Calibration of turbine test rig with impulse turbine at ... · npsarchive 1968 lenzini,m. calibrationofturbinetestrigwith impulseturbineathighpressureratios by martinjosephlenzini

NPS ARCHIVE1968LENZINI, M.

CALIBRATION OF TURBINE TEST RIG WITHIMPULSE TURBINE AT HIGH PRESSURE RATIOS

by

Martin Joseph Lenzini

DUDLEYKNOX LIBRARYJJ^WSTGRADUATESCHOOLMONTERar, CA 93943-5101

| •JHO^ 'U0|>(30I9 .

aaciNia JilHS

UNITED STATESNAVAL POSTGRADUATE SCHOOL

THESIScalibration of turbine t^st rig *vit;i

IKi '^LS£ TURBINE AT HIGH PRiiSSURi; RATIOS

by

i'iartin Joseph Lenzini

June 1968

This document is subject to special export con-trols and each transmittal to foreign governmentor foreign nationals may be made only with priorapproval of the U. S. Naval Postgraduate School.

CALIBRATION OF TURBINE TEST RIG .tflTH

IMPULSE TURBINE AT HIGH PRESSURE RATIOS

by

Martin Joseph LenziniCaptain, United States Marine CorpsB.S., University of New Mexico, 1961

Submitted in partial fulfillment of the

requirements for the degree of

AERONAUTICAL ENGINEER

from the

NAVAL POSTGRADUATE SCHOOLJune 1968

ABSTRACT

The Transonic Turbine Test Rig of the Turbo-Propulsion

Laboratory, Department of Aeronautics, of the Naval Post-

graduate School was designed to investigate the performance

of turbines with transonic or supersonic rotor inlet

velocities. The test rig has provisions for testing single

stage axial turbines at high pressure ratios and at variable

axial and radial clearances. The present study describes

the calibration of the turbine test rig with an impulse

turbine at high pressure ratios. The turbine stage consists

of a double circular-arc rotor with sharp leading edges and

a stator with converging nozzle type blading. The results

of the flow rate calibration and labyrinth seal leakage tests

are described. The instrumentation necessary to separate

rotor and stator losses is also discussed.

DUDLEY KNOX LIBRARYNAVAL POSTGRADUATE SCHOOLMONTEREY, CA 93943-6101

TABLE OF CONTENTS

SECTION PAGE

1. INTRODUCTION 13

2. TURBINE DESCRIPTION 15

3. TEST INSTALLATION 17

k. PLOW NOZZLE CALIBRATION 22

5. PLENUM LABYRINTH SEAL LEAK RATE 2 5

6. ANALYSIS AND DATA REDUCTION 29

6.1 General 29

6.2 Plow Rate Determination 29

6.3 Stator Entrance properties 31

6.^ Stator Discharge properties 31

6.5 Rotor Discharge Properties 37

6.6 performance Parameters ^1

7. DESCRIPTION OF TURBINE TESTS Al4

8. RESULTS AND DISCUSSION OF TURBINE TESTS ... k7

9. RECOMMENDATIONS S5

REFERENCES 88

APPENDIX I: COMPUTER PROGRAMS FOR FLOJ RATE

DETERMINATION AND TURBINE TEST

DATA REDUCTION 39

APPENDIX II: EVALUATION OF THE FORCE ACTING

ON THE STATOR ASSEMBLY BY THE

STATOR DISCHARGE PRESSURE 140

4v J '

LIST 0? ILLUSTRATIONS

FIGURE PAGE

1. Blade Profiles Circular-Arc Rotor

Converging Stator 57

2. Converging Stator (View showing

stator entrance) 58

3. Circular-Arc Rotor (View showing

rotor entrance) 59

4-. Mean Radius Blade profile

Converging Stator 60

5. Mean Radius Blade Profile Circular-Arc

Rotor with Sharp Leading Edges ........ 6l

6. Blade Profile Circular-Arc Rotor

with Blunt Leading Edges 62

7. Blade Profile Contoured Rotor 63

8. Blade profile Converging-Diverging

Stator 6k

9. Piping Installation, Transonic

Turbine Test Rig 65

10. Transonic Turbine Test Rig 66

11. Turbine Blading Arrangement 67

12. Closure Plate Assembly

Transonic Turbine Test Rig 68

13. Closure Plate Calibration Rig (View

showing closure plate assembly setup

for calibration run) 69

14. Closure Plate Assembly Installation


15. Turbine and Shroud Details 71

16. Circular-Arc Rotor and Bearing Assembly

Mounted in Rotor Bearing Stand ... 72

17. Floating Stator Assembly ... 73

18. Flow Nozzle Discharge Coefficient


19. Referred Labyrinth Seal Leak Rate


20. Referred Labyrinth Seal Leak Rate


21. Modified Referred Labyrinth Seal Leak

Rate Transonic Turbine Test Rig 77

22. pressure Survey at Stator Entrance

Pressure Ratio = 2.0 78

23. Velocity Diagram of Turbine Stage 79

24. Thermodynamic process of Fluid

in an Axial Turbine Stage 80

25. Total-Static Efficiency at Various

Axial Clearances 81

26. Referred Moment versus Referred RPM

Pressure Ratio = 2.0 82

27. Efficiency Total-Static Initial

Turbine Tests 83

28. Referred Moment versus Referred RPM

Initial Turbine Tests 84

29. Stator Torque Capsule Variation with

RFM Transonic Turbine Test Rig 85

30. Stator Torque Capsule Variation with

Temperature Transonic Turbira Test Rig 86

31. Stator and Rotor Loss Coefficient versus Hood

Temperature at 13O0O RPM and pressure

Hatio of 2.5 87

TABLE OF SYMBOLS

Latin

A Cross-sectional area (in )

a Plow channel throat diameter (in)

a Speed of sound (ft/sec)

b Blade width (in)

2 2C Conversion factor, 2gjc (ft /sec - °R)

Pc Specific heat at constant pressure (BTU/lb - R)P

D Diameter (in)

m

P Force (lb )

2g Gravitational constant (32.174 lb - ft /lb - sec )

h Blade height (in)

HP Horsepower

h Differential pressure across turbine flow nozzlew (in H

£0)

J Conversion factor (778.16 ft - lb /BTU)

K .Jork coefficient (dimensionless

)

K. Head coefficient (dimensionless)is

Flow nozzle discharge coefficient (dimensionless)n

M Moment (ft - lbf )

M Absolute Mach number (dimensionless)

M Relative Mach number (dimensionless)R

N Rotational speed (rpm)

P Total pressure (psia)

P Static pressure (psia)

R Gas constant (ft - lb ./lb - °R)i m

R Reynolds number (dimensionless)

R Mean radius (in)

r Radius (in)

r Labyrinth pressure ratio (dimensionless

)

r* Theoretical degree of reaction (dimensionless)

s Distance between blades (in)

s Entropy (BTU/lb - °R)m

T Static temperature (°R)

T Total temperature (°R)

t Blade thickness at trailing edge (in)

t Static temperature (°P)

t. Total temperature (°F)

U Peripheral velocity (ft/sec)

V Absolute velocity (ft/sec)

jJ Relative velocity (ft/sec)•

W Flow rate (lb /sec)x m '

Y-, Expansion factor (dimensionless)

Z Number of blades in a row

Greek

CI Absolute floxv discharge angle (degrees)

& Coefficient of thermal expansion of flow nozzle(dimensionless

)

/j Relative flow discharge angle (degree)

*y Ratio of specific heats (dimensionless)

O Referred pressure (dimensionless)

L, Loss coefficient (dimensionless)

Tj Efficiency (dimensionless)

(j Referred temperature (dimensionless)

q Area restriction factor (dimensionless)

10

($> Flo.\r function (diniensionless)

cj) L Referred labyrinth seal leak rate (in )

cJ>LM Modified referred labyrinth seal leak rate (in )

Angular velocity (radians/sec)

Subscripts

a Axial direction

ax Area normal to the axial direction

CI Closure plate

D Dynamometer

E Equivalent thermodynamic property

h Blade hub

h Hood

is Isentropic

L Labyrinth seals

m Mean streamline

n Flow nozzle properties

Stator entrance properties

P Labyrinth plenum properties

R Rotor

REF Referred value

S Stator

t Blade tip

th Theoretical value

u Peripheral direction

1 Stator discharge properties

2 Rotor discharge properties

11

SECTION 1

INTRODUCTION

Turbines for modern gas turbine plants and jet pro-

pulsion units must operate at high pressure ratios. It

is advantageous to use stages with supersonic or transonic

flows in the rotating rows thereby limiting the number of

stages and increasing the specific work output. Although

the efficiency of such stages may prove to be somewhat

lower than that of rotating rows with subsonic flows, they

are desirable for use in low-weight power plants. An

application presently under consideration by NASA is the

use of a single-stage supersonic turbine in a hydrogen-

fueled open-cycle auxiliary space power plant [ll.

Very little quantitative information on supersonic and

transonic turbine performance is available in the literature.

Therefore, a Transonic Turbine Test Rig was built at the

Naval postgraduate School, Monterey, California. The test

rig was designed by Dr. M. H. Vavra of the Department of

Aeronautics to determine the effect of different blading

arrangements on turbine efficiency and to separate the

total losses of the turbine into those of the rotating and

the stationary rows of blades. With the Transonic Turbine

Test Rig, investigation of turbine performance for transonic

and supersonic rotor inlet velocities is possible.

The present study is concerned with the installation

modifications and calibration tests necessary to obtain

meaningful data for transonic or supersonic turbine perform-

ance analysis at high pressure ratios. Initial calibration

13

tests using an impulse turbine with a stator, consisting

of converging nozzle type blading, and a rotor with circu-

lar-arc profiles with sharp leading edges are described.

