Experimental Investigation of a 4.54-inch- Mean-diameter Three-stage Reverse-flow Reentry Turbine

N64-I0184

D-2076

NASA TN

NASA TECHNICAL NOTE

Z I---.

z

EXPERIMENTAL INVESTIGATION OF A

4.54- IN C H- MEAN- D IAME TE R

THREE-STAGE REVERSE-FLOW

REENTRY TURBINE

by Donald E. Holeski and William T. Win tucky

Lewis Research Center

Cleveland, Ohio

NATIONAL AERONAUTICS AND SPACE ADMINISTRATION - WASHINGTON, D. C. . NOVEMBER 1963

TECHNICAL NOTE D-2076 EXPERIMENTAL INVESTIGATION OF A 4. 54-INCH-MEAN-DIAMETER THREE-STAGE REVERSE-FLOW REENTRY TURBINE By Donald E. Holeski and William T. Wintucky Lewis Research Center Cleveland, Ohio NATIONAL AERONAUTICS AND SPACE ADMINISTRATION EXPERIMENTAL INVESTIGATION OF A 4.54-INCH-MEAN-DIAMETER THiLEE-STAGE REVERSE-FLOW REENTRY TURBINE By Donald E. Holeski and William T. Wintucky SUMMARY A three-stage A.54-inch-mean-diameter reverse-flow reentry turbine was designed and evaluated experimentally in cold air. This turbine was designed for a blade- to jet-speed ratio of 0.244, a total- to static-pressure ratio of i_.5_ and a rotative speed of 22j_36 rpm.

At design speed and pressure ratio_ the equivalent specific work output ob- tained was 38 Btu per pound at a static efficiency of 0.57, which compare to de- sign values of AI.8 and 0.63, respectively. A single-stage test of the turbine showed thatj because of overexpansion_ the first stage produced slightly more than its design work. A study of the weight-flow variations through the turbine affected by leakage flows is also presented.

INTRODUCTION As part of an NASA research program on turbines of high specific work and low weight flow, a three-stage reentry turbine was designed_ fabricated_ and evaluated experimentally in cold air. The turbine had a 4.54-inch mean diameter and was designed for requirements in the general area of small pump drives for rocket propellants. Multistaging was accomplished by passing air through a single rotor three times. This turbine is defined as a reverse-flow type be- cause air movement through the second stage is in the opposite direction to air movement in the first and third stages. This type of a turbine may also be re- ferred to as a cross-flow reentry turbine.

Descriptions and advantages of reentry turbines are given in previous NASA reports (refs. i to _). The following are the primary advantages of reentry turbines. The later stages act to cool the single rotor and_ thus_ to maintain lower rotor-blade operating temperatures in the initial stages than in similar stages in full-admission turbines. Higher inlet temperatures therefore can be used for a given maximum rotor-blade operating temperature, which results in higher power output per pound of fuel expended. Reentry turbines permit longer rotor-blade heights than full-admission turbines for the same applications. The use of a single rotor gives a lighter and less complex rotating assembly.

This report presents information on the turbine design and the results of

evaluating the turbine experimentally with cold air as the driving fluid. The

experimental investigation was conducted by (i) investigating the complete unit

and (2) evaluating the first-stage performance alone to determine the percentage

of the total work being produced by the first stage. A study of the welght-flow

variation through the turbine was also madeto gain knowledge of the internal

flow leakage.

SYMBOLS

A

cross-sectional inlet area, sq ft c rotor-blade chord, in.

gravitational constant, 32.17 ft/sec 2 g 2_ specific work output 3 Btu/ib J mechanical equivalent of heat, 778.16 ft-lb/Btu rotor-blade height, in.

