E AMEICA SOCIEY O MECAICA EGIEES 864 4 E. 4 St., Yr, .Y. 400

e Sociey sa o e esosie o Saemes o oiios aace i aes o i is- cussio a meeigs o e Sociey o o is iisios o Secios o ie i is uicaios iscussio is ie oy i e ae is uise i a ASME oua aes ae aaiae om ASME o iee mos ae e meeig ie i USA Copyright © 1985 by ASME Downloaded from http://asmedigitalcollection.asme.org/GT/proceedings-pdf/IGT1985/79429/V001T02A015/2477464/v001t02a015-85-igt-41.pdf by guest on 27 September 2021 A CtEfftv rfrn vlpnt f th 6A6 rbprp Cprr

SUKASA YOSIAKA nd KEE S. UE Aeoyamics eame a Wiey Caaa Ic

ABSTRACT = T/288.3 (°K) : enthalpy rise coefficient = ip/ri The largest member of the PT6 engine family, the PT6A-65, was : flow coefficient = Cm/U developed in the early 1980's and went into production in September 1982. The : pressure rise coefficient = (JOH )/U for this engine consisted of four new axial stages combined with an existing centrifugal stage on a single shaft. This paper gives a brief description of Subscripts the studies leading up to the choice of the compressor configuration and a more c : choking detailed examination of the development of the chosen compressor to the is : isentropic required performance level. m : meridional component The development of this compressor presented a two-fold technical max : maximum challenge. Firstly, the limited space in the small compressor gas path did not p : peak efficiency point permit the effective use of conventional total pressure and temperature probes : surging point for performance evaluation. Secondly, the short time available for development o : engine inlet excluded some attractive corrective measures such as the redesign of some of the 3 : diffuser exit axial blade rows because the time required would have jeopardized the meeting of the tight development deadline. The first problem was overcome by a combination of limited wall static pressure measurements and an extensive use INTRODUCTION of numberical flow analysis codes. This approach proved to be quite cost- effective. The second was solved by the adaptation of an existing fully The first PT6 turboprop engine, PT6A-6, was marketed in 1962. By the analytically-designed research axial stage to the first stage position in the axial mid 1970's this ancestor had sired an extensive series of turboprop engines compressor. covering a take-off thermal power range from 450 shp to 1120 shp (gearbox limited) under sea level standard day conditions. In 1974, Pratt & Whitney NOMENCLATURE Canada Inc. (P&WC) was forecasting a demand in the 1980-1990 time period for a turboprop engine in the 1200-1300 shp range for high-speed, high-altitude

A cross-sectional area at the leading edge of the 1st stage rotor (m) executive aircraft applications. A series of variations were

C absolute velocity (m/sec), or chord length (mm) studied in an effort to satisfy the predicted performance requirements. In the

diffusion factor case of the compressor for this engine the number of possible configurations had

i absolute total enthalpy change across the compressor or been reduced to three by the spring of 1978. These were (A) the addition of a (Joule/ kg) zeroth stage to an existing compressor; (B) the combination of a newly designed

MCA multiple circular arc four stage axial compressor with an existing centrifugal stage (4A + IC); (C) a m mass flow (kg/sec) completely new 3A + IC compressor. Configuration A was eliminated in spite A = (ria-fri) (kg/sec) at a constant speed of its low technical risk and low development cost advantages due to insufficient absolute total pressure (kg/ern ) performance improvement over the current PT6 models. Configurations B and

Ps static pressure (kg/cm-) C were predicted to be about equal with respect to performance. However, the

PR pressure ratio former was considered to be a lower risk design than the latter due to the lower

R rotor, or radius (mm) pressure ratio per stage for the axial stages and the use of the existing high

S stator performance centrifugal stage. In addition the 4A + IC configuration (B) was

S/M surge margin = HPR/MV-0/6)] s/[PR/M/0/6)], — I x 100 70 expected to have a higher uprating potential than the 3A + IC (C).