Several instrumentation difficulties were experienced

during the initial tests which required a number of modi-

fications to the test installation. Tests are described

which were carried out with the turbine after the different

modifications of the Turbine Test Rig.

The author wishes to express his deep appreciation to

Dr. M. H. Vavra for his guidance during the experimental

work and for his help in the reporting of the study.

Thanks are also given to Mr. J. E. Hammer for his generous

assistance during the project.

14

SECTION 2

TURBINE DESCRIPTION

The turbine investigated is a single stage axial flow

turbine of the impulse type which was designed for transonic

rotor inlet velocities. The rotor has double circular-arc

blade profiles with sharp leading edges and is one of three

rotors presently available for the Transonic Turbine Test

Rig, which hereafter is referred to as the TTR. The stator

used during the present turbine tests has converging type

nozzles. Also available is a stator with converging-

diverging type nozzles for supersonic stator discharge

velocities. The three rotors and the two stators are inter-

changable, and any stator-rotor combination can be tested

with the TTR. Of the six available combinations only one

was tested because of time considerations and the delays

because of the TTR modifications. Figures 1, 2, and 3 are

photographs of the stator and rotor of the turbine stage

tested. Scale drawings of the mean radius blade profiles

of tnis stator and rotor are shown in Pigs. 4 and 5»

respectively, pertinent turbine dimensions are listed in

Table I. The blading parameters indicated in Table I are

at the mean radius of the stage.

Another rotor which can be run in the TTR has double

circular-arc blade profiles with blunt leading edges. This

type of rotor, which is shown in Fig. 6, could be used in

high temperature applications where blade cooling is neces-

sary. A third rotor, whose blade profiles are shown in Fig.

L5

7, has blade profiles with gradually changing curvature.

Figure 8 is a photograph of the blade profiles of the con-

verging-diverging stator.

TABLE I

IMPULSE TURBINE DIMENSIONS

Converging Stator and Circular-ArcRotor with Sharp Leading Edges

ITEM SYMBOL STATOR ROTOR

Number of Blades Z 31 60

Blade Height (in) h « 0.690 0.932

Blade tfidth (in) b 0.975 0.750

Hub Radius (in) h 3.895 3.826

Mean Radius (in) Rm 4.240 4.292

Tip Radius (in) Ht

4.585 4.763

Blade Spacing (in) s O.8594 0.444

Trailing EdgeThickness (in) t 0.024 0.020

Throat Diameter (in) a 0.205 0.1313

Throat Area (in2 ) Ath 4.385 7.3^8

2Axial Exit Area (in ) A

ax18.382 25.134

16

SECTION 3

TEST INSTALLATION

The TTR installation and instrumentation has been

described by Commons [2], Therefore, the information pre-

sented here will be concerned with the salient features of

the TTR only. However, modifications made to the TTR for

the impulse turbine tests will be covered in more detail.

The working fluid for the TTR is air which is supplied

by an Allis Chalmers VA 312 Compressor. As shown in Fig. 9>

the supply air enters the turbine test cell through the

inlet valve attached to tank 1 which is manually operated,

and normally in the open position. The turbine inlet valve,

which was originally a manually operated valve, was re-

placed by a remote controlled electrically operated .butter-

fly valve. Both the turbine inlet valve and the exhauster

inlet valve could then be operated simultaneously from the

control room. This change reduces the time spent to set

and maintain a desired pressure ratio if the TTR is operated

with the exhauster.

A scale drawing of the cross-section of the TTR is

shown in Pig. 10. Air enters the floating armature assembly

radially from a plenum which is instrumented with total

temperature and total pressure probes. Labyrinth seals,

with 0.005 inch radial clearance between the armature

assembly and the plenum, limit the leakage flow to about ?

per cent of the turbine flow rate. A conical screen is

fitted in the armature assembly to reduce the possibility of

17

damage to the turbine by foreign objects. The air flows

through the conical screen into the stator plenum which is

instrumented xvith five fixed total pressure protes , one

3-hole survey probe, and two total temperature probes. In

addition a so-called bullet probe is installed at the down-

stream end of the armature assembly, which measures total

pressure and total temperature. The arrangement of these

probes is shown in Fig. 11.

The closure plate assembly, also shown in Fig. 11, was

completely redesigned for the impulse turbine tests. The

moment on the closure plate was obtained from six equally

spaced torque flexures. The flexures, which are 0.025 inch

thick and extend radially from an inner support to an outer

ring, are equipped with two strain gages. The strain gages

are arranged to measure the bending moments on the flexures

in the axial plane. The inner support is fastened to the

closure plate force flexure. The force flexure consists of

four webs of 0.080 inch thickness which are also instrument-

ed with strain gages. These gages measure the bending

moments applied to the flexure by the sxial force acting

on the closure plate. The signals from both sets of strain

gages was read on a Daytronic model 700 strain g&ge digital

indicator. Figure 12 is a photograph of the closure plate

assembly which shows the force and torque flexures. The

closure plate was calibrated on a specially built calibra-

tion rig by applying known forces and moments with various

combinations of weights. The arrangement of the calibration

rig is shown in Fig. 13. Figure 14 gives two views of the

18

closure plate assembly installed in the TTR.

Stator hub and tip static pressures are measured in

the cavities between the stator assembly and the closure

plate, and between the stator assembly and shroud, respec-

tively. These static ports are shown in Fig. 11. The

outer shroud is instrumented with seven static pressure

taps, P through p ? , spaced at 0.25 inch intervals from15 -*-

about the mid-rotor plane to the downstream end of the

shroud. Tne last four static taps determine the shroud

pressures needed in the momentum analysis of the inlet

guide vanes.

There are seven shroud inserts available with differ-

ent inside diameters. All inserts are cylindrical with the

exception of two, one with a five degree slant and the

other with a ten degree slant, for tapered rotor blade tips.

Only the cylindrical shroud insert with an inside diameter

of 9.5^6 inches was used in the present tests. The arrange-

ment of the shroud, the shroud insert, and the seven static

pressure taps is shown in Fig. 15.

Different radial tip clearances are obtained with a

particular shroud insert by reducing the rotor diameter.

The radial clearance used for the present tests was 0.010

inch.

The turbine rotor, shown in Fig. 11, is supported by

two sets of precision ball bearings which are lubricated by

oil mist. Two photographs of the rotor in the bearing

stand are shown in Fig. 16. The axial clearance betxêen

the stator and rotor is varied by sliding the rotor bearing

19

assembly in the bearing stand. The minimum axial clearance

is limited by the distance by which the closure plate

extends beyond the trailing edges of the inlet guide vanes.

Operating at 20,000 rpm and pressure ratios of k or more,

the minimum axial clearance is about 0.1 inch.

The stator assembly, shown in Fig. 17 » is supported

by flexures which permit measuring of the reactions of the

stator discharge flow by means of reluctance type force

gages. One of the flexures was instrumented with strain

gages for the final three runs of the impulse turbine. The

results obtained with the strain gage and the reluctance

capsule measurements are discussed in Section 8.

An air dynamometer capable of absorbing 200 HP at

20,000 rpm is used to measure the turbine torque. The

torque is measured by a reluctance type force capsule

which is attached to a 20 inch long lever arm that is

fitted to the dynamometer housing. The force gage limits

the angular rotation of the dynamometer housing to about

0.25 degree. Originally a so-called direct reading spring

capsule, which turns by about 30 degrees at maximum torque,

was used as a bearing housing for the dynamometer. At

the small rotation of the dynamometer housing it was

believed that the coil spring which serves as the measuring

element of the capsule would not affect the readings of the

reluctance gage. This assumption was proven false and it

was necessary to remove the coil spring (Section 8). Simi-

lar to the rotor bearings, the dynamometer bearings are

lubricated by oil mist.

20

All pressures are measured by mercury manometers ex-

cept the pressure difference across the flow nozzle, which

is read on a water U-tube manometer. All temperatures are

measured with Iron-Cons tantan thermocouples using an ice

bath as a reference. The hood temperature, to which

reference is made in this study, was measured by a thermo-

couple located in the plastic casing of the stator torque

capsule. This casing shields the thermocouple from the

flow of air in the hood. The location of the thermocouple

is shown in Fig. 17.

21

SECTION 4

PLOW NOZZLE CALIBRATION

The turbine flow nozzle installation and the calibra-

tion techniques used are described by Eckert [3l. Early

tests by Eckert indicated that the nozzle n ischarge co-

efficient was a function of nozzle supply pressure. Fur-

ther investigation by Naviaux at nozzle supply pressures of

20, 22, and 24 psia showed that the nozzle coefficient

was a function of Reynolds number only [4]]. The latter

result was obtained with the equations used by Eckert and

an expansion factor Y, for nozzles instead of sharp edge

orifices in accordance with the ASME Power Test Codes [5].

Nozzle supply pressures for the turbine performance

tests normally vary between 30 and 42 psia. Because of

past experience with the calibration of the TTR flow

nozzle it was decided to carry out additional tests at

supply pressures of 24, 29, 34, 39 » and 42 psia to verify

the results which Naviaux obtained at lower pressures.

The test data were reduced by using the IBM 3oO Computer

of the Naval Postgraduate School. The data reduction program

was similar to that of Naviaux with the exception that the

specific gravities of mercury and water in the manometer

were corrected for the temperatures in the control room

and that only the flange taps of the sharp edge orifice in

the calibration pipe were used. The specific gravities

of water and mercury as a function of temperature, and the

standard conversion factors used for data reduction, were

22

obtained from the International Critical Tables [61.

The program description and the reduced data are given in

Appendix I. In Pig. 18 the results of the nozzle calibra-

tion tests are plotted as a function of Reynolds number.