N turbine speedj rpm P absolute pressure_ ib/sq ft R gas constantj 53.35 ft-lb/°R T absolute temperature_ OR U blade velocity, ft/sec V absolute gas velocity_ ft/sec W relative gas velocity, ft/sec w weight flow_ ib/sec c_ absolute gas-flow angle measured from tangential direction, deg relative gas-flow angle into and out of rotor measured from tangential direction_ deg Y ratio of specific heats ratio of inlet total pressure to U.S. standard sea-level pressure_ P'/P* function of Y used in relating weight flow to that using inlet conditions at U.S° standard sea-level atmosphere, T static efficiency, ratio of blade power to ideal blade power based _s on total- to static-pressure ratio squared ratio of critical velocity at turbine inlet to critical _cr s velocity at U.S. standard sea-level temperature, Vcr/_cr speed-work parameter, gj Ah

Um

v blade- to jet-speed ratio, _/2gJ_qid average stage density, ib/cu in.

Pay ¢ ratio of rotor-channel area to sum of rotor-blade cross-sectional area and rotor-channel area Subscripts: cr conditions corresponding to Mach number of i Id ideal m mean-radlus value u tangential component x axial component 0, i, 2,...,i0 station nomenclature, fig. I Superscripts: !

absolute total state w U.S. standard conditions TURBINE DES IGN As mentioned in the INTRODUCTION, this turbine was designed for requirements in the general area of small rocket-propellant pump drives. The following over- all design values, which are corrected to U.S. standard sea-level conditions at the turbine inlet, were chosen:

Equivalent weight flow, wSw_J_cre/6,!b/sec ............... 0.03841

Equivalent blade velocity, um/e-_J_cr,ft/sec ............... 444.7

Preliminary equivalent specific work output_ L_h/0cr _ Btu/ib .... 47.58

Blade- to jet-speed ratio, v ....................... 0.244 Stage speed-work parameter, _ ....................... 0.5 A preliminary design was made by using these values and assuming that all the weight flow passed through each stagej this design resulted in a turbine having three stages with a mean diameter of 4.54 inches. The stage temperature and pressure distributions through the turbine and the velocity diagrams were calculated.

Since leakage flow between stages, which consists of flow through the disk seals and annular flow that is pumped by the rotor, is known to occur_ a simpli- fied estimate of the flow through each stage was made. The amount of leakage flow was assumed to be equal to the total flow that could be passed clrcumferen- tially from one stage to another by using the static density of each stage, the mean rotor veiocity_ and the total cross-sectional area of the annulus. This resulted in a welght-flow variation through each stage. The stator area of each stage was adjusted to be compatible with the weight-flow variation, and no changes were made in the velocity diagrams. The change in stator areas, and, thus_ in weight flow of each stage, resulted in a change in the overall specific work output of the turbine. As a result, the final overall design requirements of the turbine were as follows: Equivalent weight flow, w 8W_Tcr e/5 , ib/sec ............... 0.03841 Equivalent blade velocity, Um/Sw_cr_ ft/sec ............... 444.7 Equivalent rotative speed_ N 8_cr_ rpm ................ 22_436 Equivalent specific work output, _h/0cr _ Btu/ib ............. 41.8 Overall total- to static-pressure ratio, PI/P9 .............. 14.5 Blade- to jet-speed ratio, w ....................... 0.244 Blade height, Z_ in .......................... 0.265 Tip clearance_ in ............................ 0.007 Axial clearance between blade rows, in .................. 0.010 Overall turbine efficiency, qs .................... 0.63 The final stage design characteristics were as follows: Design characteristic Stage First Second Third 0.89 Total to static efficiency 0.67 0.64 Total- to static-pressure ratio 2.04 2.37 3.00 Ratio of stage weight flow to 1.0 0.74 0.90 turbine weight flow Work output_ percent of total 0.38 0.28 0.34 Partial admission losses that were not included in this study can be ex- pected to reduce the stage efficiencies.

In the preliminary design_ the work split among the three stages was assumed to be equal. It can be noted in the preceding table, however, that the final- design work split was unequal because of the weight-flow variation through the turbine.