T absolute total temperature (°K) Consequently, the 4A + IC configuration was the final choice . blade thickness (mm) This paper begins with a brief description of the aerodynamic design of the

U rotational speed at mean radius (m/sec) four axial stages for this new compressor. This is followed by a more detailed

X axial length (mm) discussion of the performance development of the compressor. The blade or vane stagger angle (deg.) development was technically challenging for the following reasons: P/1.033 (kg/cm ) I Conventional total pressure and temperature probes for the evaluation of compressor adiabatic efficiency compressor stage performance could not be used effectively in this unit

Presented at the 195 eiig Ieaioa Gas uie Symosium a Eosiio eiig eoes euic o Cia — Seeme 1-7 195 because of the small gas-path dimensions (see fig. 1). They had the = 1 undesirable effects of limiting flow at high speed and of introducing excessive losses. 7 1 2. Due to the tight development schedule, it was impossible to undertake a redesign of the compressor without imperiling the meeting of the „- 9 development deadline. Even the time allowed for the optimization of stage matching was severely limited for the same reason. E 9

-631' Downloaded from http://asmedigitalcollection.asme.org/GT/proceedings-pdf/IGT1985/79429/V001T02A015/2477464/v001t02a015-85-igt-41.pdf by guest on 27 September 2021 E 1.6 o oo eaie Mac ume "a is" E 1.4 Sao asoue Mac ume "a u" [150- 1.2 1.0 eicio 0.8 -es esus 0 6 1 Ei ° 4 - 1 oi Imee 9 97 (WA/(/A esig R4 S4 S1 9 = 1 (% R1 45 39 2 50 3 aes •7 15 blades 99 mm) 0 50 100 15 93 1 770 97 4 (inches)0 2 4 X 6 8 cc 97 19 99 igue 1 Meiioa aia comesso gas a (oigia esig 93 c_ 97 As a result of these constraints the compressor performance development 87.2 was limited to gas generator tests with minimum gas-path instrumentation. 5

AXIAL COMPRESSOR DESIGN 75 80 5 9 95 1 15 (iogo/o / (oA/o esig (° The four axial stages of the new compressor were designed during the second half of 1978 using the well established P&WC axial compressor design igue Comesso ma (coig 1 system. The design point pressure ratio for the axial stages was 4.26:1. The compressor had a "front loading" type of work distribution which corresponds approximately to a constant enthalpy rise per stage. A maximum design The philosophy of no variable vanes, which is one of the traditional diffusion factor D, of 0.45 was selected to ensure an adequate surge margin features of PT6 ensuring high engine reliability, was maintained for without variable geometry. This limit combined with the relatively high stage this new design. However, an interstage bleed arrangement located between the loading implied in the design pressure ratio resulted in a high tip speed design. fourth stator and the impeller, as on the other PT6 models, was used and a jet- Consequently, the inlet relative flows to the first three rotors are transonic with flap system was added to the intake case. These features have been more than the tip relative being 1.49, 1.17 and 1.05 for the first, second and adequate to avoid possible mismatch problems between the front and rear stages third stage rotors, respectively. The meridional gas path geometry of the axial at start-up and idle. They are both shut off at all power rating points. stages, (fig. 1), displays an essentially contant average radius. The correponding design point mean line Cm distribution was also nearly contant. Table 1 COMPRESSOR PERFORMANCE DEVELOPMENT contains the design values of various aerodyamic and geometric parameters for the axial stages. By the time the engine development was embarked upon in late 1979 two customers had decided to use this engine, designated PT6A-65, on their new commuter aircraft. In addition, one of them had a plan to develop an executive ae Rotor Stator Rotor Stator Rotor Stator Rotor Stator version of the new airplane. This implied that the compressor, originally ow 1 1 2 2 3 3 designed for high-speed, high-altitude applications, was now required to provide Ie ai age eg 15 353 1 59 35 5 7 high performance down to 9 1 of design speed, N 1 , for hot day take-off, while Ei ai age eg 5 11 59 1 9 13 5 maintaining the low sfc capability up to at least 105% N1, 1 for high altitude Ie mea age eg 5 15 577 77 5 7 571 9 cruise. This had to be accomplished without variable geometry. Ei mea age eg 55 35 5 11 39 3 1 Soiiy 19 11 19 1 19 1 15 13 Test of Baseline Compressor (Config. 1) ickess/co 51 7 1 5 1 59 aco 39 77 39 359 3 33 In November 1979 the first was taken between 89.7% and A ages om aia 100% N 1 on a gas generator which had been allocated to the compressor development program. The test results are compared with the predicted map on Fig. 2. Note that the pressure ratio and efficiency were based on the test cell ae 1 Coig 1 aia sages Meaie esig oi aa static pressure (Station 0 of Fig. 3) and the casing static pressure (Station 3 of Fig. 3). Due to a rotor vibration indication the testing at 97.4 5 and The rotor blade were multiple-circular-arc (MCA) profiles to 100% N 1, was limited to low pressure ratios. The test data showed e oowig provide good low loss characteristics in the high subsonic and low supersonic performance characteristics. flow environment. The stator vanes consisted of NACA65-Series airfoils on I. The measured peak efficiencies are within three-quarters of a point of the circular arc camber lines. To reduce manufacturing costs the second, third and target values up to 94.9% N o . fourth stage stators were made from constant-section strip-stock which was 2. Flow and pressure ratio along the target running line are considerably lower coined to provide the correct stagger and camber. The blades and vanes of the than the design intent. At the design speed, for instance, they are lower than compressor were designed to satisfy not only the aerodynamic requirements but the design values by 3.0% and 4.4%, respectively. also to provide adequate steady state and vibratory stress margins and FOD 3. The pressure ratio versus flow characteristic, at a constant speed, is resistance. The rotor blades, discs and spacers were made of Titanium alloy to significantly steeper than the predicted map. take advantage of its low weight and high strength. The stator vane material, on 4. The measured peak efficiency points appear at a higher operating pressure the other hand, was AISI 410 stainless steel because of its cost advantage. ratio than the target.