Above a Reynolds number of 7(10-5) these results differ by

less tnan 1 per cent from those found by Naviaux. Below

a Reynolds number of 7(10 ) Naviaux' s nozzle coefficient

decreases sharply to a value of 1.002 at a Reynolds number

of 4.2(10^). The nozzle coefficients obtained from the

present tests decrease less sharply below a Reynolds number

of 7(10 ) resulting in differences of between 2 and k per

cent at Reynolds numbers between 4.2(10 ) and 6(10 ). Since

the present study was concerned with flows over a wider

range of Reynolds numbers, considerably more data were

taken at Reynolds numbers below 6(10 ).

An analytical expression for the nozzle discharge

coefficient as a function of Reynolds number was obtained

by using the method of least squares. This expression,

which represents a fourth order polynomial approximation

to the reduced data and is also plotted in Fig. 18, is

-1 -7 -13 2Kn = 9.32928x10 + 4.268322x10 Re - 6.151495x10 R

e

+ 3*895006xl0~19

R<:i

3- 9.138062xl0"

2R (1)

where

K = nozzle discharge coefficientn

Re = Reynolds number referred to nozzle diameter

23

The maximum deviation between reduced data points and the

analytical curve in the operating range of the impulse

turbine between Reynolds numbers of ^(10 ) and 8(10 ) is

0.3 per cent.

2^

SECTION 5

PLENUM LABYRINTH SEAL LEAK RATE

The method used for the plenum labyrinth seal leak

tests and the associated instrumentation are described by

Eckert [3l. Although the measuring techniques remained

unchanged, the tests were carried out over a wider range

of operating conditions.

The purpose of these tests was to find a simple ana-

lytical expression for the determination of the leak rate

as a function of the pressure ratio across the labyrinth.

This expression should cover the entire operating range of

the TTR. To accomplish this goal two series of labyrinth

leak tests were performed. The so-called hooded configura-

tion of the TTR was used for both test series. Figure 9

represents a schematic of the installation of the TTR and

shows the exhauster with the necessary piping for hooded

operation. Labyrinth leak rates at pressure ratios be-

tween 1 and 6 were measured in both series of tests.

The first test series was performed before the impulse

turbine was tested. Prom these tests a referred leak rate

as a function of labyrinth pressure ratio was determined

which was based on the actual labyrinth flow rate obtained

from the square-edged orifice data for different conditions

2in the plenum. The referred leak rate (in ) is

^=^y%r < 2 >

25

where

rf - labyrinth flow rate (Ibm/sec)

2P. as inlet plenum total pressure (lb /in )

T = inlet plenum total temperature ( R)tp

R - gas constant for air (ft-lb /lbm-°R)

2g = gravitational constant (lbm-ft/lb -sec )

The results of the first series of labyrinth leak tests

and the analytical expression derived from them are shown in

Fig. 19. It can be seen that the labyrinth leak rate be-

comes choked at a pressure ratio of about 3« The referred

labyrinth leak rate for the choked condition is equal to

0.073. This value corresponds to a flow rate between O.Co

lbm/sec and 0.1 lbm/sec, depending on inlet total conditions.

After the first twelve test runs of the impulse turbine,

inconsistent values of turbine flow rates were noted.

Analysis of these runs indicated that the inconsistency

was probably due to errors in the labyrinth leak rate

(Section 8). Re-evaluating the data from the first series

of labyrinth leak tests, it was noticed that the hood tem-

perature from run to run did not vary by more than 8 F and

remained practically constant during each run. However,

during normal operation of the TTR with the turbine installed,

the hood temperature is a function of the turbine discharge

temperature and varies with turbine speed. Furthermore, the

variation of hood temperature at different pressure ratios

can be as much as 90 F. The difference in the range of

hood temperatures during the first labyrinth tests and

26

during normal operations of the TTR made it necessary to

perform a second series of leak tests. For this series a

two inch pipe with a gate valve was connected from tank 1

to the TTR hood. Since the air temperature in tank 1 can

be changed by about 100°F, it was possible to control the

hood temperature.

Three tests were performed at hood temperatures of

59 F, 91 F and 116 F, each for the whole range of pressure

ratios of the TTR. The results of these tests are given

in Fig. 20, which shows the referred leak rate of Eq. (2)

as a function of labyrinth pressure ratio. From this

figure it is seen that the labyrinth leak rates depend on

the hood temperature. Above a pressure ratio of 3 the

referred leak rate varies by 12 per cent for a change in

hood temperature of 57 F. Since this temperature can varyo

by 90 F during the turbine tests, an expression for the

leak rate was empirically obtained which is independent

of hood temperature and inlet plenum conditions. This

expression is obtained by multiplying Eq. (2) with a cor-

rection factor which depends on the hood temperature and

the inlet plenum total temperature. This so-called modified

referred leak rate <±> was found to be

^LM L(3)

wnere

4> = expression from Eq. (2)

2?

t s= inlet plenum total temperature (°p)

t = hood temperature ( P)h

Figure 21 shows <$» as a function of labyrinth pressure

ratio from the tests at the three above-mentioned hood

temperatures. An analytical expression forCjX-™ as a

function of labyrinth pressure ratio was obtained from the

method of least squares, by approximating the test data by

a fifth order polynomial,, The resulting expression is

d> = -0.1004586 + 0.2122579r - 0.108l851r2

+

0.0276576^ - o,003^89933r + (4)

0.0001726733r

where

r = labyrinth pressure ratio = p. /p22

P = hood static pressure (lbf/ln )

Equation (4) is plotted in Pig. 21. The maximum deviation

between the test data and the analytical curve is 4 per cent

for the operating range (1.0<r<4„0) of the impulse turbine.

The IBM 360 Computer was used to reduce the data for

both series of labyrinth leak tests. The computer program

for the data reduction is presented in Appendix I together

with the output for the second series of tests.

28

SECTION 6

ANALYSIS AND DATA REDUCTION

6.1 General

The TTR is instrumented to obtain data for a one-

dimensional performance analysis of single stage axial

turbines. It is assumed that steady axisymmetric flow

conditions exist at the entrance and exit of the blade

rows and that the flow on the mean stream surface is re-

presentative of the flow through the whole stage. The

mean stream surface is assumed to exist at the mean radius

R = _E "(5)

where

R = radius of stator blade tip (in)

R = radius of stator blade hub (in)

The TTR data were analyzed on the IBM 3&0 Computer

at the Naval Postgraduate School. The computer program

is described in Appendix I, which give samples of print-

outs for runs 32, 33 a^ 3^.

6.2 Flow Rate Determination

The flow rate through the turbine is the difference of

the flow rate through the flow nozzle and the labyrinth

leak rate,

. . . (6)a = *r - *fTn L

29

where

irf = nozzle flow rate (lbm/sec)

dj - labyrinth leak rate (lbm/sec)

The nozzle flow rate is obtained from the nozzle flow

equation given by Commons (Reference 2, p. ^6). However,

the nozzle discharge coefficient K is determined from

£q» (1). Moreover, the constant in Commons' equation was

found to be incorrect. The correct equation for the deter-

mination of U isn

where

D = nozzle throat diameter (in)n '

a = coefficient of thermal expansion of the flown nozzle (dimensionless

)

K = nozzle discharge coefficient from Eq. (1)(dimensionless

)

Y = expansion factor (dimensionless)1

h = differential pressure across the pressure tapsw at 68° F (in H

20)

P = absolute static pressure at upstream pressurenoz tap (lb/in2 )

T = temperature at upstream pressure tap (°R)

The leakage flow rate through the plenum labyrinths is

obtained from

'L'», = ,

LM -^ ^^ (8)T - T-^ ~

Ttp

30

1.2

where ($> is the modified referred labyrinth leak rateLI'i

obtained from Eq. (^-)

6. 3 Stator Entrance properties

The total pressure p. at the stator entrance is takento

as the average of the data obtained with the five fixed

total pressure probes. A radial survey with the 3-hole

flow probe at this location indicated a maximum variation

of 0.75 per cent in total pressure from stator hub to stator

tip. The results of this pressure survey are presented in

Fig. 22. The total temperature T. is obtained from twoto

Temperature-Kiel probes.

6. k Stator Discharge properties

The stator discharge properties can be obtained from

trie momentum and moment of momentum equations applied to

the fluid in the stator assembly. These two fundamental

equations yield the axial and peripheral velocities from

which the other discharge properties can be derived. In

addition, the axial velocity component can be obtained

also from the equation of continuity applied to the

stator exit. The equations used in the stator analysis

are presented here without derivation, since they are

given by Messegee [7~1. From the theorem of angular

mementum

Vul= 12(Ks

+ Kcl )s/;' H

»l (9)

31

where

V , = peripheral component of absolute velocity atstator exit (ft/sec)

M = moment acting on stator assembly, measured bys a reluctance gage (ft-lb )

M , = moment acting on closure plate (ft-lb )

H - mean radius at stator discharge (in)ml

The two values of the average axial velocity component

obtained from the momentum and continuity equations are

used to establish a parabolic change of the pressure at the

stator discharge between the measured pressures at hub and

tip such that both methods yield the same results. These

calculations are carried out by an iteration procedure of

the computer program. However, the resulting pressure

distribution cannot be verified experimentally. The

stator exit pressure distribution is first assumed to be

linear between the stator hub and tip pressures. Prom the

momentum equation

Rti

V = 6/i[Fs

+ Fol- F - 2n / p^drl (10)

where

V = axial velocity component at stator dischargeal (ft/sec)

P = force acting on stator assembly, measured bys reluctance gage (lb )

F = force acting on stator assembly by closure plate,°1 measured by strain gages (lb )

P - sum of pressure forces acting on stator assemblyless force due to the stator discharge pressure(lb

f)

