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THE AMERICAN SOCIETY OF MECHANICAL ENGINEERSy^ 345 E. 47th St., New York, N.Y. 10017

C Yi C The Society shall not be responsible for statements or opinions advanced in

l ^J papers or discussion at meetings of the Society or of its Divisions or Sections,

® or printed in its publications. Discussion is printed only if the paper is pub-lished in an ASME Journal. Papers are available from ASME for 15 monthsafter the meeting.

Printed in U.S.A.

94-GT-115

SOLAR TURBINES INCORPORATED "TAURUS 60"GAS TURBINE DEVELOPMENT

Vern Van LeuvenDepartment of Mechanical Design Engineering

Solar Turbines IncorporatedSan Diego, California

ABSTRACTThe Taurus gas turbine was first introduced in 1989 with

ratings of 6200 HP for single shaft and 6500 HP for twin shaftconfigurations. A new design of the single shaft third stageturbine rotor and exhaust diffuser brought its power to 6500 HPin 1991. A program was initiated early in 1992 to identifyopportunities to further optimize performance of the Taurus.Thorough investigation of performance sensitivity tothermodynamic cycle parameters has resulted in significantimprovement over the original design with no change in firingtemperature. Aerodynamic and mechanical design changes wereimplemented in 1993 which raised Taurus performance to 7000HP and 32% thermal efficiency. Selection of the final designconfiguration was the outcome of performance maximizationversus cost increase, durability risk and loss of commonality withprevious engines. This paper details these changes and the designselection process.

INTRODUCTION

HistoryThe Taurus 60 gas turbine (Figures 1 and 2) was originally

introduced to the market in 1989. Table 1 summarizes originalISO performance and significant features of the single and twoshaft versions. The engine configuration is similar to the SolarCentaur 50 with the most important changes being an additionalstage at the front of the compressor and a new design two stagepower turbine for the two shaft version. The compressor stage isa scale of the Solar Mars first stage blade while the two shaftpower turbine was a new design. In 1991 the third stage turbineand exhaust diffuser for the single shaft version was redesigned.Engine performance improved by 390 HP and -534 BTU/HP-hr.Quantities of engines produced from 6/90 to 10/93 were (22) 6200HP single shaft, (27) 6500 HP single shaft and (39) 6500 HP twoshaft engines.

Program ObjectivesIndustrial gas turbine designs in all size ranges are

continually responding to evolving application demands for higherengine performance and the Taurus 60 is no exception.Competitive challenges and the potential for design improvementusing advances in design and analysis technology motivated Solarto initiate a development program for the Taurus 60. Goals of theprogram were to deliver a product to the market in 18 months,maintain existing cost normalized by power, incur no sacrifice indurability and produce performance shown in Table 2.

Program SummaryThe program objectives provided a tough challenge to the

new product introduction (NPl) teaming approach newlyimplemented at Solar. The basis of NPI teaming is to assignresponsibility for product development to a cross functional teamso that communication and participation throughout all disciplinesare maximized. An NPI team was assigned by management andthe team's first task was to complete conceptual product design.

Fundamentally there are two ways to improve theperformance of a gas turbine. The cycle (pressure ratio and firingtemperature) can be changed or the component efficiency can beincreased (this includes reducing cooling air flow and leakage).Brainstorming sessions were held to create a list of options toachieve the program objectives with the criteria used in theevaluation being performance, durability risk, cost, developmenttime and product commonality.

Once preliminary design was completed and the optimumconfiguration determined detail design work was initiated. Atimeline was prepared summarizing the program, and tasks wereorganized with consideration to component lead times, manpowerand test rig availability. More refined performance, temperatureand stress evaluations were begun and definition for necessarytesting was created. Previous experience from Centaur 50 andoriginal Taurus 60 were also heavily utilized in this program.

Presented at the International Gas Turbine and Aeroengine Congress and ExpositionThe Hague, Netherlands — June 13-16, 1994

Copyright © 1994 by ASME

Downloaded From: https://proceedings.asmedigitalcollection.asme.org/ on 07/11/2018 Terms of Use: http://www.asme.org/about-asme/terms-of-use

FIGURE 1. TAURUS 60 SINGLE SHAFT ENGINE CROSS SECTION

RE93133

FIGURE 2. TAURUS 60 TWO SHAFT ENGINE CROSS SECTION

Both the single and two shaft designs were completed onschedule and the cost target was achieved. Performance of enginestested to date (Table 2) show that performance goals were alsoexceeded.