2 Suge / Saio 3 ege o coke o s / o Saio /

- " 3 o si 3 o s o s _ o s s 3 o S o s euce ow caaciy Downloaded from http://asmedigitalcollection.asme.org/GT/proceedings-pdf/IGT1985/79429/V001T02A015/2477464/v001t02a015-85-igt-41.pdf by guest on 27 September 2021

igue 3 Isumeaio scemaic as sage An interpretation of the performance characteristics described above can be summarized as follows: igue 5 yica mui-sage comesso eomace

Characteristic 1 Although the predicted improvement in peak efficiency of The inadequate flow capacity of a rear stage may be due, as previously approximately one point during the increase in the rotating speed from 89.7% to mentioned, to an actual lack of sufficient flow area in the rear stage or it may be 94.9% N o did not occur the compressor efficiency was, nevertheless, considered due to an inadequate work capacity in one or more of the upstream stages. Data reasonable for an initial development test. It should be noted that peak reported in [3] and [4] support this argument. Some of these are plotted in Fig. 6 efficiencies were not reached at 97.4% and 100% 1\1 . in terms of operating range and surge margin as a function of the stator vane stagger angle. The data from a 16 stage axial compressor taken from [3] are for Characteristic 2: Obviously either one of the four newly designed axial Configurations C and D at 85% and 90% N . Data for speeds below 85% o were not used since the e])-tp characteristic of the first stage indicated that it was compressor stages or the centrifugal stage was choking prematurely. According already stalling. The data available at speeds aoe 9 D wee too sparse to to a choking criterion developed in [1]* the first stage rotor is not choking even permit an accurate evaluation of surge margin. The information taken from the at design speed since (14),/d(N/f8) is highly positive at all speeds listed 12 stage axial compressor of [4] has been limited to that for stator vane stagger (see fig. 4). angles within ± 5 degrees of the baseline value in order to eliminate possible inlet guide vane or first stage rotor choking. Fig. 6 shows a general increase in operating range and surge margin as stage areas and work capacities are increased by reducing stator stagger.

