of the steam boiler drum.

Category 1 2 3

Steam pressure,bar abs.

43 43 125 160

Category 3 covers a model thermal power station and is beyond the scope of thisDesign Guide.

Consult Furnace Section about the consequences of BFW supply failure. ForCategory 2, BFW pumps are normally required to have 100% availability in orderto 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 pumprequires a steam turbine driver with the steam supply taken either from theboiler or from a secure source. See Appendix A.

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


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3 CHOICE OF NUMBER OF PUMPS

Arrangement 1

One 100% duty pump together with one standby pump. Current practice is tomake 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. Thisarrangement 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 isimportant when delivery lines are long or heat exchangers are includedwhich can quickly generate steam upon BFW flow stoppage.

(c) Increase overall power efficiency when the steaming rate is expected tobe less than 50% of plant rating for long periods, thus allowing operationwith one pump.

(d) Relax constraints on steam system control.


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Special Arrangements

The following arrangements should NOT be used without thorough systemsanalysis:

(a) Two running pumps each rated at 100% duty in order to guarantee BFWsupply continuity at 100% of rated flow.

(b) 'Minimum duty' standby pump sets, either rated at a low pressure and sorequiring rapid boiler depressuring before they can be started, or rated ata 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 andmay have an unstable Q-H characteristic. Nevertheless, where the steamdemand varies widely, with campaigns of low flow operation, consider the use of3 or more such pumps in parallel so that the number of running pumps can beadjusted to meet the BFW demand.

4.2 Peripheral Type

The peripheral type pump is well suited to BFW service for small capacitiesbecause its Q-H characteristic is inherently stable and because BFW issufficiently pure and free from suspended solids to make acceptable this pump'ssensitivity to erosion. Specify this type of pump in preference to reciprocatingpumps.

4.3 Multistage Centr ifugal Type

Most BFW pumps are of the horizontal shaft multistage centrifugal type. Thechoice of the number of stages is not a precise determination. Some guidance is

given in Clause Cl.4 of GBHE-EDG-MAC-1014.


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

As the number of stages increases the bearing span increases and the rotorbecomes more sensitive to dynamic effects. As a first estimate take the limitingnumber 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 thiscalculation.

Where the number of stages thus calculated is less than the number of stagesrequired 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 NPSHrequired by the main pump.

The conventional arrangement has the booster directly coupled to a doubleended electric motor which also drives the geared main pump.


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5 DRIVERS

5.1 Steam Turbines

The advantages of using different designs of main and standby turbines aresufficiently large to outweigh the reasons given in Clause 2 for selecting identicalmain and standby pumps.

In normal operation the failure rate of small high efficiency turbines is significantlyhigher than that for the conventional small turbine which has only 1 or 2 rows ofimpulse blading with large non-critical clearances. Consequently the preferreddriver arrangement is to have main pumps driven by electric motors or highefficiency turbines and the standby pump driven by a conventional turbineprovided with slowroll and quickstart facilities.

Standby turbines of normal construction are held on slowroll to ensure instantreadiness. For this purpose current practice is to leave the emergency steamstop valve open and the inlet autostart valve closed but bypassed by a smallrestrictor whose orifice size is empirically adjusted to give the required slowrollspeed. The autostart valve actuator is damped to give a valve stroking time of10 seconds (for a linear valve characteristic) in order to avoid speed overshoot.Check that the slowroll speed is above the hydrodynamic limit for journalbearings. 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 thesteam supply system upon starting. Such a steam demand can be supplied directfrom the boiler. However, in process plant these turbines are normally suppliedfrom an intermediate pressure header. Conventional control systems cope betterwith reductions than with increases in steam demand; for sensitive systems thepreferred driver for the standby pump is an electric motor. Then, current practiceis to specify Arrangement 2 with each main pump having a turbine driver.

For the emergency condition of electrical supply failure; specify the turbine driverto be capable of developing rated power when exhausting either to a securesteam system or to atmosphere via a pressure relief valve.