DISCUSSION

Conceptual DesignA substantial period of time was spent identifying all

potential design changes which might be made to achieve theperformance goal. A list of options was compiled and evaluatedusing the established criteria as illustrated in Table 3. Severaldifferent combinations of the potential modifications would satisfy

the performance goal but most involved compromises in cost,schedule and risk.

A configuration which increased firing temperature andpressure ratio was seriously considered in the beginning of theconceptual stage. This required additional stages added to the aftend of the compressor and a new gas producer turbine design. Thecompressor stages would be scaled form existing Solar Marsengine stages and would not require a lengthy development.However, the new GP was to have new nozzles and would requirea directionally solidified first stage GP turbine blade. The thirdstage nozzle would also have to be redesigned to provide theconnect flow characteristics. Although this configuration met theperformance goals, the increased engine cost of about 15%,

PA


TABLE 1. ORIGINAL PERFORMANCE OF TAURUS 60 ENGINES

Parameter Original Single Shaft Taurus 60 Original Two Shaft Taurus 60

Power (1) 6200 HP (Pre 1992)6450 HP

6500 HP

Cycle Efficiency (1) 29.6% (Pre 1992)30.3%

31.0%

Mass Flow 46.0 Lbm/s 46.0 Lbm/s

TR1T 1850°F 1850°F

Pressure Ratio 11.2 11.2

Exhaust Temperature 901°F 910°F

Speed 14950 RPM 15000 RPM GP14300 RPM PT (max)

Compressor Stages 12 12

Variable Comp Stages 4 4

Turbine Stages 3 4

Note: 1. All values represent no duct losses, natural gas fuel and 59°F, sea level, 60% relative humidity air.

TABLE 2. PERFORMANCE OF UPRATED TAURUS 60 ENGINES

Engine Power Efficiency

Single Shaft Taurus 60 6900 HP 32.0%Program Goal

Two Shaft Taurus 60 6950 HP 32.0%Program Goal

Single Shaft Taurus 60 6947 HP 32.0%Average - 8 Engines

Two Shaft Taurus 60 7007 HP 31.9%- 1 Engine

Note: 1. All values represent no duct losses, natural gas fuel and 59F, sea level, 60% relative humidity air.

longer program development time, and loss of commonalty withthe current fleet of Taurus engines led to closer evaluation ofother options.

The most attractive configuration was found by improvingcomponent efficiencies. This could be accomplished in a shortertime while maintaining fleet commonalty, low engine cost, andwithout increased durability risk. Efficiency improvements weremade for both compressor and turbine components in the newTaurus 60. In the compressor, the 0-stage blade was modified, tipclearances were reduced, improvements were made in surface

finish and the guide vane settings were optimized. The turbinealso had the tip clearances reduced as well as leakage and coolingflow reduction. One option that was in the concept initially wasa longer exhaust diffuser for both single and two shaft engines.This option was eliminated because the performance gain did notoutweigh the significant package changes required and loss ofopportunity for customers wishing to uprate their units. In thisinstance, the early design conceptualization involving SolarPackage Engineering and Customer Services groups facilitated bythe NPI process was of great benefit.


TABLE 3. MATRIX OF PERFORMANCE IMPROVEMENT OPTIONS

Modification Performance Cost Risk DurationSensitivity

Raise TRIT High High Med 24 MNew 1 st, 2nd GPTBlade Mat.

Raise Pressure High High Med 18 MRatio Add Comp Stages

Redesign GP High High High 30 MTurbine

Redesign PT High High High 30 M

Improve Diffuser Med High Low 24 MNew Package

Reduce Cooling, High Low Low 12 MLeakage Flows

Optimize 0 Stg Med Low Low 6 MComp Blade

New Air Inlet Unknown Med Low 12 M

Reduce Comp High Low Low 4 MTip CL Med w/Rub Coat

Reduce Turb High Low Low 4 MTip CL

Improve Comp Low Low Low 4 MSurface Finish

Optimize Comp Med Low Low 2 MGuide Vanes

KWl^zmm

Zero Stage Blade . The zero stage blade was a scale of theMars first stage compressor stage which was originally designedin 1975. Recent mechanical analysis of the blade with finiteelement methods revealed that the deflections due to centrifugaland aerodynamic loading were slightly different than what wasoriginally calculated. Today's capability of finite element softwareto account for large deformations where the load vector relativeto the position of the structure changes significantly withdeflection (Figure 3) was not available at the time of originaldesign.