Symo e We (% Cage o sagge A 3 5 i S1 oug S15 3 9 ------__ ♦ 1(? (1 c 5 esig

Coig 1

5 1 95 1 (/9/(/ esig (%

igue Cokig ow coeicie (coig 1 Oeig Cosig — — — Characteristic 3 & 4: For a single stage compressor the pressure ratio versus Cage o sao sagge age A (eg flow characteristic tends to steepen as the rotor blade stagger angle increases. For a multi-stage compressor there is a general trend for the pressure ratio versus igue Eec o sao sagge age o comesso flow characteristic to steepen as the flow capacity of the rear stages is reduced. eomace caaceisics This phenomenon may be explained by the following simplified model. When a multi-stage axial compressor is operated at a speed less than but The stage characteristics of Config. I of the PT6A-65 compressor were close to design speed the typical 4)-ip and + - characteristics are as shown in calculated from a limited supply of measured shroud static pressure data (see Fig. 5. The first stage operates slightly on the stall side of peak efficiency while Fig. 3 for instrumentation). The calculation method was basically an analysis the last stage operates on the choke side. Consequently, the overall compressor version of the off-design performance prediction method of [5] The efficiency is comparable to that at design speed and is fairly constant over a characteristics shown in Fig. 7 indicate that the first stage lacks work at the limited flow range. As discussed in [2] and also in the Appendix the range of 4) design flow coefficient and that the third stage may be providing more work values between surge and the verge-of-choke point tends to widen as one than the design intent. However, the measured static pressures on which the proceeds downstream in a compressor. This effect is particularly evident in the stage characteristics were based are influenced significantly by shroud surface rear stages which are all operating in the negative slope region of the [4 - tp irregularities. In addition, the flow blockages and deviation angles used in the characteristics. Now, suppose the flow capacity of the last stage has been program were based on simple routines. Consequently, the absolute accuracy of reduced, (solid symbols of Fig. 5). Then the 4)-ip and characteristics of the the stage characteristics obtained was felt to be relatively low and they were used last stage are shifted to lower (I) as shown in Fig. 5. The flow coefficient of the to provide qualitative guidance only. Thus it was concluded from the test results front stages in the compressor decreases with a corresponding increase in If of Config. 1 that the newly designed four stage axial compressor was either the of the compressor is maintained the rear stage, which producing lower work than expected or had an inadequate flow area in one of controls choke flow, is forced to operate at a much lower ip value and thus at a the rear stages. The lack of work could be due to an inadequate deviation lower efficiency than before. Consequently, the operating range diminishes and allowance in the original design. To solve the performance problem the the peak efficiency point is shifted toward the surge point. following options were considered. 1. Redesign the first stage. This option was not viewed with enthusiasm since it would have been * Numbers in brackets designate references at the end of the paper. expensive and would have required approximately six months to do the

3 redesign and to procure hardware. However, in order to provide for the dollars). Therefore, two first stage rotor discs giving blade stagger reductions eventuality that it might be the only solution it was decided to do the redesign of two and four degrees and one disc, with a stagger reduction of two to be prepared for possible hardware procurement in the future. degrees, for each of the second, third and fourth stage rotors were procured. 2. Adapt an existing stage to the first stage position. A high performance research axial compressor stage with a 1.63 pressure Test Results of Config's. 2 Through 5 ratio capability [6] was felt to provide the most appropriate replacement for Config. 2 with a four degree reduced stagger first rotor (see Table 2 for a the existing first stage. A short study indicated that the incorporation of this definition of the compressor configurations) through Config. 5 with the stagger research stage in place of the original first stage of the PT6A-65 compressor reduced two degrees in all four rotors were tested in succession. The resulting would produce a compressor with the desired performance due to its reduced compressor maps are displayed on Figs. 8 to 11. An examination of these test

aerodynamic loading and increased flow capacity. Although hardware costs results produced the following observations: Downloaded from http://asmedigitalcollection.asme.org/GT/proceedings-pdf/IGT1985/79429/V001T02A015/2477464/v001t02a015-85-igt-41.pdf by guest on 27 September 2021 would be as high as for Option 1, the design elapsed time and performance 1. The compressor behaved essentially as expected. The pressure ratio versus risk would be low. Consequently, work was started on this option flow characteristics and the peak efficiency points shift towards choke as the immediately. flow capacity of the rear stages, particularly the fourth stage, is increased either by reducing the stagger of their rotor blades or by increasing the - 5 1s sage loading of upstream stages.

14 A esig Coig 1s sage sage 3 sage sage Cent'l stage Si• • esig oo Sao oo Sao oo Sao oo Sao 3 - - (1 (S1 ( (S (3 (3 ( (S Imee iuse (- 1Q as I esige (a/ Symos a/ (/c1/( eki design (%) age; - 3 3 a/ ° oe- a/ ° oe- a/ A 97 sagge sagge a/ ° oe- a/ ° oe- a/ ° oe- a/ - - 9393 sage sagge sagge sagge 99 5 ° oe a/ ° oe- a/ ° oe- a/ ° oe- a/ (•A, A sagge sagge sagge sagge 9 Moiie Moiie a/ 3 - 3 eseac eseac G> • 1 oo sao 7 Moiie Moiie I a/ Ii i aius a/ - ICI 3 sage eseac eseac euce oo sao I y 51m ( i A 1 A A411r". 3 - S 3 ae eeome comesso coiguaio - 5 2. Fig. 12 shows that &lid (NATO) of Config. 4 decreases between 97.4% and 100% N . This implies that the first rotor, as designed, is approaching 11 choking at design speed at a value of + c = 0.496 which is 3.3% lower than the design value. On the other hand, Config. 5 with the first rotor stagger 3 - reduced two degrees, giving a 5.4% increase in throat area relative to the 7 - 1 9 Ce sage I 7 1 3 5 3 5 th 9