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5.2 Electric Motors

Current practice is to use only induction type motors arranged for direct-on-linestart.

Consult Electrical Section on the advisability of connecting motors to differentunit 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 consequenteffects on both pump design and layout. Standby pumps are preferably direct-drive. Gears that remain stationary need special consideration; the oil consoleshould then be provided with an electric motor driven auxiliary oil pumpcontinuously 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 ofthe normal driver is much lower than that of the pump. This case arises when thenormal driver is an electric motor connected to an unreliable electric supplysystem and the alternative driver is a steam turbine. Now the failure rate ofCategory 2 pumps is of the order of 0.5/year, consequently this arrangement isnot justifiable for most sites with access to the CEGB grid, when the supplyfailure rate is of the order of 0.1/year.

The dual drive arrangement may be used for power recovery, where the steamturbine is in continuous operation and the motor can act as an inductiongenerator exporting surplus energy as electrical power.

6 DUTY

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


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6.1 Differential Head

When calculating the required differential head at rated flow, with throttle controlvalves WIDE-OPEN, the delivery vessel pressure is taken as the set pressure ofthe boiler drum relief valves. This step does not guarantee that rated water flowcan be then delivered because relief valves require a further increase in pressurewhen passing their rated steam flow. The steep pump characteristic canaccommodate this pressure increment (of 5 - 10%) at some marginally lowerflow.

Pumps may require a correction to the simple head calculation; refer to AppendixB.

6.2 Capacity

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

6.3 NPSH

When calculating available NPSH note that some boilers have only singleelement level control: this demands either a generous NPSH allowance foracceleration head or a heavily damped throttle control valve, to cope withtransient boiler upsets.

BFW pumps may require unusually large values of NPSH. As speed or capacityincrease, keeping Sn constant maintains hydraulic performance but the intensityof local cavitation increases. There is a limiting value of NPSH for a givenmaterial and impeller construction. For typical castings in 13% Cr steels takethis 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 theoperating range 85 - 105% of the capacity at BEP; Snshould be further reducedif a wider range of operational capacity is required.


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


7.1 Limit Due to Instability in Head/Flow Characteristic

Typical pumps have Ns~ 0.06 for which the minimum flow to avoid instability is20% of the pump capacity at BEP. As Ns increases, this minimum flow limit alsoincreases 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 anda bypass valve are currently limited to Category 1 installations.

Current practice to ensure the minimum flow limit is to employ a simplecontinuous bypass through a let-down MULTIPLE restrictor back to the inletvessel. This continuous bypass bas been used for process purposes. Thecapacity of the pump is then increased to supply both the bypass and thedelivery flows.

Large category 2 pumps may merit a dedicated control system to maintainminimum flow by opening the bypass valve when the delivery flow falls below thespecified 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 deadbandzones of about 0.25% when specified to NEMA Class 'D'. For typical pumpand system characteristics, this limits the stable range of control on asingle pump to about 70% of BEP capacity; below this capacity ordinarythrottle valve control is used with the speed held constant.


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

(b) Fluid couplings provide significant excitation torques at discretefrequencies and have been identified as the cause of impeller failures.Variable speed electric motors of the Schrage or commutator typesimilarly provide excitation torques.

8 TRANSIENT EFFECTS

8.1 Reverse Flow upon Trip

Process heat exchangers in the delivery of BFW pumps constitute accumulatorscapable of supplying steam which may run the pumps in reverse rotation. Currentpractice is to provide a reliable non-return valve in each pump discharge line andNOT install reverse rotation locks. For' steaming' exchangers the risk is greaterand 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 steamflashing in the main pump where booster pumps supply BFW through feed-heaters giving water temperatures> 200C at the main pump inlet. The rate ofdeliberate speed adjustment should be limited to give a pressure reduction ratenot 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 thermalstratification in the casing and thence to casing distortion. Preventive measuresare:

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


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(b) The provision of a bypass round the discharge non-return valve in order toensure a sufficiently large reverse flow. The required flowrate depends onpump size and layout but lies in the range 2% to 10% of the pump ratedcapacity.