32

P = static oressure at radius r at stator exit1 (lb

f/in2 )

The last term of Sq. (10) is then evaluated by assum-

ing that p varies parabolically from hub to tip. A

factor € is introduced so that the shape of the parabola

can be changed to satisfy continuity considerations. prom

the derivation presented In Appendix II

HtlTT,

2n Plrdr = -Fhl

Rill

5 p

(1 -i €)R* + H*,^, - (2+€)R :

tl ti ill til

(11)

3 ti(2 + €)Hti " R

tiRhi ~ & + € > Hhi

where

P = hub static pressure at stator discharge (lb /in)hi i

Pt -, = tip static pressure at stator discharge (lb /in)

The integral of Eq. (10) can be represented by an

average stator discharge pressure p_ , multiplied by thelav

stator exit axial area. Using Eq. (11),

hilav

(1 + € )Rtl + Rtl Rhl - (2 +e )Rjhi

BtlR<hi

(12)

tl< 2 +€ > fl

ti - Htl\l - (1 + €)Khi

Rti - R

hi

33

From the continuity equation

Val

='to

CAlFlav /

1-J1 **&2R2 Vul - T

1

to 112

- Vul

/ _

(13)

where

1A-,= effective axial flow area at stator exit (in )

2 2= conversion factor, 2gjc (ft /sec - °R)

p

= specific heat at constant pressure (BTU/lbm-°R)P

The axial velocity component obtained from Eqs . (10)

and (13) varies directly with the stator discharge pressure

Since this pressure is a function of the factor 6 , an in-

crease or decrease in 6 will increase or decrease the

axial velocity, respectively. The solutions of these

equations are matched by varying € until the velocity

calculated with Eq. (10) equals the velocity calculated

from Eq. (13). This iteration is possible because the

value of Eq. (10) changes more rapidly for a change in

6 than the value of Eq. (13). The absolute velocity at

the stator exit is then

2 9 i

1 L al ul J (1*0

The static temperature T, at the stator exit is found from

the energy equation, or

1 to

V,

2gjc(15)

P

^

From Fig. 23, which represents a velocity diagram of a

turbine stage, it can be seen that the angle of the abso-

lute flow at the stator discharge is

ax

= Tan'1

(Vul/Val ) (16)

Further, the relative velocity W- at the rotor inlet has

an axial component

J = V (17)al al

and a peripheral component

W , = V , - U. (18)ul ul 1

where, with N representing the rotor speed in rpm,

U -NnRml

(19)

Thus, the relative velocity is

"l-p'.J+lJ]* (20)

The angle of the relative flow at the rotor inlet is

&- Tan_1 iVV < 21 >

The speed of sound of the air at the stator exit is

ax

= [7gRT ]* (22)

where V is the ratio of the specific heats of the working

fluid. The absolute and relative ftach numbers of the flow

35

at the s tat or discharge are

Mx

= V1/a

1(23)

M^ = Vax(24)

In accordance with Pig. 2^, the stator loss coefficient

L is defined as3 s

t( T1 ;

Tlis) _ < T1 - Tlis)bs = " ATlls <Tto -Tlls )

(25)

where

"lisxto| Pto

T,. = uA^\y ^)

Also, from Fig. 24,

2

t>s v 2

lth

where

2

LthV = 2gjc (Tto - T ) (28)

p zo lis

The stator efficiency is

V, - i - £ s ( 29)

The so-called flow function ($> is given by Vavra (Refer-

ence 8, pt. I, p. C24) as

th to

36

where A = stator throat area given in Table I (in ).th

For isentropic conditions the flow functioncjx is obtained1 s

from

4>. =IS to) to

(3D

where p., = stator throat static pressure (lb /in ).vi 1 I

The stator throat pressure is assumed to equal P untillav

the flow through the stator becomes choked. For choked

conditions the pressure ratio P., /P in Eq. (31) is taken

th to

as the critical pressure ratio, which for air is equal to

0.5283. A stator blockage factor can now be defined by

£-*'** (32)

Therefore, r represents that percentage of throat area

which would be necessary to pass the flow if the expansion

process through the stator were frictionless

.

6. 5 Rotor Discharge Properties

The rotor discharge properties are obtained by the

application of the moment of momentum equation, the continu-

ity equation, and the energy equation to the fluid passing

through the rotor. The flow in the rotor will be treated

with respect to a relative coordinate system. In this

manner the fundamental laws, applied to the rotating row of

blades, will yield discharge properties analogous to those

obtained for the stator.

3?

Prom the moment of momentum equation the peripheral

component of the absolute discharge velocity is

^1 12MD S

u2 ' \2 ul \2*(33)

where

M = moment acting on the dynamometer measured by a" reluctance gage (ft-lb^)

R = mean radius at rotor discharge (in)m2

The peripheral component of the relative velocity is

tf « = V - Uu2 u2 2(3<0

with

"2

" u^ (35)

Introducing the so-called equivalent temperature T

as defined by Vavra (Reference 8, pt. Ill, p. G^), the

energy equation for relative flows becomes

E

LE ~ ^1 + 2gTc~ 2gjc '

X2 2gJc

P p p

(36)

Using the energy equation in this form with the contin-

uity equation, the static temperature at the rotor dis-

charge is found as

T_ =

2 2

y- 1 ft*i- fcy

•2 /tf „2

W R / u2 - T.

i _,2

- 1PgAg \2gjc

V

(37)

38

where

2P = hood static pressure (lb /in )

2A = effective axial flow area at rotor discharge (in )

From Fig. 24, the total temperature at the rotor discharge

is

t. = t - At, (38)t2 to ri

The temperature drop AT , is proportional to the work

generated by the turbine stage or

At = —2- (39)vJ WO J

where

CO = rotational speed of rotor (rad/sec)

From Euler's turbine equation, AT is alsow

At = x ul 2 u2(4o)

In accordance with Fig. 24, the absolute and relative

velocities at the rotor discharge are

V2

= C< Tt2 " T2> 2eJc p]5 (*Ds

rf = [(T - T )2gjc ]* (42)2 Ji <c p

The axial velocity components of V and rf are

V = w = [V* - V *]* (43)a2 a2 u 2 U2"

1

39

Prom Pig. 23, the angles of the absolute and relative veloc-

ities at the rotor discharge are

a2= Tan" 1

(Vu2/Va2 ) (U)

P2= Tan"

1(Wu2/Wa2 ) (45)

.

With the equivalent temperature of Eq. (12), the rotor loss

coefficient- is obtained from

t = 2 " 21s(46)bR T

E " T2is

with

To, = T,

p, \î I p, \21i218 ElP

fil, .,,

lavr = iJr 7 (47)

where

p = total equivalent pressure at the stator dis-El

2charge (lb /in ). From Pig. 24, the rotor loss coefficient

of Eq. (46) is also

£ =1-2 (48)

" W2th

where

">2th " 2SJcp< Tl " T21s>

+ Wl

+ U2 " U

l " ^*p«l-T2i«' (49)

The rotor efficiency is

^-i-^ (50)

40

6,6 performance Parameters

For the evaluating of the overall performance of a

turbine stage it is advantageous to use dimensionless

coefficients. The performance parameters presented in

this section are those given by Vavra [9].

The overall stage efficiency is the ratio of the work

generated by the turbine stage and the isentropic enthalpy

drop across the turbine from the total conditions at the

stator inlet to the static conditions at the rotor dis-

charge. Therefore, the so-called total-static efficiency

is obtained from

nD co

At., wc j

V--^--k^~ (51)

where, as shown by Pig. 24,

P., \7-1

z

p..7]

toj

AT.s

= TtQ

[l - |^- 7 ] (52)

AT. can be expressed also byis

C2

At = —^— (53)is 2gjc

P

where G is the theoretical velocity obtained by an isen-o

tropic expansion from P to p .

to 2

The theoretical degree of reaction r* is that fraction

of the isentropic enthalpy drop of the turbine stage which

is used up by the rotating row of blades. It is a measure

41

for the acceleration of the relative flow in the rotor.

From Fig. 24

r* 1 - lth(5*0

Using Eqs. (26), (28), (52) and (53) the degree of reaction

at the hub or the tip of the blading can be expressed by

7 - 1

r* = fa/p, iV-i

— (55)

\*2

where p» = p. . or p (lbf/in ).

The isentropic head coefficient K, is used to estimateis

the number of stages necessary to handle a given isentropic

enthalpy drop at a given speed u . It is defined as

Kis ,Un

(56)

The work coefficient K is a measure of the actual work

that the stage generates per unit mass of fluid at a given

speed u, , or

At4

U,/2gJc(57)

The peripheral speed U was selected to make K and1 is

K dimensionless since it is usually a fixed quantity deter-

mined by rotor stress considerations.

42

Referred values of flow rate, rotational speed, dyna-

mometer moment, and horsepower are obtained by using the

NASA reference system. They are defined by

(58)

with

"ref o

N«t— = NREF /JF (59)

Mdbe?-»d/S (60)

_ HPMDW

*bf"S^"550S^ (61)

e=Tto/5!8.4 (62)

S= P+Vl*.7 (63)to

*3

SECTION 7

DESCRIPTION OF TURBINE TESTS

During the present study the TTR was operated for 191

hours, 8^ of which were used for calibration tests with the

impulse turbine installed. The initial tests consisted of

the runs conducted before the second series of labyrinth

leak tests, and the final tests consisted of three turbine

test runs that followed these labyrinth tests.

During the first run of the impulse turbine foreign

object damage to the circular-arc rotor blading was en-

countered. The damaged rotor, however, could be saved by

cutting back the leading edge of the blade row by 0.125

inch and shaping each blade as shown in Fig. 6. Another

circular-arc rotor with sharp leading edges was installed,

and the tests were resumed.