igue 7 Sage eomace caaceisics (coig 1 9 3. Modify the original airfoils. The geometric modification with the lowest cost and leadtime is the restaggering of rotor blades. A camber change of rotor blades requires a redefinition of the airfoils and the design of the appropriate tooling. The restaggering of stator vanes entails the provision of a new stator shroud esig assembly with restaggered vane slots since the stator vanes are brazed into oi shaped slots in the shroud. In the case of the rotors, which were made 1 up of separate blades inserted in dove-tail slots in a disc, the restaggering (iuo(iuo esig esig could be accomplished merely by broaching the slots at the appropriate angle 9 in the disc. Therefore rotor blade restagger became the primary means of 97 = 1(% changing geometry. Since the axial compressor featured a "front-loading" 00 type of design it was decided to increase the aerodynamic loading of the third 0 9 and fourth stage rotors by reducing their stagger angles by 2 degrees. This 93 modification would also result in an increase in the flow capacity of these 97 7 97 stages. The effect of this change on part-speed compressor performance was 9 99 estimated to be insignificant. On the other hand reducing the stagger of 93 rotors one and two was expected to cause a significant efficiency 7 deterioration at part speed. However, it was felt that there was considerable 97 _ merit in testing these modifications as well since the results would lead to a fuller understanding of the performance characteristics of the compressor. 5 This extra knowledge could, in turn, facilitate a more rapid and accurate diagnosis of the performance problems and lead, hopefully, to a quick 75 5 9 95 1 15

solution. The potential technical gain was felt to be well worth the extra (o/-/ / (m1o/o esig (°/ engineering effort and cost since the turn-around time of a gas generator test is only one or two weeks at a cost of about $25,000 (1980 Canadian igue Comesso ma (coig

aseie oo yies a increase of 4.0% in choking flow without any indication of first rotor choke. Based on these data it was estimated that a 4.7% increase in the "as designed" first rotor throat area would be adequate to pass the design point flow with a satisfactory choke margin. This throat area increase was incorporated in the research stage rotor which was being adapted to replace the original first rotor of the PT6A-65 compressor. The decreased choke flow of Config. 2 relative to Config. 5 was n attributable to the excessive aerodynamic loading experienced by the first stage rotor when its stagger was reduced four degrees. The 1)-ip

characteristics of this stage are flat up to 100% 1\1 and all indications are that Downloaded from http://asmedigitalcollection.asme.org/GT/proceedings-pdf/IGT1985/79429/V001T02A015/2477464/v001t02a015-85-igt-41.pdf by guest on 27 September 2021 the stage is operating in a stalled regime. This inefficient operation of the first stage pushes the third and fourth stages into premature choke at part speed. There was no evidence from this series of tests that the centrifugal stage was running under choking conditions. 3. The deterioration of the peak overall compressor efficiency of Config. 3 relative to the baseline (Config. 1 is small as was expected. In fact, the 1 overall compressor efficiency of Config. 3 is slightly better than Config. I, along the estimated compressor running line, since the peak efficiency points 1 for Config. 3 have moved towards choke. 9

c cc _ 7

1= 7 5 75 5 9 95 1 (iicii/o / (o/o esig (

igue 9 Comesso ma (coig 3 1

i 1 9

1 9 cc _ 7 e 9 a

5 1

1 igue 11 Comesso ma (coig 5

9 1 esig 97 Coig (/gg/(esig Coig 3 = 1 (% Coig Coig 5 cc 7 5

-

5

75 5 9 95 1 15 5 9 95 1

(ioA/o I (1A/o esig (°i (//(/-E/ esig (-

igue 1 Comesso ma (coig igue 1 Cokig ow coeicie (coig oug coig 5

5

4. The efficiency deterioration of Config. 4, with the stagger of rotors two, The modification of the research stage involved a radial shift of the rotor three and four reduced by two degrees, is at least one point relative to Config. and stator airfoils by a factor of 1.018 to match the spanwise relative Mach 1. This change is considered to be excessive. Config. 4 did achieve the desired number distribution. In addition the stagger of the rotor blades was increased pressure ratio at each speed and the shape of the speed lines was very similar 1.5 degrees to provide the desired flow capacity. Since the research stage had a to the design intent but the flow was less than required. These characteristics higher ni/A (183.1 kg/sec/m ) and pressure ratio capacity than the baseline first suggested that it might be possible to raise the engine operating line on the stage these modifications were mutually compatible. In addition the closure of compressor map to maintain the "design" pressure ratio versus speed the rotor improved the high incidence angle situation existing on the research relationship. This should permit the compressor to run at or near the peak stage stator. A through-flow analysis was performed to define the meridional efficiency zone and would produce a cycle pressure ratio gain at a given flow. gas path precisely and the final geometry is compared with the original design on