It is ESSENTIAL to maintain adequate thermal insulation around the pumpcasing.

A simple hole drilled through the non-return valve element suffers erosionand generates noise: the recommended arrangement is an external pipedbypass embodying a multiple restrictor which can be empirically adjustedduring commissioning.

This arrangement re-introduces the risk of reverse steam flow fromsteaming process heat exchangers; consequently a reverserotationdetector should be provided to trip shut the discharge block valve.

The distribution of water flows through the stationary pump isindeterminate unless the natural downward drift of cooled water isencouraged. Study the detail arrangement of the pump in conjunction withthe manufacturer to identify possible stagnant zones. Extraction of coolwater from such zones may entail the provision of an ancillary pump forreturn to the de-aerator. Select a glandless canned motor pump for thisduty to avoid problems of air ingress.

(c) Temperature gradients can be prevented by positively isolating the pumpto create a truly stagnant water condition. The pump then has to withstand'cold start' conditions of thermal shock. Both these requirements aredifficult to realize; consequently this scheme should be considered only asthe 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 meshstrainers are NOT recommended because they clog easily and bring pumpoperation to a stop.


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Pre-commissioning should include acid cleans and wash-through procedureswhich adequately remove scale from the -pump inlet system before the pump isfinally connected and started..

10.2 Casing Piping Arrangements

Separate inlet, bypass and balance water lines should be provided for eachpump. Manifolding each functional group of lines close to the de-aerator to avoidmultiple 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 andstandby systems share common elements.

(b) Pressure drop variations may be excessive; this is particularly true ofbalance 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 ofvalves accounts for a large percentage of cases of damage to BFW pumps.

It is important to avoid accumulation of steam bubbles and to ensure pumppriming: accordingly the rise in the discharge piping adjacent to the pump shouldbe continuous and exceed a slope of 1 in 40.

The line connecting the booster pump to the main pump inlet should be providedwith a bypass line and orifice restrictor directly returning to the supply vessel inorder to guarantee continuous venting. The flow should not be less than 10% ofthe booster pump BEP capacity.


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11 SEAL COOLING

11.1 Packed Glands for Category 1 Pumps

The gland enclosure drain should be piped to a safe disposal point where effluxof flash steam will not obscure operationally important items.

When the pump inlet water temperature exceeds 70C, cool flush water should beinjected 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 shallcool the flush water to a temperature not exceeding 65C with cooling water orambient 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, apressure limiting system should be provided, for example:


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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 isrun to drain in order to avoid seat distortion due to high or fluctuatingcooling water pressure.

(b) Specify mechanical seals for Category 2 pumps.

The seal circulates BFW through an external loop which includes a heatexchanger with cooling water as the cooling medium. This cooler ismounted above the seal in order to promote thermosyphon action duringpump standstill.

The standby BFW pump is normally required to run for a short period afterfailure of the cooling water supply. For an orderly shutdown it is sufficientto have an overhead reservoir capable of providing the required flow forabout 10 minutes.

SECTION TWO PUMP CONSTRUCTION FEATURES


12.1 Limit Due to First Stage Cavitation Damage

The usable capacity range of a pump narrows as S nincreases because flowrecirculation in the larger impeller eye results in increased cavitation damagepotential a flows away from the stage best efficiency point.

When Sn(NPSH p)1/4> 0.8 check manufacturer's offer by comparing available

NPSH against curve of 'onset cavitation NPSH' for the recommended minimumflow point.

12.2 Limit Due to Instability in Head/Flow Characteristic

Manufacturers find it difficult to shape the pump characteristics: experienceindicates that all tenders should be thoroughly checked to verify that the stabilityrequirement has been satisfied.


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Note that the classic design methods based on impeller vane angle as thefundamental parameter do not afford the insight given by the area ratio designmethod - 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 metalsurface. It is not related to the inlet NPSH.