Tne next five runs were carried out witnout the ex-

hauster system at a pressure ratio of 2.0 and different

axial clearance A X between the stator and the rotor. The

turbine was tested at A* equal to 0.200, 0.250, 0.265,

0.300 and 0.350 inch. As stated in Section 3 the radial

rotor tip clearance was 0.010 inch for all runs. Data

were taken at rotor speeds between 1^,000 and 20,000 rpm,

where the lower speed was imposed by the maximum torque

absorption of the dynamometer.

After the optimum clearance A -£ was determined, the

exhauster was installed for so-called hooded operation of

the TTR. Tests were conducted at turbine pressure ratios

kk

of 1.5, 2.0, 2.5, 3.0, 3.5» and 4.0 at speeds up to 20,000

rpm. Several runs were conducted at pressure ratios of 2.0

and 3»0 to insure that consistent results could be obtained.

The minimum speed possible with the presently installed

dynamometer varies with pressure ratio. The minimum speeds

for the above-listed pressure ratios were 9,000; 11,000;

12,500; 14,000; 15,000 and 16,900 rpm, respectively.

A particular turbine pressure ratio can be set by

different combinations of stator inlet pressure and hood

pressure. However, to adopt a standard procedure, the

stator inlet pressure was kept at 5 inches of mercury above

the atmosphere, and the pressure in the exhauster was

varied to obtain the desired pressure ratio up to a value

of about 2.7. For higher pressure ratios the maximum

vacuum of about 17 in. Hg was maintained and the stator

inlet pressure was increased.

From the initial test results, discrepancies were

found in the turbine flow rates and the loss coefficients

of stator and rotor. As discussed in Section 8, the second

series of labyrinth leak tests was then undertaken to in-

vestigate the leakage flow problem. The unrealistic loss

coefficients could be attributed to inaccurate measurement

of the torque which is exssrted on the stator assembly.

Inaccurate stator hub and tip static pressures were

thought to be a secondary cause for this discrepancy.

For the final three runs the reluctance type force

capsule used for measuring the stator torque was disconnected

45

and one of the horizontal torque flexure shown in Fig. 1?

was instrumented with strain gages to measure the torque of

the stator assembly c Additionally the static taps for the

measuring of the hub and tip pressures at the stator dis-

charge were modified and static taps were arranged to

determine the pressures at the hub and tip radius of the

stator throat section as shown in Pig. 15.

The final tests were conducted with the hooded con-

figuration of the TTR. For two of the three runs data were

recorded at various speeds between the minimum possible

and 20,000 rpm at pressure ratios of 2.0 and 2.5. The

third run was carried out at a fixed speed of 13,080 rpm

and a constant pressure ratio of 2.5o For this run the

stator inlet total temperature was varied between 135 and

I65 F. Appendix I lists the raw data that were recorded

for each run.

k6

SECTION 8

RESULTS AND DISCUSSION OF TURBINE TESTS

Figure 25 gives the measured total-static turbine

efficiencies as a function of the referred speed at differ-

ent values of the axial clearance Ax. It is seen that

the maximum efficiency of 83. per cent was obtained for

an axial clearance of 0.250 inch.

Increasing Ax to 0.265 and 0.30 inch produced optimum

efficiencies of 82.2 per cent and 82.8 per cent, respec-

elyj indicating that the values obtained for Ax = O.265

inch might be doubtful. At a reduced axial clearance of

0.20 inch the optimum efficiency is about 82.2 per cent,

equal to that obtained for Ax = 0.35 inch.

The efficiency, as defined by Eq. (51 )> depends on

the dynamometer torque MD and the mass flow rate A, for

given values of turbine pressure ratio p. /p , inlet totaluo 2

temperature T. , and rotational speed N. The influence of

dynamometer torque can be seen by the plot of referred

dynamometer moment against referred rotor speed of Fig. 26.

Figure 26 shows these data for values of AX of 0.250,

O.265 and 0.300 inch. The graph shows that the values of

referred dynamometer moment at Ax = 0.250 inch and

Ax = 0.300 inch lie on the same curve, whereas the data

for Ax = O.265 inch form a curve that is parallel to but

below the curve for the other clearances. Therefore, it

can be concluded that the low efficiency obtained with

Ax = 0.265 inch was due to low values of measured

4?

dynamometer torque. During the tests the dynamometer seemed

to be functioning normally, and no reason could be found to

explain the lower readings. However, during the next

several runs it was noticed that the dynamometer readings

would sometimes fluctuate by 10 or more counts as the temper-

ature of the dynamometer housing increased. It was found

that the fluctuations were due to expansion and contraction

with temperature of the coil spring that was located inside

the capsule which served as the bearing housing of the

dynamometer. After removing this spring consistent values

of dynamometer torque were obtained. The results of the

runs with the hood attached with Ax = 0.250 inch indicated

that there were inconsistencies in the data necessary for

the calculation of the efficiencies and the stator and

rotor loss coefficients. These discrepancies will be

discussed by comparing the data of runs 21, 23 and 24

carried out with the hood, with the data obtained from run

20 where the turbine discharged into the atmosphere. Runs

20 and 21 were carried out at a pressure ratio of 2.0, and

runs 23 and 24 at a pressure ratio of 3.0.

The efficiency as a function of head coefficient K.is

for the four runs is shown in Fig. 27. It can be noted

that the efficiency was different for the runs at equal

pressure ratios. Moreover, the maximum efficiency was

obtained at a value of K. of about 3.7 for runs 20, 21,is

and 23 whereas for run 24 the maximum efficiency occurred

at Kis

=-- 4.3.

48

Figure 28 shows the referred dynamometer moment as a

function of referred speed. It is evident from this graph

that the difference in efficiencies at a pressure ratio of

3.0 was due to the higher values of the referred dyna-

mometer moment obtained during run 2k, The non-linear

shape of the referred moment curve accounts for the

different value of K at which the maximum efficiency foris

run 2k was obtained.

During later runs it was noticed that fluctuations

occurred in the dynamometer moment readings. Examination

of the reluctance gage showed that a lead from a cannon

plug to the gage had broken within the insulation. This

faulty lead may have affected the dynamometer readings

during run 2k.

The referred moments for runs 20 and 21 plotted against

referred speed lie on the same curve. Equal referred

moments for a given referred speed and pressure ratio indi-

cate that the difference in efficiency must be due to

differences in mass flow rates. Since it was believed that

errors in turbine flow rate were due to wrong values of

labyrinth leakage, the second series of labyrinth leak tests

was undertaken as discussed in Section 5. As stated earlier

it was found from these tests that the labyrinth leak rate

is a function of hood temperature. However, the new

labyrinth leak rates used for the reduction of the data

from runs 20 and 21 did not account for the difference in

efficiency of 1.7 per cent since the variation in hood

temperature between the two runs was only 12°p. Without

/+9

further testing it is not possible to explain the differences

in the efficiencies of Pig. 26.

Equation (25), (27), (46) and (48) show that the stator

and rotor loss coefficients depend on the discharge proper-

ties after the rows of blades and can be obtained from the

calculated velocities., These velocities are obtained with

the methods explained in Section 6. The peripheral com-

ponent of the absolute velocity after the stator is deter-

mined primarily from the stator torque measurements. It

is about 3 to k times larger than the axial velocity

component. Thus a variation in peripheral velocity in-

fluences the losses more than an equal percentage variation

in axial velocity. Figure 29 shows the stator torque read-

ing as a function of speed for runs 20, 21, 23, 24 and 25.

All the runs clearly indicate a decrease of the order of

15 per cent in the stator moment as the speed is increased

from the minimum rotor speed to 20,000 rpm. It can be

shown that a 3.5 Pe ** cent variation in the stator moment

will change the rotor loss coefficient by about 0.10,

hence it can be concluded that the measured stator moments

are responsible for the inconsistent values of the losses.

Therefore, it was decided to monitor the read-out of the

stator torque capsule during the second series of labyrinth

leak tests to determine if a change in hood temperature

would Influence its reading. Since a closure plate was

placed over the stator discharge during these tests, any

variation of the stator torque capsule reading from its

calibration setting had to occur because of different

50

thermal expansion of the capsule and the frame to which the

capsule is attached. Figure 30 is a plot of the torque

capsule readings for different hood temperatures. The

latter are given in millivolts above the electrical read-out

of the instrument at the calibration temperature. The

calibration temperature, which changed for each run, was

the hood temperature at which the capsule was set to zero.

Figure 30 shows that the read-outs of the stator torque

capsule are strongly affected by the hood temperature.

The read-outs varied by 10 to 15 per cent of full scale

read-out over a 56°F temperature range. To separate the

overall losses into stator and rotor losses with a suffi-

cient degree of accuracy, the variation of the stator torque

must be less than 2.5 Per cent of full scale read-out. The

axial stator force, which was monitored also during these

tests, varied by less than 1 per cent of full scale read-

out in the same temperature range. Since both reluctance

capsules are identical except for their operating range,

it was concluded that the large variation in the stator

torque capsule read-out was due to the thermal expansion

of the frame to which the capsule is fastened. This con-

clusion is supported by Fig. 1? since the U-shaped frame

to which the capsule is attached is made of aluminum

whose coefficient of thermal expansion is at least twice

that of the capsule and its attachment rods.