In addition to the previously mentioned characteristics_ the following as- sumptions and requirements were made: (i) Velocity head at exit of each stage lost In diffusion (2) Total-pressure loss for each stator of A percent (3) Constant hub and tip radii through each blade row (4) Constant blade angles for all stator blades (5) Constant rotor-blade entrance and exit angles The velocity diagrams for the turbine were constructed at the mean radius and appear in figure i. This figure shows that the first-stage stator was de- signed to operate subsonic, but near choking (V/Vcr = 0.918), while the second- and third-stage stators were designed to choke.

Stator The stator blades were designed to be similar to the first-stage stator blades of the turbine described in reference 5, which were designed to turn and accelerate the flow to near sonic velocity in a smooth coverging channel. The critical velocity ratios at the exit of each stator were 0.918_ 0.988_ and 1.082 for the first, second, and third stages, respectively.

The stator-blade profiles and solidities for all three stages were exactly the same; only the number of passages varied. The number of passages was 6j i0_ and 22 for the first_ second_ and third stages_ respectively. The percentage arc of admission was ii.7_ 19.5, and 42.9 for the first_ second_ and third stages 3 respectively. The suction surface of the blade downstream of the throat was a straight lime.

The resulting dimensions and profiles of the stator blades are presented in figure 2(a).

Rotor The rotor-blade design presents a compromise between geometry_ diffusion, and reaction. Since a reverse-flow turbine _as being used_ a primary considera- tion was that the geometry of the blade be symmetrical about a plane that passes through the midchord of the blade parallel to the plane of rotation.

In order to obtain a low total diffusion, a high solidity of 2.63 was used.

Also, by arbitrarily increasing the channel width to 0.053 inch at midchord from

0.0¢9 inch at the leading edge (an increase of approximately 8 percent), it was

assumedthat someof the diffusion on the suction surface would be transferred

to the pressure surface. For a rotor-blade chord of 0.25 inch, 150 rotor blades

were required to give a solidity of 2.65. In figure 2(b) are shownthe dimen-

sions and profiles of the rotor blades.

The rotor blades were designed for constant relative velocity entering and

leaving the blade. The values of relative critical velocity entering and leaving

the first, second, and third stages are, (W/Wcr)2 = (W/Wcr)3 _- 0.538, (W/Wcr)5

(W/Wcr) 6 = 0.586, and (W/Wcr)8 = (W/Wcr)9 . 0.654 (fig. i).

Reentry Ducts The reentry ducts were designed to reduce the flow velocity from the rotor to a low value before it is carried to the stator inlet of the next stage. A development drawing of the turbine at mean diameter showing the reentry ducts is presented in figure 3. The reduction in velocity was accomplished by a transi- tion section that consisted of a 0.3-inch-long straight portion followed by an annular flare of 0.3-inch radius on the inner and outer walls. This design re- sulted in a flow path that increased in height from 0.265 inch at the rotor exit to 1.2 inches in the low-velocity portion of the duct, giving an area ratio of 4.5.

The reentry-duct inlet ends were placed arbitrarily (i) at the same circum- ferential position with respect to the trailing edge of a stator blade at one end of a stator row (position A, fig. 3) and (2) at the opposite end of the stator at a circumferential position corresponding to an extension of the straight portion of the stator-blade suction surface (position B, fig. 3). The two reentry-duct collectors and the exit collector were placed in the same manner.

Because of the reverse-flow design, the flow path to the subsequent stage was very short compared to the crossover-type reentry turbine in which the flow always enters the same side of the rotor. In the subject turbine, the airflow made only one turn of approximately 178 ° per stage, thus making the reentry ducts smaller than those of the crossover-type turbine. The exploded cutaway drawing of the 4.54-inch-mean-diameter three-stage reentry turbine in figure 4 shows the major components.

APPARATUS, INSTRUMENTATION, AND PROCEDURE The apparatus used in the evaluation of the performance of this turbine con- sisted of an inlet and an exhaust systems, a speed-reducing gear box, and an eddy-current-type dynamomenter. Figure 5 shows the experimental turbine instal- lation.