Peripheral and Barske type stages have open impellers which can be fullymachined and which can be used at heads up to 1500 m when manufactured in18/8 austenitic steels.

Centrifugal pump stages with closed impellers and bladed diffusers of optimumdesign for high efficiency are manufactured as castings. Errors occur when usingtraditional pattern-making and casting techniques. A convenient classification isby differential head h across one stage, as follows:

(a) h < 40 m

No special care needed.

(b) 40 m


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(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 beforbidden. After adjustment by machining the vane tips (but not theshrouds) to smaller diameter, the upper surface of the vane should besmoothly faired. The trailing edge angle is limited to 0.2 radians.

Impellers should not be balanced by local removal of material from thedisc/shroud surfaces. Such surfaces should run true.

(e) h > 660 m

Stages above 660 m head need thorough investigation of the design andmanufacturing 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 theworks performance test. Such small clearances quickly wear to some equilibriumvalue; the consequent rubbing increases the risks of rotor seizure or inducedwhirl.

(a) Clearances should not be less than the values given in the following tablebased on BS 4500.


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(b) Flexible rotors sag appreciably at rest but approach the nominal geometriccenterline during operation. Eccentric clearance settings are NOTacceptable; bearing and casing centerlines should coincide.

Bushing wear-rate can be dramatically increased when the pump is run atmuch reduced speed, eg during slowroll. Then either:

(1) the shaft sag should be less than the minimum radial clearanceobtained after allowing the assembly tolerances.

OR

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


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13.3 Clearances Affecting Hydraul ic Forces

Most multistage centrifugal BFW pumps have vaned diffusers. The radial gapbetween the impeller vane tips and the diffuser vane leading edges affects thepressure pulsations at the blade passing frequency, the radial hydraulic forcesand t he broadband noise level.

Let C be the radial gap, R the impeller vane tip radius and E the radialeccentricity of the shaft centre to the diffuser geometric centre.

Then C E > 0.02R

Irregularities in castings may cause variations in the measured gap C. Theinspecting engineer should verify that these measurements fall within the band of100% to 140% of the least value found for C. Any machining of the leading edgesof the diffuser vanes should be followed by hand finishing to produce the roundedentry profile.

Pumps with double volutes should be subject to inspection to verify that both theradius to the tip and the leading edge profile of each volute cutwater is identical.

Then, using the same nomenclature

C - E > 0.04R

Impellers should have an ODD number of vanes to reduce perturbing torques.


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14 ROTOR

14.1 Rotor Dynamics

Multistage centrifugal pumps in Categories 1 and 2 normally have operatingspeeds above the first lateral critical speed obtained by assuming the pump runon air; they rely on the damping provided by the pumped water to suppressdynamic effects. Unwanted perturbations chiefly arise from operating at partcapacity, producing large low-frequency radial hydraulic forces.

Category I pumps should have rolling element bearings, within the limits set bySpecification GBHE-MAC-18-06A.

Future systems using feed heaters may require the main BFW pump to have astiff shaft capable of running dry for short periods.


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Current practice is to supply the BFW pump directly from the de-aerator andpermit the use of flexible shaft pumps which cannot withstand running dry. Thepump shaft diameter and bearing span may be varied by pump designers for agiven hydraulic duty, rotor speed and number of stages. Guidance on acceptablerelationships of these parameters is given in Fig 14.1 which is based uponavoidance of excessive shaft bush wear. Note that such wear has led to shaftfracture in Zone A.

Category 2 pumps should have the second critical speed of the rotor, as thenatural frequency in air, not less than 130% of the maximum continuous speed.This frequency is easily measured and is a useful inspection check onmanufacture.