A limited series of performance and handling tests on a full engine confirmed Fig. 14. The new gas path resulted in a modest increase in the aerodynamic Downloaded from http://asmedigitalcollection.asme.org/GT/proceedings-pdf/IGT1985/79429/V001T02A015/2477464/v001t02a015-85-igt-41.pdf by guest on 27 September 2021 the feasibility of this idea. It should be noted that a surge line survey was not loading of the second stage rotor. This was accepted because it resulted in a carried out since a blade and vane vibration check with strain gauges had not substantial reduction in the cost of modifying and reprocuring the second stator been completed. The - characteristics of Config. 4 are shown on Fig. 13. assembly. A meanline performance analysis indicated that this new configuration could be expected to meet the compressor performance requirements for the 5 - - 1s sage PT6A-65 engine. This compressor geometry was identified as Config. 6. A esig Test Results of Config. 6 esig 3 - Config. 6 was tested in March, 1981. As shown on Fig. 15 the required efficiencies along the compressor running line were achieved up to 96.2% N Symos and the design mass flow was slightly exceeded. The peak efficiencies at 97.4 and (//( a esig ( 100% N , could not be obtained due to excessive turbine inlet temperatures since - 3 this was a gas generator test. This compressor achieved target efficiency at all A97 speeds below 97.4% N . At 97.4% N the test data showed that the

compressor was capable of producing design point pressure ratio and efficiency sage - 93 but at a correspondingly lower flow. At 100% N and design pressure ratio the 1 A 99 efficiency was significantly lower than the design intent. This compressor - 9 configuration was tested as part of a complete engine, with a higher than design 3 - 3 operating line, and demonstrated better than target sfc at both take-off and • 1 cruise conditions. However, the turbine inlet temperatures were found to be - 3 sage higher than desired. Since it was considered difficult to provide art additonal improvement in compressor performance in the short time available, rematching A between the compressor and turbine was the only realistic approach to this problem. 3 - 3 To achieve a solution it would be necessary to move the peak efficiency zone closer to choke and at the same time, lower the compressor running line. - 5 This rematch could be expected to decrease the uie ie emeaue ecause i esus i a ow icease comie wi a eucio i comesso essue 9 A aio I aio there would be a reduction of the 4H/U of the highly loaded 3 -

7 - age om =102 eiciecy ocus A •• Ce sage • 1 7 3 5 3 5 -77 98 igue 13 Sage eomace caaceisics (coig E 96 As a result of this series of tests it became clear that the first stage was anemic in terms of flow and work capacity. The attempts (Configs. 3 and 4) to increase the work of the downstream stages did not provide the needed perfor- mance improvements. Thus the replacement of the original first stage by a suitably modified version of the existing research stage became a high priority exercise. 100

- - c 0 E 9 6 - 150 Radius 4 -100 cc 99 2 50 (/o/(1o esig 7 97 93 = 9 ( 0 0 350 300 250 200 150 100 e o( 14 12 10 8 6 4 (inches) 5 9 95 1 Axial lenght, X (io/-/o I (o-o esig (I

igue 1 Meiioa gas a o eeome comessos igue 15 Comesso ma (coig

compressor turbine. However, the shift of the peak efficiency zone of the Aciee age compressor towards choke appeared to be difficult to achieve without adversely -7;1 eiciecy ocus [eiciecy ocus affecting the compressor performance at design speed. Contrary to Config. I the first rotor of Config. 6 was choked at design speed as shown in Fig. 16. Thus, if the first rotor operating condition could be &1 shifted to lie just at the verge-of-choke the overall compressor efficiency could be expected to improve significantly without a significant loss of flow since the 9 movement would be occurring where the characteristic was essentially vertical. In order for this to be accomplished without an increase in overall compressor