The arrangement of the apparatus is shown schematically in figure 6. High-

pressure dry air was supplied from the laboratory air system. The air was heated

by an electric heater so that the turbine-inlet temperature was 600 ° R and was

filtered to remove dirt particles. After passing through the turbine, the air

was exhausted into the laboratory low-pressure exhaust system. With a fixed

inlet pressure, a remotely controlled valve in the low-pressure exhaust line was

used to maintain the desired pressure ratio across the turbine. The airflow was

measuredby a calibrated flat-plate orifice located in the air-supply line.

The power output of the turbine was absorbed by an eddy-current-type dyna-

mometer that was cradle mounted for torque measurements. The torque-force

measurementwas madewith a calibrated commercial strain-gage load cell.

The rotational speed of the turbine was measuredwith an electronic counter

in conjunction with a magnetic pickup and a shaft-mounted gear.

An inspection of the turbine madeprior to testing showedthat the turbine

b!ading was not fabricated to the desired dimensions. All the stator flow areas

and the rotor flow area were less than the design values. The area measurements

of the first-stage stator were approximately i0 percent less than the design

area. The second- and third-stage stator areas were approximately 7 and 3 per-

cent less than the design values, respectively. The rotor flow area was also

less than design, by approximately 12 percent. Although these area values are

somewhatdifferent from the design values, it was felt that these differences

would not change the major results significantly.

Figure 3 shows the positions of the instrumentation. All pressures were

measuredwith wall taps connected to electrical pressure transducers. Tempera-

ture probes were placed at the meanradius.

All data were recorded by an automatic digital potentiometer and were proc-

essed through an electronic digital computer. Experimental data were taken over

a range of inlet total- to exit static-pressure ratio from approximately 7 to 32.

At each pressure ratio, the turbine rotative speed was varied from 40 to I00 per-

cent of the design equivalent speed in 10-percent increments of speed. Eighty-

percent speed data were not taken, however_ because of a severe vibration in the

apparatus at this speed. The turbine-inlet temperature was approximately 600 ° R,

and the inlet total pressure was approximately 125 pounds per square inch abso-

lute. The friction torque of the bearings and seals was obtained by motoring

the shaft and rotor (rotor in vacuum) over the range of speeds covered in this

investigation.

The turbine was rated on the basis of static efficiency. The inlet total

pressure was calculated from the weight flow, the inlet static pressure, and the

inlet total temperature as follows:

Y

PI = PI + 2-

, ; The single-stage turbine investigation was conducted similarly to that of the three-stage turbine. The exceptions were that (i) an airbrake absorption dynometer was used (described in ref. 6) and (2) only one pressure ratio across the turbine was used_ this being near the pressure ratio across the first stage when the three-stage turbine was tested.

RESULTS AND DISCUSSION The results of this investigation are presented in three sections. The first section_ Overall Performance_ presents the performance results for the complete turbine. The second section, First-Stage Performance_ presents the re- sults of testing just the first stage as well as a statement on the work split among the three stages. The third section_ Weight-Flow Distribution, discusses the weight-flow distribution as a result of an analysis of the leakage flows.

Overall Performance The overall performance of this turbine appears in figures 7 and 8. Fig- ure 7 presents the variation of turbine specific work output Zkh/Scr with total- to static-pressure ratio P_/P9 for lines of constant blade speed. This figure indicates that_ as the total- to static-pressure ratio was increased from 7_ its lowest value, the specific turbine work output increased rapidly at first. With further increases in pressure ratio_ the work output continued to increase_ but not as rapidly; at the higher pressure ratios, the work output reached a limiting value for a particular speed. For design speed (Um/e_J_cr = 444.7 ft/sec), the work output varied from 28.7 Btu per pound at a pressure ratio of 7 to a maximum of 41.3 Btu per pound at a pressure ratio of about 26. Work output remained con- stant at this maximum up to the maximum pressure ratio of 32. A specific work output of 38 Btu per pound was obtained at the design pressure ratio of 14.5, which was 9 percent less than the design specific work output of 41.8.