For the final series of tests one of the horizontal

torque flexures shown in Fig. 17 was instrumented with

strain gages to measure the stator torque, and the

51

reluctance torque capsule was removed. Additionally, as

shown in Pig. 15 » the outer diameter of the closure plate

rim was reduced by about 0.030 inch to avoid impingement of

flow on the rim which could falsify the readings of the

static hub pressure. This modification was necessary

since the closure plate assembly could not be centered

perfectly with the stator assembly. Moreover, as shown in

Fig. 15 » a cylindrical shim was placed behind the upstream

end of the shroud insert which blocked the upper portion

of the stator tip static port, reducing its radial width

from O.O69 inch to about 0.025 inch, to ensure a more

accurate measurement of the stator tip pressure. From

iSq. (10) it is evident that correct measurements of the hub

and tip pressures are necessary to obtain accurate values

of the axial velocity component after the stator.

Results of tests indicated that the readings obtained

by the strain gages attached to the torque flexure were

influenced by the hood temperature also. Figure 31 shows

the stator and rotor loss coefficients obtained from the

data of a run at a constant rotor speed and constant

pressure ratio. In Fig. 31 these loss coefficients are

shown as functions of the measured hood temperature. It

is seen that reasonable values of the loss coefficients

were obtained only if the hood temperature did not differ

by more than about 2°F from the temperature at which the

torque flexure strain gages were calibrated. During this

test it was noticed also that the temperature difference

along the flexure was about 70°F, which made it impossible

52

to compensate the strain gage circuit for temperature. Thus

it can be concluded that accurate measurements of the

stator moment are not possible with the torque flexure or

the reluctance capsule as presently installed in the TTR.

However, it is felt that accurate measurements can be

obtained with the reluctance capsule if it is relocated as

recommended in Section 9.

The change in the axial velocity due to the modifica-

tions to the stator hub and tip static pressure ports was

negligible. It was found that the tip pressure decreased

by about 2 per cent and that the hub pressure remained

unchanged from values recorded for similar runs during

earlier turbine tests.

As stated in Section 7 the stator throat was instru-

mented with static pressure taps at the hub and tip radii.

The throat hub static pressure was found to be the same as

the hub pressure measured in the gap between the closure

plate and the stator hub. Whether this condition truly

occurs or whether it occurs because of a leak in the

measuring line can be verified only by additional tests.

The throat tip pressures measured at an overall turbine

pressure ratio of 2.5 were 15 per cent higher than the

theoretical pressure for choked conditions. The theoretical

critical pressure ratio for air is 1.89 and that obtained

from the measured stator throat tip pressure was 1.60.

This indicates that either the flow is not choked at the

tip at a overall turbine pressure ratio of 2.5 or that

the pressure tap is located upstream of the actual stator

throat. Tests at higher pressure ratios would show

whether the last-mentioned condition exists

.

5*

SECTION 9

RECOMMENDATIONS

The rotor and stator losses cannot be separated

accurately with the present instrumentation of the TTR

because of the difficulties associated with the measure-

ment of the moments that act on the stator assembly. It

is felt, however, that accurate measurements of the stator

torque are possible if the force capsule is arranged near

the back strut of the cradle that supports the stator.

The capsule should be mounted vertically with one end

connected to an arm attached to the cylindrical inlet pipe

and the other end to a steel frame which is bolted to the

cradle. Using steel for the frame reduces the differential

thermal expansion of the frame and the capsule. To reduce

temperature effects further, an enclosure should be built

around the frame and the capsule into which a small amount

of ambient air would be blown from the atmosphere to keep

the capsule and frame at constant temperature, A similar

arrangement has been successful in reducing the temperature

effects on the dynamometer force capsule (Reference 2, p. 33)

Further tests should be carried out at a number of

turbine pressure ratios between 2*0 and ^.0 to determine

whether accurate measurements of static pressure can be

obtained from the stator throat hub and tip taps as present-

ly installed, or whether these taps need to be relocated.

It is suspected that the static pressure line to the stator

55

throat hub tap has become disconnected in the cavity

between the closure plate assembly and the stator hub.

This possibility should be investigated before further

tests are carried out.

56

FIGURE I

BLADE PROFILESCIRCULAR- ARC ROTORCONVERGING STATOR

57

** s

FIGURE 2

CONVERGING STATOR(VIEW SHOWING STATOR ENTRANCE)

53

v^ '** '/'I

0ff\\\\\

FIGURE 3

CIRCULAR-ARC ROTOR(VIEW SHOWING ROTOR ENTRANCE)

59

LUq:

LU

a: q:

a. o

£*<_J

co c5

Q LU< >or z^ o<LU

60

0.020

0.75

0.020

0.1314

s = 0.4440

FIGURE 5

MEAN RADIUS BLADE PROFILECIRCULAR-ARC ROTOR WITH SHARP LEADING EDGES

6l

FIGURE 6

BLADE PROFILECIRCULAR-ARC ROTOR

WITH BLUNT LEADING EDGES

62

I fc I

FIGURE 7

BLADE PROFILE

CONTOURED ROTOR

63

FIGURE 8

BLADE PROFILE

CONVERGING— DIVERGING STATOR

64

!si

65

§Q:

&&

£S3

18$

66

67

• »

1 ^# i

• î * Vv

•

9

FIGURE 12

CLOSURE PLATE ASSEMBLYTRANSONIC TURBINE TEST RIG

63

€

UJ

*9

View a

View b

FIGURE 14

CLOSURE PLATE ASSEMBLY INSTALLATION

TRANSONIC TURBINE TEST RIG

7C

Stator Tip

Throat

Shroud Insert

Shroud

P P P P P16 17 18 19 20 21

010

9.l70Dia.

Stator Hub-Throat Pressure

Ax+

Shim Installed For

Final Turbine Tests

Portion Removed For

Fina I Turbine Tests

7.652Dia

9.546 Dia.

Rotor

FIGURE 15

TURBINE AND SHROUD DETAILS

71

View a

View b

FIGURE !6

CIRCULAR -ARC ROTOR AND BEARING ASSEMBLYMOUNTED IN ROTOR BEARING STAND

n

73

7'4

I'llft 1J1VU Mt/37 HlNIUAGt/7 a3Hi/3J3i/

_ I —J h<0 8 v~

75

76

77

JTfO-0*QUf

CD

X

fO

O £fO w

UJa:

3COCOUJcr

<i-o

CMI-UJ

CVJ

CM

UJOz<I-

LU

or

o o

CO oUJ 1-

<1-

cr <3 o:

CD >-UJ

U. UJ rr> 3CO

_> COCO UJ

UJQ.

olO

o4.30 3.9

( Nl) smavd

V- 1

a:

COCOUJa:0-

?8

UJ

CO

UJzm

toCM

U.UJtr

3CD 2z *

o<

oo-JUJ>

79

FIGURE 24THERMODYNAMIC PROCESS OF FLUID IN AN

AXIAL TURBINE STAGE

80

1

1

%- s> k Co

&

31

52

00

4

AON3IOIdd3J J

CD

CVJ

O

lf> <1-C/>

UJ1-

K 1

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REFERENCES

1. Vanco, Mo r , "Thermodynamic and Turbine Character-• istics of Hydrogen-Fueled Open-Cycle Auxiliary Spacepower Systems," NASA TM X-1337, 196?

.

2. Commons, p M. , "Instrumentation of the TransonicTurbine Test Rig to Determine the performance of Tur-bine Inlet Guide Vanes through the Application of theMomentum and Moment of Momentum Equations," NavalPostgraduate School Thesis, September 1967*

3. Eckert, R H. , "Determination of Flow Rates, TransonicTurbine Test Rig," Naval Postgraduate School TN 66T-1,January 19 66.

k. Naviaux, J. C. , "Transonic Turbine Test Rig ExhausterSystem Tests and Tests of a Reaction Turbine," NavalPostgraduate School Thesis, December 1966, p. 68 c

5. Flow Measurement , Chap Q 4, Part 5» Supplement to ASMEpower Test Codes, ASME, New York, N Y„ » 1959? P« 7 4.

6. National Research Council, International CriticalTables , McGraw Hill, 1928.

7. Messegee, J. A., "Influence of Axial and RadialClearance on the performance of a Turbine Stage withBlunt-Edge Non-Twisted Blades," Naval PostgraduateSchool Thesis, September 1967, pp. 112-119,

8. Vavra, M. H. , Problems of Fluid Mechanics in RadialTurbomachines, pts I, II, III and IV, "Von KarmanInstitute Course Note 55a," Rhode-Saint-Genese

,

Belgium, Von Karman Institute for Fluid Dynamics,March 1965.

9

.

Vavra , M . H . , Aero-Thermodynamics and Flow in Turbo -

machines , John tfiley and Sons, Inc., New York, N Y ,

I960, pp. 418-^38.

88

APPENDIX I

COMPUTER PROGRAMS FOR FLOW RATE DETERMINATION

AND TURBINE TEST DATA REDUCTION

A. Program FLOCAL .

This program calculates the flow nozzle discharge

coefficient of the Transonic Turbine Test Rig by comparing

the flow through the nozzle with that through a standard

ASME square-edge orifice e The inputs for this program are:

CardNo. Format Units Fortran Description

1 13 L Number of runs to beprocessed.

2 13 N Number of* data points ina given run. Entries oncards 2 through 11 arerepeated for each run.

Barometric pressure.

Temperature of Hg. ofbarometer

Control room temperature.

Differential pressureacross flow nozzle, mul-tiple entries.

Reference for nozzlepressure, multiple entries.

Flow nozzle upstream staticpressure, multiple entries.

Flow nozzle temperature,multiple entries.

Orifice temperature atupstream pressure tap,multiple entries.

8F10.*4- in.Hg. PFL Orifice upstream staticpressure, multiple entries.

89

3 3F10.4 i n . Hg

.

PBAR

°F TCL

°F TCR

4 8F10.4 ln.Hgb DH

5 8F10.4 in.Hg. PATM

6 ano.4 in. ;ig. PNOZ

7 8F1C.^ m.v. TNOZ

8 8F10.4 m.v. TTD1

10 8F10.4 in.Hg. PREF

11 8F10.4 cm.H2

DPFL

Reference for orificepressure, multiple entries

Pressure difference acrossorifice with flange taps,multiple entries

.