It is extremely difficult to predict by analysis alonecompressor performance sensitivity to changes in blade incidencethus rig or engine testing is essential in the evaluation process.Rather than just correcting a known problem, testing was used tofurther optimize the stage incidence. The objective was to increase

flow as much as possible without sacrificing efficiency. Theairfoil incidence angle was modified and back to back enginetesting was performed to compare against the original design. Onecompressor case half was removed, the blades were changed outand the engine was rerun all in the same day. By changing outblades in the test stand a minimum amount of uncertainty wasintroduced in the evaluation. The result of several tests was ablade incidence angle which produced additional airflow and nodetriment to compressor efficiency.

Increase in power is summarized in Table 4. There were noadditional material costs associated with the change in newproduction and no mechanical risks were introduced.

Compressor Tip Clearance ReductionCompressor performance is known to be very sensitive to tip

clearances. An optimum value of running clearance is generallybelieved to be between 1.2 - 1.5% of chord length. Finite element

4


TABLE 4. RESULTS OF PERFORMANCE ENHANCEMENTS

Modification Performance Improvement

Zero Stage Blade Twist 114 HP0.0% Efficiency

Compressor Tip Clearance Reduction 180 HP0.7% Efficiency

Compressor Guide Vane Optimization 50 HP0.4% Efficiency

GP Turbine Tip Clearance Reduction 77 HP0.37% Efficiency

Single Shaft PT Cooling Reduction 92 HP0.29% Efficiency

Two Shaft PT Cooling Reduction 157 HP0.45% Efficiency

Piston Seal Leakage Reduction 110 HP0.19% Efficiency

U 1 - U2 = Difference in orientation of structureto load vector after loading. REM1

FIGURE 3. ZERO STAGE COMPRESSOR BLADEDEFLECTION UNDER LOAD

modelling of the compressor rotor and a component tolerancestackup study showed that there was potential for optimization oftip clearances. Adequate margin of safety must be included in thedesign because heavy nibs in the compressor cannot be tolerated.A rub can result in material build up at the contact area whichmay not clear away. If this happens, the rubbed material buildupcan continue to grow resulting in damage to the blading. Also, ifa stationary airfoil rubs against the rotor heating from friction cancause distortion, further tubbing, loss of material strength andpossible component failure. A tub tolerant coating could beintroduced to permit occasional rubs by design, however theexpense of a rub coating and the risk of it not adhering

sufficiently outweighed the benefit of tighter clearances. It wasdecided to reduce tip clearances as near as possible to optimumwithout permitting the possibility of tub.

Finite element methods were used to calculate the rotor radialgrowth due to thermal and centrifugal loads and static structurethermal displacements were calculated. Positional tolerance of therotor relative to the cases involved a stackup of four features onthe front end and three features on the aft end while diametraltolerance consisted of two features. Clearance values were then setso that rubs would never occur. To confirm proper machining andassembly, clearances were measured during build of each engineby alternately removing each compressor case half and measuringgaps with feeler gages. A development test engine was run toverify analytically predicted minimum clearances by brazing insmall diameter instrumentation tubing to act as rub pins. Worstcase thermal transient condition clearance was measured by thepins and compared favorably with design predictions.

Performance improvement due to tip clearance reduction wasdetermined both by comparison of test engine performance builtwith the tighter clearances against data from previous productionengines, as well as a back-to-back test with the compressorreturned to original clearances. The measured performancesensitivity to tip clearance compared well with analyticalpredictions. No cost and minimal durability risk are associatedwith this change.

