0262 1762/06 2006 Elsevier Ltd. All rights reserved WORLD PUMPS January 200628

In light of the deregulation of theUK electricity industry, powerstations needed to becomeflexible in terms of generating loadand operating times, with numerousstarts/stops and part-load operationbecoming the norm. At FiddlersFerry, a 2000 MW coal-fired powerstation commissioned in the 1960s,the fundamental design of a 12 000kW boiler feed pump (BFP) and agingassociated ancillary equipment led tosignificant reliability problems whenfaced with these new operatingdemands. Moreover, the powerstation had been given a finite lifeand the associated reduction inmaintenance spend had furtheraggravated the situation. However,with the increasing cost of gas andrelatively low price of coal, olderfossil fuel fired power stations becamemore economical to run.

Faced with this situation, Flowserveworked with the customer tounderstand the root causes of thepump and system problems. Anupgrade solution was proposed, usinglife cycle costs (LCC) to calculate thereduced maintenance and operating

costs on a return-on-investment(ROI) basis. The increasedreliability/availability is expected to significantly increase the pumpsmean time between overhauls(MTBO) to 70 000 hours (eightyears). The first of the upgradedpumps is now installed, operatingvery satisfactorily and assisting withthe generation of 25 MW moreelectricity than with the original BFP.

The initial scenario

Fiddlers Ferry power station consistsof four units generating 500 MW,each containing a Mather & Platt16-16 Plurovane main boiler feed pump (MBFP) directly driven by a steam turbine, togetherwith two start & standby boiler feed pumps (SSBFP) driven byvariable speed motors. Up to the1990s, the station generated at fullload power totalling 2000 MW, as perthe original design. During this period the MBFPs were operatingalmost nonstop. This optimumcondition resulted in a reasonablereliability of the pumps.

With deregulation of the electricityindustry, and the subsequentincreased competitiveness of themarket, the station had to adopt amore-flexible operating regime,generating only at times wheneconomically viable. Two-shiftingbecame the normal mode ofoperation with the pumps onlyoperating at peak demand timesbetween low and high pump loads.The result was that the pump unitswere being started and stopped anever-increasing number of times.This had a significant effect on thereliability of the boiler feed pumps.

Observation of the cartridges duringoverhauls revealed a number offailure modes. The pumps wereusually stopped because theirefficiency had deteriorated so muchthat their performance did not meetthe demands of the boiler. The performance degradation was aresult of high internal wear. In addition, the first stage impellersuffered from severe cavitation andrecirculation damage, often resultingin an out-of-balance force thatincreased the vibration levels of thepump above alarm-trip levels. The maintenance spend on thepumps was far above average and,with the foreseen requirements onthe pump, the situation would only get worse.

Flowserve had worked with thispower station previously, looking atthe reliability of the SSBFP, whichhad an MTBO of around 600 hours.As such, the company was familiar

Upgrading boiler feed pumps in aUK coal-fired power station

The reduction in running costs following a carefully considered pump modification caneffectively pay back the initial expense, as in the case of the upgrade of a series of boilerfeed pumps in a UK 2000 MW coal-fired power station where life cycle costs were used tojustify capital investment. Peter Irlam, an aftermarket technical services manager atFlowserve Pump Division, provides the details.

f e a t u r e r e d u c i n g r u n n i n g c o s t s

Figure 1. Appliedapproach to problemsolving.

WORLD PUMPS January 2006 www.worldpumps.com 29

with the designs of the MBFPs andproposed the course of actionoutlined in Figure 1.

History and fieldmeasurements

The history and field measurementswere collated by a week-long sitesurvey carried out by a team of fivepump improvement engineers. The agreed agenda with the customerwas to:

Review MBFP current operatingprocedures

Determine the pump requirements(demands)

Review historical files detailingpump failures (internal andexternal)

Obtain system and machineinformation

Perform customer archive review Interview maintenance,

operations and engineeringpersonnel

Measure hydraulic performanceand dynamic behaviour of theMBFP

Obtain historical MBFP main-tenance costs.

On completion of the site survey, adetailed engineering report wasproduced for the customer indicatingthe areas of the pump design andoperation contributing to the poorreliability. Some major observationswere made.

From 1998 to 2003 the number ofstart/stops on the MBFP increasedfrom an average of 100 per unit to239 per year (Figure 2).

Before 1993, when the stationintroduced two-shifting operation,the average MTBO was 44 months.After 1993, over a period of ten years,the average cartridge MTBO was 21 months.