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FIG 14.1 HORIZONTAL MULTISTAGE PUMPS WITH RADIAL FACEMECHANICAL SEALS


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NOTES ON FIG 14.1

1 At each impeller location the shaft may be grooved for a retaining ring orbe stepped in diameter. Now the effective diameter should correspond tothe shaft response acting as a beam in bending; for this it is sufficientlyaccurate 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 bymeasurements on the assembled rotor. The expression

Z X d5Z + 2 L3

Is then replaced by:

k. w. D

as a close approximation where:

W = weight of rotor (excluding coupling spacer) kg

. = measured deflection of rotor at midspan mmD = diameter of shaft/casing bushes mmk = constant ~including g at 9.81 m/s

2) of value

8.6 x 10-6

14.2 Rotor Mechanical Balance

Each rotor element should be individually balanced before assembly on to theshaft. Dummy half-keys are necessary to ensure correct balancing. For theseelements the balance quality according to BS 5265 should be better than G6.3and 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 sleeveslocked by end shaft nuts. Such rotors should NOT be balanced after pre-assembly but checked as Clause 14.4.


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Rotors for horizontal split casing pumps are fully assembled and have eachimpeller individually located. The check balance of the assembled rotor shouldshow 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 torsionalcritical frequency of the pump/driver system, calculated for the rotor running in airshall be within 20% of the vane pass frequencies. This margin can be reduced to6% when the critical frequencies have been measured when running in water.These margins include the normal variation in speed of induction electric motordrivers and an allowance of 0.6% for variation in to. frequency of the motorelectric supply. The speed regulation performance of steam turbine governorsshould be individually assessed.

14.4 Rotor Assembly

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


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CATEGORY 1 PUMPS WITH RING SECTION CASINGS

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

CATEGORY 2 PUMPS WITH HORIZONTAL SPLIT CASINGS

The impellers are individually mounted on a stepped shaft so that the bores ofthe impellers progressively increase. This provision demands careful numberingof impellers.

High speed rotors have the impeller bore machining limits adjusted to retain atransition fit when rotating at the maximum continuous speed. This requires aninterference fit upon assembly.

It is essential that impellers do not swash, i.e. when the rotor is assembled andsupported on 'V'-blocks, a dial indicator bearing axially against the machinedimpeller shroud surface should show no variation in reading upon rotor rotation.

CATEGORY 2 PUMPS HAVING BARREL CASINGS WITH RINGSECTION CARTRIDGES

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

CATEGORY 2 PUMPS HAVING BARREL CASINGS WITH AXIALLYSPLIT CARTRIDGES

Treat as pumps having horizontal-split casings.


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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 andstability. Check that:

(a) Packing is one of the standard metric sizes of 6, 8, 10, 12.5, 15 mmsquare section.

Current material used is Walkers Fortune 417.

Shaft sleeves are provided, of 13% chromium steel preferably BS 970 420S 45 hardened to 240 - 280 VH and ground to a surface finish better than0.4 mRa. It is essential that the sleeve surface is concentric with theshaft: this should be verified at the time when the rotor is inspected forstraightness.

(b) When the pump inlet water temperature exceeds 70C, cool flush water isinjected through a lantern ring.

15.3 Mechanical Seals for Category 1 Pumps

Category 1 pumps may be fitted with mechanical seals having water cooledseats in order to avoid the need for a cooled BFW flush supply to a packedgland.

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-lockedzones.


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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 pumpbalance 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 sealdiameter) exceeds 20 m/s.

The seal circulates BFW through an external loop which includes a separator anda heat exchanger with CW as the cooling medium. This cooler is mounted abovethe seal in order to promote thermosyphon action.

Particles of iron oxide are prevented from reaching the seal faces by a magneticparticle separator.


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16 AXIAL HYDRAULIC THRUST BALANCE METHODS

16.1 Opposed Impeller Configuration

This demands a complicated casing shape; a specific assurance on castingquality control should be obtained from the foundry nominated by the pumpmanufacturer.

A thrust bearing is required to take the residual thrust: to ensure that this thrust isunidirectional 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. Thebalancing 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 theclearance in the preceding piston section. Efforts to increase efficiency bydecreasing the leakage flow through a reduction of ~ should be resisted,otherwise the pump becomes very sensitive to particulate contamination of thefeedwater.