Air Inlet EvaluationAnother area for potential performance improvement analyzed

was the air inlet collector and duct. Previous testing withtraversing probes showed a potential flow distortion in the airentering the compressor inlet guide vane. Computational fluiddynamics (CFD) modeling with the ACE (copyright CFD

5


Research Corp.) computer code also showed some potential flowfield distortion (Figure 4). However, the flow velocities in thisarea are low and the influence of this distortion on compressorperformance was not known. A back-to-back test was designed toshow the potential for compressor performance improvementbefore changing the production design. Static pressure taps alongwith Kiel Probes were used to show the flow distortion, andengine performance was tested, first with the inlet collector thenwith the collector removed (see Figures 1, 2). There is adequatespace on the engine package so that negligible flow distortionoccurs entering the compressor with the collector removed. Thissimulates the lowest flow distortion expected from a collectorredesign. The test revealed that any flow distortion created in theair-inlet was not reducing engine performance. Therefore, nochange to the existing air intake was made for the new Taurus 60.

RE93135

FIGURE 4. STATIC PRESSURE DISTRIBUTIONIN COMPRESOR INLET

Compressor Surface FinishThe compressor flowpath materials are stainless steel and

ductile iron which require coating for optimum corrosionprotection. The fact that the surface finish of the compressor hasa significant effect on performance is obvious to anyone that hasever washed a dirty gas turbine in service. However, therelationship of uniform surface roughness measured with aprofilometer to compressor performance is not well quantified.Taurus 60 compressors had previously been coated with Sermatel725 coating with a typical surface roughness of 65 pin. Sermatelalso offers a coating with a surface roughness objective of 10 pin(0.010 in cutoff) called Sermatel 5380 DP ([Mosser, 1988]. Sincethe difference in cost was not great and the corrosion protectionwas expected to be as good as the 725 coating, the team chose topursue the 5380 DP coating. Surface roughness testing wasperformed on a Solar airfoil uncoated and with both coatings and

the results are summarized in Table 5. The 5380 DP coatingunfortunately did not meet its roughness objective. Still, the finishwas better than previous coating so evaluation continued.

TABLE 5. SURFACE ROUGHNESS TEST RESULTS

Component Seimatel 725Sermatel5380DP

Compressor Blade 65 pin 60 on

Compessor Vane 51 pin 36 pin

Note: Roughness measure ment made using 0030.in cutofflength profilometer.

Back to back performance testing without introducingvariables more significant than the predicted benefit from thecoating could not be performed. To be conservative, noperformance increase was predicted from the new coating. Sincecost and durability risk were low and there may also be a benefitof improved dirt fouling resistance the new coating wasimplemented in the design.

Compressor Guide Vane OptimizationRecent expertise gained by the aerodynamic group in the use

of a powerful multivariable optimization routine which is part ofthe RS1/Discover (copyright BBN Software Products Corp.)computer program permitted fast and accurate optimization of theguide vane settings. Optimization variables were individualpositions of the four guide vane stages and the objective variableswere engine horsepower and specific fuel consumption. A matrixof the variables was first generated using 17 differentcombinations of guide vane settings. The computer program fitsa regression to the data and then solves for the vane settings atmaximized objective variables. Some iteration around the solutionis required as the regression fit improves with more data. Figure5 displays the Taurus 60 compressor efficiency as a function offlow with isometric lines shown for power and cycle thermalefficiency. Points were plotted on this curve during theoptimization to graphically show how the compressor flow andefficiency are interelated and what tradeoffs must be made whenchoosing cycle power or efficiency as an objective.

Optimization was performed using two different productionengines to improve accuracy of the result as compared to theaverage of production engines. Results from this testing is shownin Table 4. No durability impact or cost increase is incurred withthe change of guide vane settings.

Turbine Tip ClearancesIn the attempt to reduce turbine tip clearances, teaming

between the design and manufacturing engineers provedbeneficial. Manufacturing suggested an improved machiningprocess that would reduce the tolerance on the first and secondstage turbine nozzle tip shoes. Previously the rub tolerant coating


+0.75

'Svx +0.50

0 o`y x?o0

+0.254 ° xr0

0 0

^-0.25

moo -0.50

- 0 . 75 G Thermal Efficiency %

-1.00

o +2>-CZw

LL +1U-wUI—m 00

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RE93144M

:OOLING FLOW SCHEMATIC

. It was suspected that there could bethe piston ring seal on the first stagepaint tests confirmed the high leakage