Figure 8 shows the variation of efficiency _s with blade- to jet-speed ratio for lines of constant blade speed. This figure also shows a dashed curve drawn through the peak efficiency point of each speed line. This curve shows that_ at the design blade- to jet-speed ratio of 0.244_ the peak efficiency ob- tained was 0.57 as compared to the design value of 0.63. Figure 8 also shows that for each speed there was a large variation of efficiency with blade- to jet- speed ratio. The high pressure-ratio data are at the left portion of each curve in figure 8. As the pressure ratio was increased from the value where peak ef- ficiency occurred_ the curves broke away from the peak efficiency curve. This was caused by losses that occurred as the third stage approached limiting load- ing. As the pressure ratio was reduced from the value where peak efficiency oc- curred, the curves again broke away from the peak efficiency curve. This was caused by incidence losses that were becoming significant.

Figure 9 presents the variation of static pressure through the turbine at the different measuring stations (sho_m in fig. 3) for overall pressure ratios of approximately 7_ 9_ 15_ and 32 at design speed. Also shown is the design pressure variation through the turbine. As shown in the figurej the measured

variation of pressure approximated the design variation of pressure. The first

stage showedsomeoverexpansion_ which may have been caused by an area mismatch

due to flow leakage losses from the first stage. The second-stage pressure

variation was as designed. The thlrd-stage variation showedunderexpansion as a

result of the fixed design overall pressure ratio and first-stage overexpansion.

The measured equivalent weight flow of this turbine was 0.03739 pound per

second and was constant over the range of pressure ratios and speeds investi-

gated_ which indicates that the first-stage stator was choked. The first-stage

stator flow coefficient was calculated to be 0.97. The measured weight flow was

approximately 3 percent less than the design weight flow of 0.03841 pound per

second for the first stage_ which was designed to be near choking. As previously

mentioned_ the stator flow areas were less than the design values. Since the

second- and third-stage stator areas were closer to design values than was the

first-stage stator_ choking of the first-stage stator resulted_ although the

second- and third-stage stators were designed for choking.

First-Stage Performance

To determine the percentage of the work being produced by the first stage_

a test of the first stage alone was conducted. Figure I0 presents the perform-

ance results as a plot of static efficiency _s against blade- to jet-speed

ratio w for a pressure ratio of 2.33. The maximum efficiency was 0.595_ which

is less than the design value of 0.69. As mentioned in the DF_SIGN SECTION_ a

portion of this efficiency loss was anticipated because partial-admlssion losses

were not included in the design study. Additional efficiency losses may also

have been due to the smaller than design flow areas that were mentioned in the

section APPARATUS_ INSTRUMENTAT_0N_ ANDPROCEDURE. Data were taken at only one

pressure ratio (2.33) - this being near 2.26, the pressure ratio across the first

stage when the three-stage turbine was tested. With the assumption that the

single-stage turbine efficiency was constant (0.595) for a small change in pres-

sure ratio in the region where the curves of efficiency against blade- to jet-

speed ratio are nearly flat_ the specific work output of the first stage was

computed to be 15.5 Btu per pound at the pressure ratio of 2.26. Although the

work output is less than the design value_ this stage produced 41 percent of the

measured total turbine specific work output of 38.0 Btu per pound at design

speed and pressure ratio. The obtained 41 percent of the measured total specific

work corresponds to a design value of 38 percent.

The second and third stages were not tested as single-stage units because of

the complexity of the inlet and outlet ducts of these stages. From temperature

measurementstaken during the complete-turbine tests_ however_ the indications

were that the specific work output of the third stage was less than that of the

second stage. No attempt was madeto determine the actual work split.