Of the 16 cartridge changes between1998 and 2003 (Figure 2), seven were for vibration, five as a result of pump seizure, two through

wear/erosion causing loss of pumpperformance, and only one for aplanned overhaul.

The average MTBO for unit 4 was 1.5 are required. It would not bepossible to increase the NPSHA;instead, the NPSHR would have tobe reduced to improve the marginratio.

From the impeller hydraulic designinformation, the onset of both thesuction and discharge recirculation(Figure 4) was calculated using themethods of Fraser1. The dischargerecirculation was 15001550 m3/h,less than the best efficiency point(BEP) of the pump. Also, the vaneoverlap was very small, a symptomthat contributes to a strong suctioneye recirculation. Both recirculationspresent will result in pumpperformance degradation with anassociated loss in efficiency.

The hydraulic review had pinpointedthe reasons behind the damage to the

f e a t u r e r e d u c i n g r u n n i n g c o s t s

Figure 2. (a) Historical main boiler feed pump (MBFP) cartridges changes; (b) MBFP starts per unit per year.

Figure 3. Typical recirculation damage tothe first stage impeller vanes.

(a) (b)

WORLD PUMPS January 200630

first stage impeller. Recirculation wasoccurring during normal pumpoperation leading to extensivedamage. This operating conditioncaused loss of impeller material,leading to the rotor going out

of balance. Continued operationresulted in increasing pump vibrationlevels.

Many operating parameters wererecorded during the site survey.Vibration spectra taken on thebearing housings showed the erosion-induced out-of-balance signal butalso a more-dominant peakcoincident with the vane passingfrequency (VPF). There were sixvanes in the impellers, leading to anever-present vibration signature atsix times the running speed. Thisvibration signature was caused by theinteraction with the five vanes in thediffuser. It is well known thatimpeller/diffuser vane combinationshaving a numerical difference of onemaximize VPF vibration. Furtherdetailed analysis using in-houseproprietary software programs PumpDoctor and Pulsatr predictedsevere pulsations, with rotorvibrations at 6x running speed (see Figure 5).

Data B represents the magnitude anddirection of the resultant radial forcefor 60 of rotor rotation, caused bythe hydraulic interaction betweenthe impellers and diffusers [NB: DataA does not apply to this analysis].The orbit of force rotates 360 forevery 60 of rotor movement. Hence,for one revolution of the rotor theresultant force rotates six times,equating to VPF as measured. Notealso that the magnitude of the forcefrom the origin equates to about ten.

The same software programs wereused to predict the vibrationbehaviour of a proposed hydraulicdesign with seven impeller vanes and

nine diffuser vanes. Pump Doctorpredicted some pulsations, rotorvibrations at 28x running speed(Figure 6). Data B representing theresultant radial load no longer rotatesbut has a random pattern. Inaddition, the magnitude of the forceis of the order of one.

Hydraulicperformance

Investigation of the headflowoperating point of the pump showedthe pump performance curve was notmatched to the actual operatingconditions. The BEP of the existinghydraulics was too high for the actual100% operating point of the pump.As a consequence, the pump wasoperating below peak efficiency, asituation made even worse whenoperating at part load.

From the operating data review, theunits were operating for 50% of thetime at 100% block load (489 MW),and 50% of the time at 60% blockload (292 MW).

Once the operating scenario wasunderstood, the hydraulic designcould be tailored to best meet theactual demands of the power station,i.e. providing the highest possibleefficiency at the power plants futureoperating conditions (Figure 7).

Energy savingopportunities/LCC

The power station feed-water systemwas reviewed for possible energysavings using actual operating data,and various opportunities wereidentified.

Efficiencyimprovements on theMBFP

The MBFP optimized for efficiencywith the operating scenario between100% and 60% block load, combined with new impeller/diffuserdesign, gives higher peak efficiency

f e a t u r e r e d u c i n g r u n n i n g c o s t s

Figure 5. (a) Original impeller/diffuser design. (b) Predicted radialforce (Data B) for six impeller vane/five diffuser vane combination.

Figure 4. (a) Impeller suction eye recirculation; (b) discharge recirculation.

Figure 6. (a) Proposed design with a seven-vane impeller and a nine-vane diffuser. (b) Predicted radial force for the proposedimpeller/diffuser vane combination.