;ak paths inherent in the design. A newh incorporated several ideas to reduceus leaks between the nozzle rails andur when the nozzle rail distorts undersized by trapping the rails in a rigidised at the top of the rails to furthering is now assembled into the carriering around the circumference. The new;ing comprehensive thermocouple31 paint. Back to back tests of the old)ns indicated that, as predicted, leakage

Quantification of the performanceaccomplished with the instrumented

-rs felt that a field durability test wouldiroduction implementation. Since theprogram could be met without thereed to delay implementation of thisi 5000 hour field testing results are

0.005 inches, without reducing the minimum clearance. Previousanalytical work, testing and field experience had alreadydemonstrated that the minimum clearance should not be reduced.

Tip clearances of the power turbine stages had been fullyoptimized in the original development program.

Turbine Cooling Flow ReductionThe evaluation of reducing cooling flow began with a cycle

analysis showing the sensitivity of cooling air use and leakage ateach stage. Figure 7 and Table 6 show the sensitivity of cyclepower and efficiency on cooling flows and leaks. With theseresults the team established a very high priority for evaluatingflow reductions in the power turbine. The analytical investigationbegan with an extensive network flow analysis using previous testdata to correlate the model. Cooling and leakage changes couldthen be quickly evaluated. Table 3 summarizes the potential flowreductions that were identified from this study.

-3 -2 -1 0 +1 +2

AIRFLOW, IbisRE93143M

FIGURE 5. TAURUS 60 COMPRESSORPERFORMANCE MAP

had been applied to the integral tip shoes (Figure 6) withdimensional control based on stackup of individual componenttolerances. The .improved process is to machine the coating to afinish dimension referenced from the locating features of thenozzle. This allowed the reduction of the nominal tip clearance by

AGE


TABLE 6. TURBINE COOLING FLOW PERFORMANCE SENSITIVITY

Location Horsepower Cost (HP) Cycle Efficiency Cost

A 17 0.085

B 0 0

C 108 0.178

D 108 0.178

E 128 0.279

F 132 0.297

G 148 0.374

H 183 0.546

I 183 0.546

J 176 0.511

Note: 1. See Figure 7 for location schematic.2. Performance penalties are normalized to 1% compressor mass flow.

FIGURE 8. GAS PRODUCER TURBINE PISTON RINGSEAL CONFIGURATION

Power Turbine Disk Impingement Flow. The perfor-mance sensitivity study directed the most in depth optimization ofcooling flows to the power turbine section. Analytical evaluationsof both single and two shaft turbines and previous testing during

1st STAGE CARRIER PISTONNOZZLE RING SEAL C-SEAL HOUSING

RE93138M

FIGURE 9. IMPROVED GAS PRODUCER TURBINEPISTON RING SEAL CONFIGURATION


TABLE 7. DISC METAL TEMPERATURE SUMMARY

Disk Location Original Taunts 60UprateTaunus 60 Change

Single Shaft

3rd Turbine

Rim 950°F 960F +16°F

Bore 790°F 800°F +10°F

Two Shaft (1)

3rd Turbine

Rim 975°F <1060°F + <85°F

Bore 730°F <780°F + <50F

Two Shaft (1)

4th Turbine

Rim 960°F <1060°F + <100°F

Bore 490°F <540°F + <50°F

Note: 1. Upper limit temperatures are from preliminary tests. Results of final development testingwere not available at the time of this writing.

the original design phase of the power turbine indicated thatimpingement flows to the power turbine disks could be completelyeliminated without significant rise in disk temperatures. The mainreason for the opportunity in the design is that originally thecooling system was sized for a potential thermal uprate. Aspreviously mentioned, an increase in firing temperature was notpart of this program. Testing of both the single shaft and twoshaft turbines was performed with instrumentation as shown inFigures 10 and 11. Disc metal temperature increases fromcomplete elimination of cooling air is summarized in Table 7.

Performance evaluation could not be performed with theheavily instrumented test configurations so final confirmation ofthe improvements was taken from first production tests. FromTable 4, it can be seen that the cooling air reductions in theturbine were responsible for a significant portion of theperformance improvement of the program. Excellent agreement ofactual engine performance with predictions thus validates the flowand sensitivity analyses.