Weight-Flow Distribution

A study was madeto determine the weight-flow distribution through the tur-

bine at design operating conditions. The choking weight flow through each stage

stator was determined experimentally to be 0.0374_ 0.0290, and 0.0318 pound per

second for the first, second_ and third stages, respectively. These weight-flow

values were corrected to the interstage pressures and temperatures that would

exist if the complete turbine were operated at standard sea level inlet condi-

tions. The second- and third-stage stators passed 77.5 and 84.9 percent_ re-

spectively_ of the first-stage flow. As the flow passes through a stag% a por-

tion of the flow is composedof leakage flow through the disk seals_ leakage

flow in the axial clearance spaces between the stators and rotor, and flow that

is pumpedto the next stage outlet. In order to compute these flows the follow-

ing assumptions were made:

(i) For the total flow through both the upstream and the downstreamdisk

sea!s_ the stage outlet flow density was used.

(2) For the flow through the axial clearances and for the rotor pumping

fiow_ the average density of a particular stage was used.

(3) The leakage flow in the axial clearance spaces between the first and

third stages was assumedto be zero.

With these assumptlons_ an estimate of the leakage and pumping flows was madeat

the design operating point. The values of these leakage flows are given in

table I as percent of turblne-inlet flow. The pumping flow for a particular

stage was computedby the use of the equation

w = 12OavZcUm ¢

Circumferential pressure gradients on the rotor_ caused leakage flow from the first and second stages toward the center of the rotor through the disk seals. This flow leaked back into the third stage. This leakage flow (also the leakage flow in the axial clearance spaces) was computed by iterating with as- sumed values of flow coefficients until a balance of the weight-flow distribu- tion through the turbine was obtained. This method resulted in a flow coef- ficient of 0.528 for the disk-seal flow. This appears to be a reasonable value, because the disk seal was a single sharp-edged restriction. The flow coefficient was assumed to be constant for the three stages.

When the annular flow in the axial clearance space was calculated_ a flow coefficient of 1.265 was computed in the manner just described. The rotor_ ro- tating in the same direction as the fiow_ caused the flow coefficient to be greater than i. As mentioned previouslyj no flow was assumed to exist between the first and last stages.

Thus_ it is seen that the leakage flows of this type of turbine play an im- portant part in its performance, and they must be considered in the design. A simple estimate of the flows_ like that made for this turbine deslgn_ appears to be a reasonable first estimate.

i0

SUMMARY OFRESULTS

A cold-air experimental investigation of the 4.54-inch-mean-diameter three-

stage reentry turbine yielded the following results:

i. An equivalent specific work output of 38 Btu per pound and a static ef-

ficiency of 0.$7 were obtained at design speed and pressure ratio. The values

c_rresponded to design values of 41.8 and 0.63, respectively. A maximum equiva-

lent specific work output of 41.3 Btu per _ound was obtained at and above a

total- to static-pressure ratio of 26.

2. Experimental data from the first-stage tests indicate that the first

stage produced 41 percent of the total turbine work output. Interstage tempera-

ture measurementsmadeduring the complete-turbine test indicated that the spe-

cific work of the second stage was greater than that of the third stage. _le

design work output of the first, second, and third stages, respective!y_ was

0.38, 0.28, and 0.34 of the total work output.

3. The measured pressure variation through the turbine approximated the de-

sign variation of pressure at the design operating conditions. The first stage

showedsomeoverexpansion and the third stage someunderexpansion. The pressure

ratio of the second stage was as designed.

4. The results of a weight-flow distribution study through the turbine at

design operating conditions indicated that the second- and third-stage stators

passed 77.2 and 84.9 percent; respectively, of the first-stage flow. The re-

maining flow was composedof internal flow leakage.

Lewis Research Center

National Aeronautics and SpaceAdministration

Cleveland, Ohio, August 8, 1963

REFERENCES

Design and

i. Evans, David G., Guthrie, William D.; and Wasserbauer; Charles A.:

NASA MEMO

Performance of a Six-Stage 8-Inch-Mean-Diameter Reentry Turbine.

X-487_ 1962.

2. Wong_Robert Y._ Darmstadt; David L., and Monroe; Daniel E.: Investigation of

a 4.0-Inch-Mean-Diameter Four-Stage Reentry Turbine for Auxiliary Power Drives. NASA MEM0 X-152; 1980.