B. program LABLEK ,

This program calculates the labyrinth leak rate of the

Transonic Turbine Test Rig. The referred labyrinth leakage

is computed and compared with the values obtained from the

analytical expression of Section 5. The variation between

these two values in percentage is included as output . The

inputs for this program are

:

CardNo. Format

1 13

2 F10A

12

5

6

7

13

2F10.4

8F10.4

8F10.^

8F10.^

8F10.il-

Units

in.Hg.

in.Hg.

in.Hg.

in.Hg.

m.v.

Fortran Description

L Number of runs to beprocessed*

PBAR Barometric pressure.

N Number of data points inthe given run. Entries oncards 2 through 12 arerepeated for each run.

NRUN Run number.

TGL Temperature of Hg. ofbarometer.

TCR Control room temperature

PATM Reference of labyrinthplenum pressure and hoodpressure, multiple entries.

PSPL Labyrinth plenum totalpressure, multiple entries.

PHD Hood static pressure,multiple entries.

TTPLD Labyrinth plenum totaltemperature, multipleentries.

90

8 8F10.^ m.v. TTD1

9 8F1C.4 m.v. THD

10 8F10.4 in. He. PFL

11 8F10c4 In. He. PREF

12 8F10.4 om,H2

H'rfFL

Total temperature at up-stream pressure tap of theflow orifice, multipleentries.

Hood temperature, multipleentries.

Orifice upstream staticpressure, multiple entries.

Reference for orificepressure, multiple entries.

Pressure difference acrossorifice with flange taps,multiple entries.

C. Program TTRSS .

This program reduces the data obtained with the Tran-

sonic Turbine Test Rig. The program consists of an executive

routine and 10 subroutines. The executive routine provides

the calling sequence for the subroutines. This sequence is:

1. INPUT. This subroutine reads the input test data.

The data items which have multiple entries are so

indicated in the item description, all others are

single entry items. Each test run may consist of a

maximum of 50 data points. The input data consists of

the following items:

Units Fortran Description

MM Number of runs to beprocessed

NRUN Run number. Entrieson cards 2 through 37are repeated for eachrun.

In. AXCLR Axial clearance.

CardNo. Format

1 110

2 110

2F10.^

91

5

8

9-15

16

17

18

19

20

21

22

F10.4

110

2P10.^

8F10.4

in. RADCR

in.Hg. PBA.R

N

o?

8F10.4 in.Hg.

8P10.^ in.Hg.

8P10.4 in.Hg.

TCL

°F TCH

in.Hg. PREF2

8F10.4 in.Hg. PTPL

8F10.4 in.Hg. P15-P21

8F10.4 in.H2 DH

8P10.^ in.Hg. PATM

PNOZ

8F10.4 in.Hg. PSPL

PHUB

8F10.^ in.Hg. PTIP

PHD

Radial clearance.

Barometric pressure.

Number of data pointsin the given run.

Temperature of Hg ofbarometer.

Control roomtemperature

Reference for statorplenum pressure andshroud pressures,multiple entries.

Stator plenum totalpressure, multipleentries.

Shroud pressures,multiple entries.

Differential pressureacross the flow nozzle,multiple entries.

Reference for thenozzle pressure, laby-rinth plenum pressure,stator hub and tippressures and the hoodpressure, multipleentries

.

Flow nozzle upstreamstatic pressure,multiple entries.

Labyrinth plenum totalpressure, multipleentries.

Stator hub staticpressure, multipleentries.

Stator tip staticpressure, multipleentries.

Hood static pressure,multiple entries.

92

23 8F10.4 m.v.

2Ur 8F10.^ m.v.

25 8F10.4 m.v.

26 8F10.4- m.v.

27 8F10.4 RPM

28 SF10A counts

29 8F10.^ counts

30 3F10.^ counts

31 8F10.^ counts

32 8F10.^ counts

33 10X,4A2

3^ 15X,3A2

35

36

F5.3

II

in.

TNOZ Flow nozzle temperature,multiple entries.

TTPLD Labyrinth plenum totaltemperature, multipleentries.

TTPL Stator plenum totaltemperature, multipleentries.

THD Hood temperature

,

multiple entries.

RPM Turbine rotationalspeed, multiple entries.

AXIL Stator assembly axialforce, multiple entries.

TORQR Stator assembly torque,multiple entries.

DINAR Dynamometer torque,multiple entries.

CLAXIL Closure plate force,multiple entries.

CLTRQR Closure plate torque,multiple entries.

DATE Month/Day/Year.

TTYPEB I (circular-arc rotorwith sharp leadingedges) or II (circular-arc rotor with bluntleading edges).

STATOR I (converging stator)or II (converging-diverging stator).

METHOD MF (Val determinedusing momentum andcontinuity) or CF (Valdetermined usingcontinuity alone).

RMEAN Stator mean radius.

J Number of pressureratios tested in givenrun.

93

37 "(215) NPTS(K) First data point atparticular pressureratio.

NPTSS(lv) Last data point atparticular pressureratio. These twoentries repeat inpairs, one pair foreach pressure ratiotested in given run.

2. SETCON. This subroutine consists of all the constant

factors used in the data reduction calculations.

3. CONVERT. This subroutine converts the units of the

input data into a single system compatible with the

equations of Section 6.

4. PLORAT. This subroutine computes the turbine flow

rate using the equations of Subsection 6.2.

5. STATOR. This subroutine uses the equation of Sub-

section S.k to calculate the stator discharge properties.

Subroutine MOMENT is called from STATOR when using

momentum and continuity to compute the axial velocity

component.

6. MOMENT. This subroutine determines the axial com-

ponent of absolute velocity by the application of the

momentum equation to the fluid within the stator assembly,

7. ROTOR. This subroutine computes the rotor discharge

properties using the equations of Subsection 6.5.

8. PERFRM. This subroutine computes the performance

parameters and the referred quantities of Subsection 6.6.

9. OUTPTA. This subroutine gives a detailed printed

output consising of the stator and rotor discharge

properties and the performance parameters.

94

10. OUTPUT. This subroutine prints the turbine per

formance parameters in report form.

95

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98

FLOW NOZZLE CALIBRATION FOR THE TRANSONIC TUPBINE TEST PIG

TEST SERIES

1

ATMOS. PRESS,

14.766

POINT

1

234c

6789

10

NOZZLF SUPPLY PRESSURE

39.19

DATA POINTS GAMMA

10 1.4

FLCW RATE DISCHARGE RFYNOLOS(L8M/SEC) COEFFICIENT NUMBER

4.5470 1.04339 12494^6.004.1810 1.04332 1150119. CO3.7625 1.03974 1036495.373.3435 1.C4034 921833.002.9460 1.03806 812751.442.6033 1.04049 718826.872.2851 1.03916 631762.941.8747 1.03327 519366.691.4961 1.02885 415375.441.0637 1.00591 295950.94

TEST SERIES

2

ATMOS. PRESS.

14.767

DATA POINTS

10

NOZZLE SUPPLY PRESSURE

41.95

GAMMA

1.4

POINT FLOW RATE(LBM/SEC)

DISCHARGECOEFFICIENT

REYNOLDSNUMBER

123456789

10

4.66794.24423.88773.41272.88292.55492.28681.95061.54511.0399

1.037961.038501.037341.040241.038161.C33021.032171.027201.012200.98741

1284592.0011675C8.CD1G692C4.00939151.62793601.87703979.37630379.37538148.44426719.87287806.31

99

TEST SERIES N0Z7LE SUPPLY PRESSURE

3 34.30

ATMOS. PRESS. DATA POINTS GAMMA

14.735 2 1.4

POINT FLOW RATE (DISCHARGE OFYNOLDS(LBM/SEC) COEFFICIENT NUMBER

1 4.2561 1.04044 1182130. OC2 3.9094 1.04132 1085137.00

TEST SERIES N0Z7LE SUPPLY PRESSURE

4 34.25


14.700 10 1.4

POINT FLOW RATE DISCHARGE REYNOLDS(LBM/SEC) COEFFICIENT NUMBER

1 3.4087 1.03932 944578.002 3.0387 1.04327 842217.063 2.5675 1.04209 711613.374 2.3839 1.04038 660725.445 2.1594 1.03792 598632.696 1.9635 1.03493 544912.757 1.7285 1.03260 479896.508 1.4205 1.02484 395303.199 1.1926 1.02393 332649.62

10 0.9363 1.00979 261676.87

TEST SERIES NOZZLE SUPPLY PRESSURE

5 29.35


14.682j

9 1,4


1 3.9243 1.04078 1094621.002 3.5504 1.04636 9<>07^1.623 3.1456 1.04380 878523.814 2.7296 1.04415 762850.125 2.3435 1.04390 655627.626 2.0941 1.C4002 586375.507 1.8079 1.03963 506764.068 1.4423 1.02582 404797.879 0.9900 1.02611 278830.06

100

TEST SERIES NOZZLE SUPPLY PRESSURE6 24.39

ATMOS. PRESS. DATA POINTS GAMMA14.603 4 U4

P0INT PLOW RATE DISCHARGE REYNOLDS(LBM/SEC) COEFFICIENT NOMRFP

\ 3.5521 1.04187 994823. »1

1 \ % \^¥Z 1.04286 872236.444 2.8472 1.04237 796562181

TEST SERIES NOZZLE SUPPLY PRESSURE7 24.32


14.574 6 1.4


1 '2.5438 1.04131 711068.442 2.2157 1.03960 619473.123 1.9886 1.03360 556218.694 1.7660 1.03475 493973.375 1.5451 1.03297 432556.756 1.2198 1.03049 342052.44

TEST SERIES NOZZLE SUPPLY PRESSURE

8 41.58

ATMOS. PRESS.