Rubbing occurs during transients. SULFINUZ anti-galling surface treatment hasbeen used successfully to lengthen disc life but the performance was erratic.Current practice is to use hardened chromium steels.


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The dynamic compliance of the balance disc gives a natural frequency of axialvibration of the rotor. Interaction may occur with metal diaphragm type shaftcouplings whose spacers have their own intrinsic natural frequency bands ofaxial vibration. Such an effect should be specifically assessed for pumps runningabove 65 rls.

For truly parallel balance disc faces the restoring force rapidly decreases as thefaces come nearly into contact. This effect is reduced by extending the face reliefnearly 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 beproportioned to ensure that this residual thrust is unidirectional over the normaloperating range.

There is a hydraulic self-centering action, but this is reduced if the piston isserrated.


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Current practice is to ensure that:

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

(b) the assembly positioning aligns the bush along the geometric centerlinebetween bearings.

17 CASING

17 .1 Casing Type

CATEGORY 1 PUMPS

These are invariably of the ring section type having one cell per stage, the wholebeing 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 theextension 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 thanM20. Check that the Manufacturer's Manual includes specific instructions on thesequence and method of tightening the through bolts.

This type of casing is prone to thermal distortion because the through bolts arenot in contact with the feedwater and the uniformity of bolt temperature dependson the quality of the lagging application. Consequently this type should be usedonly where the pump draws feedwater directly from the de-aerator vessel and thewater 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 pumpshave been used in double pump arrangements, both pumps being driven by adouble-ended electric motor, but connected in series for BFW flow.


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CATEGORY 2 PUMPS - NOMINAL SYSTEM PRESSURE 43 to 100 BAR

The horizontal-split casing type enables the rotor to be assembled as a unit priorto insertion into the casing.

Volute stages are often used to simplify the casting; it is then important that thetwo half-casings match at the split-line. The inspecting engineer should visuallyexamine the assembled casing to verify that the volute surfaces have beenground 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 casingand 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 ring-section pump but when the limits for rotor dynamics are approached the cartridgeassembly in an axial split arrangement should be considered.

17.2 Casing Connections

(a) Configuration

Horizontal-split casings have side connections in order to retain theadvantage of removing the upper half casing without disturbing the inlet ordischarge piping.

For ring section and barrel casings, specify top discharge and top or sideinlet connections.

Maintenance of barrel casing pumps is carried out by withdrawing thepump cartridge, leaving the casing installed. Consequently the highpressure discharge branch may be butt-welded to the piping.

(b) Rating

For Category 1 pumps specify that the inlet and discharge connections arerated for the same pressure.


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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 pumpand 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 disturbingthe pipe system.

Horizontal-split casings are torsionally very weak when the upper half-casing hasbeen removed; consequently the applied moments should be zero when theassociated piping is cold. Note that this condition occurs not only when allpumps 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 goodthermal and mechanical properties and immunity from corrosion bydemineralized feedwater.

Because the feedwater is de-aerated, castings for horizontal-split casings mayhave the Chromium content reduced to the lower limit of 5%. High qualitycomplex castings have become progressively more difficult to obtain.Consequently for horizontal-split casings, the inspection plan and weld repairtechniques should be agreed with the foundry at the time of placing the purchaseorder.

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 shouldNOT be used for shafts.


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For rubbing pairs, viz. impeller/casing wear rings and shaft sleeves/bushings, it isadvantageous for the stationary component to have the highest practicablehardness. Current practice is to specify casing wear rings in PH stainless steelaged to produce a hardness of 300-340 HV and ground to a surface finish betterthan 1.6 m R

aupon visual inspection against a comparator tablet. Impeller wear

rings are vulnerable to disturbance on heavy rubs; consequently the preferredarrangement for the highest intrinsic reliability is to integrate the wear ring intothe impeller casting. This favors the use of 13/4 Cr Ni alloy heat-treated to obtaina hardness of 240-280 VH, also ground to a surface finish of 1.6 m R a. Such arubbing 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 beused.