3. Evans_ David G.: Design and Cold-Air Evaluation of a Four-Stage 8-Imch-Mean- Diameter Reverse-Flow Reentry Turbine. NASA TMX-720_ 1963.

4. Evans; David G.: Design and Experimental Investigation of a Three-Stage Multiple-Reentry Turbine. NASA MEM0 1-16-59E; 1959.

ii

5. Rohiik, Harold E.: Investigation of Eight-Stage Bleed-Type Ttu_bine for

Hydrogen-Propelled Nuclear Rocket Applications. ! - Design of Turbine and

Experimental Performance of First Two Stages. NASA MEM0 X-475_ 1961.

6. Wong_Robert Y._ and Nusbaum_ William J.: Air-Performance Evaluation of a

4.0-!rich-Mean-Diameter Single-Stage Turbine at Various Inlet Pressures from 0.I_ to 1.88 Atmospheres and Corresponding Reynolds Numbers from 2500 to 50_000. NASA _ D-1515, 1962.

TABLEI. - DISTRIBUTION OF LEAKAGE FLOW

[Percent of inlet flow.]

Stage

First Second Third

Measured stator 10O 77.5 84.9

weight flow

Circumferential flow

pumpedby rotor: 2.6

Out 15.2 6.5

In 2.6 15.2 6.5

Disk-seal flow:

Out 8.0 1.4 0

In 0 0 9.A

Circumferential flow in axial clearance space: 3.9 Out 1.8 0 0 In 3.9 1.8 = 0.162 = 0.918 = 0.456 = 0.886 _3 = 87"990 = O. 301 _3 = 55"44° _5 = O. 449 St at i on _m_3 (a) First stage.

Figure i. - Velocity diagrams for turbine at mean radius.

= 0.162

_5 = 15.27°

¢

_5 : 28.15°

= O. 586 : O. 260 = O. 954 _6 = 87"14° _6 : 55'18° St at ion (b) Second stage.

Figure I. - Continued. Velocity diagrams for turbine at mean radius.

(V_r) = O. 159 _8 = 15.25 ° I 7 P8 : 27.86 °

_ (_I :_o_

-Vcr-8 !_ cr'8 ..... pS_r)8 = 0._ cc9 -- 85.64 ° p9 = 38.S2° 9 V _ - O. 418 cr 9 W_ui) = 0.512 Wc 9 St at ion _8 (c) Third stage.

Figure i. - Concluded. Velocity diagrams for turbine at mean radius.

Stat or-blade coordinates

x, YI? YU _ in.

in.

O. 015 O. 015 .040 •015 .075 • 080 •054 •088 • 120 .044 .088 • 160 .048 •081 • 185 •048 .076 .220 .045 • 260 •04O • 300 •055 Straight line • 340 •026 //'_-- 0. 295" --_ o 25' . Z80 •018 z Rad. j • 420 •009 O. 008" • 0O6 • 466 • 006 (a) Stator blade.

/-Axis of rotation / Rotor-blade coordinates / / X, YL' YU' in.

in. in.

\ vRad.

O. 0040 0 O. 0040 T 0. 049" \\ \ 0. 004" •0198 .010 .0054 .0654 •040 .0402 •1009 .070 .0621 • 1199 • i00 .0738 .1242 • 120 .0767 .1242 • 150 .0767 .0737 •1197 • 150 .0667 •1087 • 170 / u Rad., .0556 .0897 • 190 I 0 •004" I •0395 .0625 •210 0.095 L0.049" • 0045 •0189 • 240 •0040 •0040 .2495 (b) Rotor blade.

Figure 2. - Stator- and rotor-blade geometry and coordinates at mean radius.

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0 •08 .16 .24 . $2 .40 .&8 Blade- to jet-speed ratloj v Figure I0. - Variation of static efficiency with blade- to Jet-speed ratio for first stage at pressure ratio of 2.35.

NASA-Langley, 1963 E-2117 25