(a) (b)

WORLD PUMPS January 2006 www.worldpumps.com 31

(+6%) than the original pump (as explained above), whichgenerates large energy savings in the drive power needed from the MBFP steam turbine. This savingamounts to 420 kW which, with athermal efficiency of 36%, is equal to1167 kW of boiler heat. The MBFP inunit 4 had additional performancedegradation due to head reductionresulting from suction recirculation,hence the damage observed on unit 4impeller vanes (Figure 3). Themeasured head degradation on theunit 4 MBFP was 184 m of head,resulting in 821 kW of extra drivepower required from the steamturbine, which represents 2200 kW ofboiler heat.

Finally, steam tapped from the outletof the high-pressure (HP) stage of themain turbine reached the auxiliarysteam turbine outlet at too low apressure for it to flow back to thedeaerator after passing through theHP heaters. This required the steamto flow back to the condenser,resulting in a very poorthermodynamic efficiency. Thepower plant operated with anadditional steam tap from the mainturbine IP (intermediate pressure)stage fed into the last stage of theauxiliary steam turbine to increasethe back pressure of this steamturbine, allowing the outlet steam tobe returned to the deaerator, which ismounted 30 m higher, rather than allthe steam flowing to the condenser.

Upgrading the MBFP hydraulics togenerate a much lower drive powerfor the steam turbine could possiblyeliminate the turbine back pressureproblem, resulting in a large gain inthermodynamic efficiency.

Reduced regulatorvalve throttling

The pressure drop across thefeedwater regulator valve over thepump operating range is in excess of 10 bar. For Benson and slidingdrum type boilers this is normally

2.5 to 3 bar. This pressure droppresents an opportunity to reducevalve throttling and save energy as aresult of the lower MBFP drive power.Once the MBFPs are retrofitted withmore reliable cartridges, the entireboiler regulation can be done byspeed regulation rather than by valvethrottling combined with speedregulation.

A 10 bar reduced valve throttling at100% block load equates to a pumppower reduction of 648 kW. With aplant thermal efficiency of 36% thisrepresents 1800 kW of boiler heat.

HP injection, superheated, tapped fromMBFP

Currently, the superheated HPinjection water is tapped from theboiler economizer manifold and allinjection water going intoattemperation is heated in the high-pressure heaters. This HP injectioncould be tapped directly from theMBFP, if equipped with a specialattemperation stage impeller,certainly when energy savings byreduced valve throttling areimplemented. By increasing only 5%of the injection flow to the required

f e a t u r e r e d u c i n g r u n n i n g c o s t s

Figure 8. Thermodynamic analysis of steam cycle (where HP, IP and LP = high, intermediate and low pressure,respectively; HP5..HP7 are high-pressure feedwater heaters).

Figure 7. Comparison of existing over designed hydraulics versus new optimizedhydraulics.

Model:CasePower:HR:EFF:

FF10FF10OD504.629438.1738.14

Condensatepump

Condenser

LP heaters

Deaerator

MBFPBFPTHP5HP6HP7

Fossilboiler

HP turbine

IP turbineLP turbine

WORLD PUMPS January 200632

injection pressure, rather than all thepump mass flow, the drive power of theMBFP can be reduced for 95% of theMBFP mass flow, which is insteadincreased to the lower boiler inletpressure rather than the injectionpressure. This lower drive power for theMBFP represents 3194 kW of boilerheat savings.

The energy savings, all expressed in boiler heat, were calculated using an average cost of 2 per gigajoule (GJ) of boiler heat and 8000 hof operation per year in the 60% to 100% block load range. A combination of reduced boilerregulating valve throttling, attemp-eration injection and using a small kicker stage impeller on theMBFP, together with optimizedhydraulics, would result in boiler fuel savings of 370 000 on units 1, 2 and 3 and 505 000 for unit 4.

Flowserve conducted a full thermo-dynamic analysis of the steam watercircuit, using the commerciallyavailable Gate cycle software toanalyse the existing and modifiedconditions (Figure 8).

The customer chose not to invest inthe energy saving opportunitiesassociated with HP injection and theregulating valve. As such the LCCprojection was based on theinstallation of a new, more-efficientcartridge. The LCC model is based on:

A one-off engineering study in 2003

Installation of one upgradedMBFP per year starting in 2004

Eight years MTBO for upgradedMBFP

LCC does not include the extra 25MW generating capacity owing to the installed MBFP unit.

Figure 9 charts the summation () ofannual costs from the start of theproject to the stations expected life,where:

upgrade costs = cost of the pumpupgrade including, MBFP, SSBFP,technician, savings

yearly costs no improve =projection of annual costs to runstation with no improvements

pay back = difference betweenprojected yearly costs withoutimprovements and upgrade costs.