CATEGORY 1 PUMPS

The important feature is that steam systems in Category 1 normally do notemploy demineralized feedwater and the pump inlet temperature is less than 120C.

Early practice using bronze impellers and fittings has been discontinued becauseof erratic deterioration of bronze components and the high thermal expansion ofthis material.

Individual cells of ring section pumps may be in grey cast iron to BS 1452 Grade14 or 17 but specify theend-sections in carbon steel or nodular cast iron fordesign pressures above 16 bar g.

Impellers and diffuser rings may be in grey cast iron to BS 1452 Grade 14 or 17provided that H < 40 m, except for the first stage which should be in 13/4 CrNisteel when Sn >0.28.

Casing wear rings are normally Ni-resist. For demineralized feedwater follow thepractice for Category 2 pumps.


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19 DRIVER CONSIDERATIONS

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


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BIBLIOGRAPHY

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

Paper 1 A feed pump design concept for 660-MW generating sets, byJ 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 unitrating of their turbo generators from 500 MW to 660 MW, thedesign basis of the boiler feed pumps was completely reviewed.The prime requirement was that the pumps should be made lesssensitive to mechanical fault and capable of surviving disturbedsuction conditions without failure. At the same time they were to becapable of rapid replacement. This paper explains the decisionthat the pumps should have only two or three stages, with stiffshafts. 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, andM 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 powerstations, by A A Gasiunas, Dipl.Ing


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The paper discusses the pressure pulses that are generated whenimpeller vanes travel past stationary vanes. The influence ofgeometrical and hydraulic non-symmetry on radial forces in adouble volute casing is dealt with. Test results of a double suctionmodel boiler feed pump are presented, including descriptions ofdurability 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 BFWat 160C including measurements on through bolt temperaturedifferences and casing distortion.

Paper 7 The single-stage high-speed nuclear feed pump, by A R Bush, BMech.Eng, and J E C Valentin

Feed pumps applied to early large-size light water moderatednuclear reactors in the United States were modifications of existingsplit-casing volute-type pipeline pumps. Reliability was the keyconsideration. The final selection was a single-stage solid casingwith an integral diffuser and condensate injection seals. Themanufacturing problems are discussed, together with performancetests run on the first machines of this type.

Paper 8 Application of the thermodynamic method of measurement for thedetermination of the boiler feed pump efficiency in large electricalpower units of Electricite de France, by J S A Guitton, IngenieurEcole Centrale, and H Procaccia, Ingenieur EEIP

Paper 9 Dynamic hybrid bearing characteristics of annular controlledleakage seals, by H F Black, MSc., C Eng, MI Mech E, and D NJenssen, BSc.

The dynamic bearing characteristics of plain seals havingappreciable length-to-diameter ratio are analyzed. Someexperimental results are given and compared with theoreticalpredictions. It is shown that length-to diameter ratio has a

considerable effect on seals of appreciable length, e.g. balancepistons, and that dynamic components of bearing action due toshaft rotation are comparable with the purely hydrostatic centeringforces.


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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 actupon the rotor of a centrifugal pump with diffuser guide vanes if theimpeller is out of centre. The magnitude of these forces increasewith growing eccentricity. At very small flows, these forces becomenon-stationary. They rotate at a considerably lower speed than therotor speed. The paper discusses the effects and resulting shaftdeflections and vibrations, from measurements On twoexperimental 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 duringthe last decade by G F Arkless. COny on Steam Plant Ancill1aryEquipment Proc lnst Mech Eng 1966/7 181 (Pt.3N) 6.

4 Analysis of cavitation damage in commercial marine condensatepumps. PE Paashaus ASME - SNAME Meeting New York December1964.

This paper refers to the "bulging" of impeller shrouds upon "hammering"when operating on free suction control.

5 Horenburg, 0: Schaden an Kesselspeisepumpen. Hinweis zurSchadenverhutung durch Auswertung von Schadenstatisiken. DerMaschinenschaden 43 (1970), No 4, pp 135/147.