pressure ratio the work of one or more of the downstream rotors had to be 9 Downloaded from http://asmedigitalcollection.asme.org/GT/proceedings-pdf/IGT1985/79429/V001T02A015/2477464/v001t02a015-85-igt-41.pdf by guest on 27 September 2021 reduced. Unfortunately, reducing the work in the downstream stages would eicio —es esus have the effect of shifting the peak efficiency zone away from choke at part esig speed. Based on a detailed examination of overall engine performance it was oi concluded that the benefits of a performance improvement near design speed far outweighed any losses due to a deterioration at part speed. Therefore, it was decided to reduce the work capacity of a downstream stage and accept the risk of 1 ia uig a possible reduction in part speed compressor efficiency. Since the third and ie fourth stage rotors were already operating close to choke it was deemed unwise to further reduce their work capability. Thus, the centrifugal stage was the only 9 remaining possiblity and a decision was made to reduce the impeller tip radius. A reduction of the impeller tip radius causes the optimum match between the impeller and diffuser to occur at a higher speed and leads to a performance tr 100 deterioration at low speeds. A reduction in impeller tip radius of 5.08mm or 0.2 97 inches was chosen and the compressor was designated Config. 7. 9 1 99 97 a- 9 — 93 (/a/(/- 7 esig 97 = 99(% esig 93 Config. 5 9 95 100 Coig 7 (7-1/-/o / (rhoVA/do) esig (%

+c 5 igue 17 Comesso ma (coig 7

In additon to the development activities described above there was a continuing campaign to reduce blade and vane tip clearances, eliminate gas path irregularities and improve surface finish everywhere. Performance im- provements in the hot section of the engine also contributed significantly to the successful reduction of the initially high turbine inlet temperature. During the course of the compressor performance development it was found that the first 5 9 95 1 and second stage stator vanes of Config. 7 needed to be modified to improve (/o/(/go esig (% their dynamic characteristics. The spanwise t ,,,,,,/c distribution of the first stage igue 1 Cokig ow coeicie (coigs 7 vane which was originally nearly constant at 0.083, was changed to vary linearly from 0.07 at the hub to 0.103 at the shroud. The second stage stator, which was originally made from strip-stock with a = 0.06 was modified to have a t/c distribution varying linearly from 0.05 at the hub to 0.10 at the shroud. Test Results of Config. 7 The original chord lengths were retained. These changes to the first and second Config. 7, incorporating the reduced radius impeller, was tested one month stators solved the dynamic problem and produced no measurable change in after the test of Config. 6. The compressor map obtained from the test is compressor performance. presented in Fig. 17. As expected, flow and pressure ratio were reduced at part speed and the peak efficiency decreased at 92.3% N 1 . At about 95% the peak CONCLUSIONS efficiencies of Configs. 6 and 7 are equal and the compressor efficiency at design speed and the original design pressure ratio improved by as much as 4.2%. This Following 18 months of engineering effort including 7 configuration corresponded to a 1.4% improvement along the new engine operating line which changes the performance development of the PT6A-65 compressor was had been lowered to reduce turbine inlet temperature. Surprisingly, there was completed successfully and on schedule. The final compressor configuration met little shift of the peak efficiency zone which had been expected to move toward all of the performance and stability requirements. To the end of Dec. 1984, 120 stall due to the reduction in impeller radius. One explanation for this unexpected PT6A-65 engines had been delivered and 250,000 flight hours had been phenomenon could be that the third and/or fourth stage was operating very near accumulated. Throughout this period there has been no reported performance choke in Config. 6 even though the 4)-14., characteristics did not indicate it clearly. problem involving the compressor. In Config. 7 where the diffuser controls choking at part speed the third and It has been demonstrated that simple static pressure measurements can be a fourth stages are constrained to operate at slightly lower flows and hence may be very useful development tool for a small multi-stage axial or axial-centrifugal operating at or near their peak efficiency points. Consequently, even though the compressor. The cost-effective development of a compressor is possible if first and second stages are operating at lower efficiency at part speed in Config. through-flow and meanline performance analysis codes are used extensively with 7, the efficiency of the third and fourth stages could have improved with the the measured static pressures as part of the input data. Needless-to-say sound result that the peak efficiency zone remains essentially the same as in Config. 6. engineering judgement is necessary in the interpretation of the output data. The compressor could not be mapped up to its normal operating pressure ratio at design speed because of turbine temperature limitations. However, a ACKNOWLEDGEMENTS minor extrapolation of the test data revealed that Config. 7 gave a design speed efficiency only 0.3% less than target while exceeding the target flow and The PT6A-65 engine development was assisted financially by the Federal pressure ratio by 0.2% and 2.2%, respectively. Government of Canada. Complete engines with the Config. 7 compressor successfully demonstrated The authors are grateful to Pratt & Whitney Canada Inc. (P&WC) for the expected reduction of turbine inlet temperature without any deterioration in granting permission to present this paper and to the many P&WC Engineering the already satisfactory sfc level. Subsequent ground and flight tests proved that staff members who contributed significantly to the compressor development and the compressor had ample surge margin. Thus, Config. 7 has become the provided the authors with valuable technical advice during the preparation of production standard compressor configuration. this paper.