14.654

POINT

1

23456789

1011

DATA POINTS GAMMA

11 1.4

FLOW RATE DISCHARGE REYNOLDS(LBM/SEC) COEFFICIENT NUMBER

4.6614 1.04521 1204436.004.3785 1.04539 1216127.004.0059 1.04457 1112648.003.6474 1.04790 1013286.123.2227 1.0472* 896041.562.8031 1.04634 779872.372.5465 1/048 55 708955.312.2738 1.04442 633425.871.9652 1.03927 547810.061.5387 1.C29Q1 429389.371.1206 1.01809 313372.31

101

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139

APPENDIX II

EVALUATION OF THE FORCE ACTING ON THE STATOR ASSEMBLY

BY THE STATOR DISCHARGE PRESSURE

The pressure at the stator discharge p varies between

the pressure P at the hub and the pressure Pât the tip*hi * tl *

This pressure is not necessarily a linear function of radius

r # In the present study the pressure p will be assumed

to vary parabolically from p, , at R to P^, at R suchhi lil tl tl

that at the mean radius R the pressure is (1 + € ) P ,

ml ml

where p = (P + P )/2. The value of the factor € willml hi tl

be determined by comparing momentum results with those from

the continuity equation.

Let

pl " P

hl+ 1OTT lr-Rhl> + AP = P

hl S<-V + AP (64)

where

P - P

S = JSk *i (65)»tl " Rhl

and

Ap= A + A,r + A r2

(66)o 1 2

The constant factors A , A_ , and A are obtained from the

conditions Ap = at r = R , AP = at r = Rfc

, and

AP = P -, at r = R = (R. . + R, , )/2. Thusml mi tl Til"

140

= A + A, R + A„R '

o 1 hi 2 lil

= A + A. Rtn + A R,.o T. tl 2 tl

€P„. = A + A R + A^Rml 1 ml 2 ml

The above simultaneous equations are solved "by Cramer*

s

Rule, with

2i a - r, -

nl hi

1 R R. ?;

tl tl

1 R . R fml ml

= î-HhiJ^-\i«Bti +.%a ,

"+a

hiatl 3

W" h HM1 = (S

hl+ H

tl)/2

c = (Hti- Bhl ,Ci(V+

^ti)2 -^ B

hi+ H

ti)2 + H

hiRtl ]

c = - i (Htl- ^)3 (6?)

Thus

A = 1/C

Rhl \l

fltl hi

e pmlV y

€ Pml Âl RhlHtl

with Eq. (67)

ml (Htl- wwith Eq. {6?) t in like manner

(68)

A. =^ Pml (Rtl +

.

Rhl)

< Rtl " Bhi>2

(69)

141

4 €PA2 = "

ml

(Rtl " V (70)

Substituting Eqs. (68), (69), and (70) into Eq. (66)

AP =7^—V-T2- [

" Rhi

Rti + (Rti - Rhi) r - ^2] (7i)

.

(Rtl - "hi'

The force on the stator assembly due to the statordischarge pressure is

Rtl

ill

with Eqs. (64) and (71)

(72)

sD

«tl

2nJ ['(P,

R il:,

:

.a)r,-

. , 2r '

-1- /\pr j dr

hi

Rtl

= - (Phl" SBhl)(^-I^) + fSCEt

3.8h3] + a„ /APrdr

ahi

with Eq, (65)

'-D-" (p

hi-V H^J'î-^'^ê^^ - ^+

2n4 €

P

mi

(Rtl * R

hl>'

3. - 4- ~ 4-'

R R< Btf- Bh!)

i n ^ RUtlW )

Htl-Rhl" Rhl Htl 2 h (Bhl+Rtl ) 3 " 5

142

Simplifying

*sD - 5PhltRtl

+ Rhl

Rtl - 2R

hl^ * 3Ptl^2Rtl " **!!*«. " ^

Uiti " hi;

./ith p = (P +p )/2, the final expression for the forceml hi tl

FsD acting on the stator assembly by the pressure at the

stator exit is

Rtl

FSE = 2- / Plr dr " îCU + € >B* + R

hlRtl - (2 + € )B^]

Rhl

(73)

•iPtlL( 2^)VVtf (1 + 0finl ]

143

INITIAL DISTRIBUTION LIST

Copies

1. Defense Documentation Center 20Cameron StationAlexandria, Virginia 22314-

2

.

LibraryNaval Postgraduate School 2

Monterey, California 939^0

3. Commander, Naval Air Systems Command 1

Navy DepartmentWashington, D.C. 20360

4. Commandant of the Marine Corps (Code A03C) 1Headquarters, U.S. Marine CorpsWashington, D.C. 20380

5. Commander, Naval Ship Systems Command 1Navy DepartmentWashington, D.C. 20360

6. professor M. H. Vavra 3Department of AeronauticsNaval Postgraduate SchoolMonterey, California 939^0

7. Chairman, Department of Aeronautics 1

Naval Postgraduate SchoolMonterey, California 939^0

8. Professor A.. E. Fuhs 1

Department of AeronauticsNaval postgraduate SchoolMonterey, California 939^0

9. Professor R. D. Zucker 1

Department of AeronauticsNaval Postgraduate SchoolMonterey, California 939^0

10. Dr. E. S. Lamar (Code 03C) 1

Chief Scientist, Research and TechnologyNaval Air Systems CommandNavy DepartmentWashington, D.C. 20390

11. Commander 1

Naval Ordnance Systems CommandNavy DepartmentWashington, D.C. 20390

144

12. Dr. 0. H. JohnsonNaval Air Systems Command (Code 330 B)Navy DepartmentWashington, D.C. 20360

13. head, Department of EngineeringNaval AcademyAnnapolis, Maryland 21^02

14. SuperintendentNaval AcademyAnnapolis, Maryland 21402

15. Mr. I. Silver, Code 330propulsion AdministratorResearch and TechnologyNaval Air Systems CommandNavy DepartmentWashington, D.C. 20390

16. Office of Naval ResearchAir programs OfficeNavy DepartmentWashington, D,C. 20360

17. Capt. A. Eodnaruk, USNNaval Ship Systems Command (Code 6l40)Navy DepartmentWashington, D. C. 20360

10. Office of Naval Research (r-wjer Branch)Attn: Mr. J. L. Patton, Jr.Navy DepartmentJashington, D. C. 20360

19. Dr. F. I. TanczosTechnical Director, Research and TechnologyNaval Air Systems CommandNavy DepartmentWashington, D. C. 20390

20. Captain"

Tartin J. Lenzini100 Lakewood 'lace[ighland Park, Illinois 60035

145

UN CLASSIFIEDSecurity Classification

DOCUMENT CONTROL DATA -R&DSecurity classification of title, body of abstract and indexing annotation must be entered when the overall report is c fossified)

\ originating ACTIVITY (Corporate author)

Naval Postgraduate SchoolMonterey, California 9394C

2a. REPORT SECURITY CLASSIFICATION

UNCLASSIFIED2b. GROUP

3 REPORT TITLE

CALIBRATION OF TURBINE TEST RIG rflTH

IMPULSE TURBINE AT rilGH PRESSURE RATIOS

4 DESCRIPTIVE N O T E S (Type of report and, inclusive dates)

Thesis5 AUTHOR(S> (First name, middle initial, last name)

Martin J. Lenzini, Captain, U. S. Marine Corps

6 REPOR T DATE

Jane 19olla. TOTAL NO. OF PAGES

1447b. NO. OF RE FS

8a. CONTRACT OR GRANT NO.

6. PROJEC T NO.

9a. ORIGINATOR'S REPORT NUMBER(S)

t\AjSLCKMJâJ( &C&A-

9b. OTHER REPORT NO(S) (Any other numbers that may be assignedthis report)

T^j ^ 1 'r-irnM—

II. SUPPLEMENTARY NOTES 12. SPONSORING MILITARY ACTIVITY

Naval Postgraduate SchoolMonterey, California 9394c

I 3. ABSTR AC T

The Transonic Turbine Test Rig of t

Laboratory, Department of Aeronautics, oSchool was designed to investigate the ptransonic or supersonic rotor inlet veloprovisions for testing single stage axiaratios and at variable axial and radialstudy describes the calibration of the t

impulse turbine at nigh pressure ratios.of a double circular-arc rotor with sharwith converging nozzle type blading. Thcalibration and labyrinth seal leakage t

instrumentation necessary to separate roalso discussed.

he Turbo-Propulsionf the Naval Postgraduateerformance of turbines withcities. The test rig has1 turbines at high pressureclearances. The presenturbine test rig witn an

Tne turbine stage consistsp leading edges and a statore results of t.ie flow rateests are described. Tnetor and stator losses is

DD, F

.r.,1473 ,PAGE "40 V 65

S/N 0101 -807-681 1

147UNCLASSIFIEDSecurity Classification

A-31408

UNCLASSIFIEDSecurity Classification

KEY WORDS

I ,ix

ilse Turbine

Transonic Turbine

fruisonic Turtine Test ji t

LINK A LINK B LINK C

ROLE AT

-L

DD .?„~.1473 back.

• /N 0101-0. 07-6471 1^8 UNCLASSIFIEDSecurity Classificntion

thesL527

Calibration of turbine test i

3 2768 002 12055 2DUDLEY KNOX LIBRARY C.J

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Calibration of turbine test rig with impulse turbine at ... · npsarchive 1968 lenzini,m. calibrationofturbinetestrigwith impulseturbineathighpressureratios by martinjosephlenzini