Elimination of impingement flows served to simplify both thenozzle machining and casting resulting in significant costreduction. The cooling passage brazed covers, cast cooling cavityand impingement holes were eliminated.

Single Shaft Exhaust Diffuser ImprovementsAlthough the longer exhaust diffuser concept was determined

to be impractical, other methods of improving the diffuser'sperformance were also evaluated. The previous exhaust diffuserhad 0.12 in thick sheet metal that protruded into the flow path.This causes a detriment to performance because of the upset flowalong the walls and inhibited diffusion. In the quest to eliminatethe flow-path step the manufacturing and design engineering teamworked to improve the design. The shop floor worker whopersonally manufactures the diffusers provided the most valuableideas and a design change was realized that not only eliminatedthe step but reduced cost and manufacturing lead time. Thesuccess of this effort reveals the value of soliciting design inputthrough our cross functional teams.

• Metal Temperature • Air Temperature •PressureRE93139M

FIGURE 10. SINGLE SHAFT TURBINEINSTRUMENTATION DIAGRAM

9


• Metal Temperature ■ Air Temperature A PressureRE9314C 1

FIGURE 11. TWO-SHAFT TURBINE INSTRUMENTATIONDIAGRAM

Two Shaft Version Power Turbine Thrust BearingThe increase in pressure drop through the power turbine of the

two shaft version of the Taurus 60 resulting from the compressorperformance enhancements mandated a review of the thrustbearing load carrying capacity. Figure 12 shows the predictedthrust as a function of rotor speed. The thrust bearing design mustanalytically demonstrate a minimum oil film thickness of 0.0005in and a maximum pressure-temperature severity ratio of 0.5throughout the design speed range. Pressure-temperature severityratio is a measure of the bearing babbitt material yield stressmargin at a given temperature and pressure. The existing fixedgeometry tapered land thrust bearing (Figure 13) did not satisfythe criteria so a tilting pad bearing with a larger area wasdesigned (Figure 14). Film thickness and pressure-temperatureseverity ratio are plotted against speed for several ambienttemperatures in Figure 15. Engine testing with load cells andtemperature sensors in the thrust bearing verified analyticalpredictions.

4300-20 0 20 40 60 80 100

AMBIENT TEMPERATURE, OFRE93145M

FIGURE 12. TWO SHAFT POWER TURBINE THRUSTVERSUS ROTOR SPEED

SHAFT CENTERLINE

COLLAR1-1

,,n HOUSINGRE93141 M

FIGURE 13. ORIGINAL POWER TURBINE THRUSTBEARING DESIGN

CONCLUSIONThorough review of the Taurus 60 design from a team

representing all disciplines of Solar's industrial gas turbinebusiness resulted in an improvement in product performancewhich met the program performance goals. The program schedulewas met and no increase in product cost or durability risk wasintroduced. The performance improvements have been verified by9 production engine tests. The key to the success of the programwas the NPI teaming concept because it insured continuinginvolvement and product ownership from a wide area of expertise.

12,920

11,850

10,770

9690O0

8620

D

= 7540

6460

5380

NpT, % of 14,300 rpm

\SO

60

100

NGP = 15,000 rpm

[---faT5 = 1400°F

10


Responsibility to a focused group rather than as traditionally to afunctional area contributes primarily to these motivations. As aresult, critical evaluation of the design criteria were applied toeach concept in a systematic and unbiased fashion.

REFERENCESMosser, M. F., 1988, "An Improved Coating Process for Steel

Compressor Components - SemieTel Process 5380 DP", SAETechnical Paper 880879.

SHAFT CENTERLINE

THRUST I ILl PADCOLLAR THRUST BEARING

RE93142M

FIGURE 14. NEW POWER TURBINE THRUST BEARINDESIGN

Ta = 120°F

Ta = 68°F

^Ta

0.2 I I I I I

'

6 7 8 9 10 11 12 13 14 15 16SPEED, rpm (000)

RE93146M

0Q 0.5

0.4waU) 0.3w¢D

0.2waw 0.1wD 0W 6 7 8 9 10 11 12 13 14 15 16a SPEED, rpm (000)

RE93147M

FIGURE 15. TWO SHAFT POWER TURBINE THRUSTBEARING PERFORMANCE

1.6

1.4

E 1.2J

LL 1J

00.8

0.6

0.4

11


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