7 REFERENCES Where,

1. Yoshinaka, T., "Surge Responsibility and Range Characteristics of 1 w dri i I U=1 u Centrifugal Compressors" Proceedings of 1977 Tokyo Joint (I 112 d4 (A + g ) y-1 -1- rl 11 I Congress, 1977. dwi U= Y - 2. Geye, R.P., Budinger, R.E. and Voit, G.H., "Investigation of a (1+1 ) Y-1 High-Pressure-Ratio Eight-Stage Axial-Flow Research Compressor with _ 1 Two Transonic Inlet Stages, II-Preliminary Analysis of Overall Performance", NACA RM E53J06, 1953. d may be calculated using the change of angular momentum of the flow 3. 1 Downloaded from http://asmedigitalcollection.asme.org/GT/proceedings-pdf/IGT1985/79429/V001T02A015/2477464/v001t02a015-85-igt-41.pdf by guest on 27 September 2021 Medieros, A.A., Benser, W.A. and Hatch, J.E., "Analysis of Off-Design d+ across the rotor. Since Performance of a 16-Stage Axial-Flow Compressor with Various Blade Modifications." NACA RM E52K03, 1953. 4. Koyama, S. and Yamaguchi, N., "The Axial Compressor - The Internal U Cx, W I _ U - - tan(3 + tana, 4'1 (A-3 Flow Peeped into from Outside of the Casing -" (in Japanese), Journal of -1 [ U, I U, Cx, JSME, Vol. 74, No. 634, 1971. 1 5. Raw, J.A. and Weir, G.C. "The Prediction of Off-Design Characteristics of U, Cx, Axial and Axial/Centrifugal Compressors", SAE 800628, 1980. dy [ tang, + tana, + 6. Yoshinaka, T. and Leblanc, A.D., "Test Results from an Analytically d+i I Designed Axial Compressor Stage of 1.65:1 pressure Ratio", SAE800629, 1980. (A-4) I [ 1 [ C a/ + aa i 1 Cx, c14)1 U, U, I

APPENDIX The right hand side of Equ. (A-2) may be evaluated as follows: Fig. A - I eiiio of Velocity Triangles and Locations K : is normally greater than unity with the area ratio, (A i /AO] being the The flow coefficient of the 1st stage, +1, and that of the 2nd stage, +11, have dominating factor. the following relationship based on mass conservation in a simplified meanline flow field: F( WI ) : is total density ratio, P,I/P, II, and therefore is positive and less than unity.

dF(wi) : is the rate of change of the total density ratio with respect to the d g1 change of the pressure coefficient, WI, which is negative.

1 1s sage chP1 : is always negative in the present analysis, and moves towards zero 1 11 sage as d01/d+1 increases.

c3

igue A-1 eiio o eociy iages a ocaios

= K( ii (A-I (1 y 3 11 ¢ - Where, 1 5 - 1 1 9 ■-( 7 1 11 Y -1 = 1 y 1 = - 1 K- : approximately constant y - 1 M 2 11 / 7

C : discharge coefficient

M: absolute flow Mach number U2,1 ( + igHl 1 ) IO = : total density ratio U 1 1 y/(y - 1 C I + g1-11

Therefore,

(1 (A- 13 13 011 K [ (11 + 1 1 1 igue A- Sage-o-sage aiaio o ow coeicie aio

8 If the right hand side of Equ. (A-2) is greater than unity, this implies that For the case of the PT6A-65 compressor, typical values of the more the 4 ma — 4 m i range of the flow coefficient of the second stage is greater than influential parameters are: that of the first stage. = 0.3 to 0.55 Fig. A-2 shows the behavior of Equ. (A-2) for several combinations of the F(p) 0.8 most influential parameters. The data indicate that dfii/c141 increases with 1 dtp 2 to 3 1 increasing and decreasing (- ) (negative). An increase of F (p also K 1.3 Downloaded from http://asmedigitalcollection.asme.org/GT/proceedings-pdf/IGT1985/79429/V001T02A015/2477464/v001t02a015-85-igt-41.pdf by guest on 27 September 2021 at Nife 90% N 1 , increases (d411/d41)/K. owee(11/1 itself tends to approach unity as F(y.,) increases, since the increase of F(sp) to unity means a reduction of stage When these are substituted in Equ. (A-2) the resulting values of 11/d41 pressure ratio with the consequence that K, also, moves towards unity. are greater than unity for all conditions of interest.

9