Describes pump damage due to thermal distortion of casings and ingressof foreign bodies through strainers.

6A Honold, E : Vergleich von Kesselspeisepumpen fur hohe Enddrucke inGlieder-und Topfbauweise im Hinblick auf die Entwicklung groberKraftwerksblocke, Mitt der VGB (June 1966). No 102, pp 149/152.


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6B Honold, E : Rin kritischer Vergleich von Topfund Gliederpumpe amBeispiel der Entwicklung der Hochstdruckkesselspeisepumpen inDeutschland und den USA. KSB Technische Berichte, No 1 (October1960).

These 2 papers discuss the choice of barrel or ring-section pump casings.

7A Strub, RA : Abfall des Saugdruckes von Speisewasserpumpen beistarken Lastschwankungen. Technische Rundschau Sulzer 42 (1960),No 3, pp 41/44.

7B Wollschlaeger, K : MaBnalmen zur Verbesserung der zulassigenDruckabsenkungsgeschwindigkeit bei Kesselspeisepumpen.Energie 18 (1966), No I, pp 16/18.

7C Rahlwes, H : Untersuchung zur Klarung Speispumpen beiGleitdruckentgasung. No 106, pp 61/67. von Zulaufstorungen anMitt VGB (Feb 1967),

7D Stonner, A : Ein Beitrag zur Schadenverhutung anHochdruckkesselspeisepumpen. 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 inCentrifugal Pumps. H F Black & E A Cochrane, Sixth InternationalConferennce on Fluid Sealing 1973, 65, pp 61 - 69.

9 W Schumacher

Wear and Galling of Nitrogen-Strengthened Stainless SteelsReport in Machine Design August 1983 on paper presented at Wear ofMaterials 1983 Conference, Reston, V.a. Apr 1983. Indicates that nitrogenaddition improves strength and hardness of stainless steels but does notimprove wear resistance.

Increasing nickel content decreases the wear resistance. Best materialNITRONIC 60.


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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 bysteam turbines. The 1920's saw the introduction of the arrangement with themain 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% capacitystarting/standby pump having electric motor drive, appeared about 1955. Avariation of this arrangement used a main pump driven through a hydrauliccoupling 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 ratesand pressures which are adequate to take care of any emergency.

C5.403 A spare feedwater pump or injector, in addition to the feedwaterequipment required by PG-61.I, Section I, is preferable. Wherefeedwater pumps are electrically driven and there is no fullyindependent auxiliary source of electric supply, there shall bemaintained, ready for service, steam-driven feed pumps or injectorsof sufficient capacity to safeguard stoker-fired boilers in case offailure of electric power. This recommendation also applies toboilers fired by other methods if furnaces contain large amounts ofrefractory or are arranged to accumulate slag in the bottom.

C6.803 Normal Operation

A connection should be provided between the pump discharge lineand a point on the suction system, preferably the deaerator, as ameans of preventing overheating of the pump when it is required tooperate 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 andthe check valve.


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The recirculating line should be connected back into the suctionsystem at some point where the heat added by the pump will bedissipated before the water can re-enter the pump suction. This lineshould be provided with a stop valve and an orifice of such size asto permit recirculation at the minimum flow rate stipulated by themanufacturer of the pump.

The stop valve should be locked or sealed in the wide-openposition whenever the pump is in operation or ready for operation.

All valves in hydraulic-balancing drum-leak-off lines should belocked or sealed in the wide-open position whenever the pump isready for or in operation.

3 EXTRACT FROM PG-61 (GENERAL REQUIREMENTS FEEDWATERSUPPLY)

61.1 Except as provided for in PG-61.2 and PG-61.4, boilers having more than500 sq.ft. (47 m2) of water heating surface shall have at least two meansof feeding water. Except as provided for in PG-61.3, PG-61.4 and PG-61.5, each source of feeding shall be capable of supplying water to theboile

Boiler Feed Water Pumps

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