With the extra generating capacity of the upgraded MBFP, the pay back period will be well before the projection of 2008. From that point, substantial savings are made on an annual basis.

Original pumpdesign critique

A design review of the existing pump(Figure 10) revealed a number of

problem areas that were improved in the upgraded design (Figure 11) as follows:

Single suction first stage impeller.Suction specific speed (Nss) >14 000.This is very high for a BFP and shouldbe closer to 9500. The pump requiresa double suction first stage impelleroptimized for minimum NPSHR andreduced suction and dischargerecirculation.

Hydraulics too large. New pumpperformance is based on actualoperating conditions with BEPoptimized for 100% and part loadoperation, using proven highefficiency hydraulics tailored to fit inthe existing barrel.

Discharge head too thin. Finiteelement analysis showed deflection at the gasket area leading to steam leaks. The new discharge head is designed to ASME VIII pressurevessel calculations.

Discharge head to barrel fit. Openfit sealed by Grayloc gasket; leads to assembly issues includingmisaligned cartridges. Upgradeddesign to have metal-to-metal fitwith controlled compression gasketallowing proper stud pre-load andtorque sequence.

Balance disk design. Thiscounteracts the entire residual axial load generated within the pump and negates therequirement for a thrust bearing.However, the disks operate with a very close axial clearancemaking them unsuitable for fluidscontaining any particles or for pumps with varying operatingconditions, without a spring-loaded thrust bearing arrangement. The upgraded pump is fitted with a balance drum, which operates with a higher clearance and can tolerate changes in pump operating conditions. Thisdesign change requires the use of an associated thrust bearing toaccommodate the residual axialthrust.

f e a t u r e r e d u c i n g r u n n i n g c o s t s

Figure 9. Life cyclecosts for MBFPupgrades(1 = c. 1.48).

WORLD PUMPS January 2006 www.worldpumps.com 33

Coupling. Existing coupling is a veryheavy lubricated gear type designcausing an overhung mode predictedby the rotor dynamic analysis.Replaced with a lightweight disccoupling requiring no lubrication.

Existing condition monitoringequipment limited to accelerometerson bearing housings. New cartridge tobe additionally fitted with non-contacting radial and axial shaft probeshardwired to the on-line monitoringsystem. Embedded temperature probesare fitted to the journal together withactive and inactive thrust bearing pads.

Existing wear part materials improvedto PTA (plasma transferred arc) weldoverlay rotating wear parts and HVOF (high velocity oxy fuel)overlaid stationary wear parts. Thiscombination offers the best anti-gallingproperties whilst maintaining thepump performance.

Existing 180 bearing housingsreplaced by 360 bearing housings.These offer greater rigidity, main-taining pump alignment at all times.

Conclusion

This whole process of installation ofthe upgraded MBFP was mapped outto the customer at the beginning ofthe investigation. Each of the steps ofthe problem solving approach out-lined in Figure 1 has been completedand reported to the customer. Byworking closely with the customer, adetailed LCC model was producedaccurately reflecting the future costswith and without upgraded cartridges.This cost justification together withsuccess stories of working together in the past, made for an attractiveproposal to the customer. The firstupgraded unit was installed in August 2004, naturally not withoutteething issues. However, these werequickly resolved with the co-operation between Flowserve and thecustomer. The unit has been runningsuccessfully since then with vibrationlevels of 0.8 mm/s as measured on thebearing housings.

Reference

[1] W.H. Fraser, AvoidingRecirculation in Centrifugal Pumps,Machine Design, (1982).

Acknowledgement

The author thanks the manage-ment and engineers in the water

services group at Fiddlers Ferry Power Station.

CONTACT

Michael DaughertyMarketing communications managerFlowserve Pumps2200 E Monument AveDayton, OH 45402, USA.Tel: +1-937-226-4376E-mail: MDaugherty@flowserve.comwww.flowserve.com

f e a t u r e r e d u c i n g r u n n i n g c o s t s

Figure 11. Upgraded MBFP design.

Figure 10. Original MBFP design.

Upgrading boiler feed pumps in a UK coal-fired power stationThe initial scenarioHistory and field measurementsReverse engineeringHydraulic performanceEnergy saving opportunities/LCCEfficiency improvements on the MBFPReduced regulator valve throttlingHP injection, super heated, tapped from MBFP

Original pump design critiqueConclusionReferenceAcknowledgement