UNCLASSIFIED AD NUMBER LIMITATION CHANGEStion (SFC) turboshaft engine at 2500oF turbine inlet gas temperature would be provided. This volume discusses the fabrication, tests, and redesign

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Approved for public release; distribution isunlimited.

Distribution authorized to U.S. Gov't. agenciesand their contractors; Critical Technology; FEB1970. Other requests shall be referred to U.S.Army Aviation Materiel Laboratories, FortEustis, VA 23604. This document containsexport-controlled technical data.

USAAMRL ltr, 23 Jun 1971

CO (M CO a to 00

9

USAAVLABS TECHNICAL REPORT 69-10B

ADVANCEMENT OF SMALL GAS TURBINE COMPONENT TECHNOLOGY

ADVANCED SMALL AXIAL COMPRESSOR

VOLUME II - TEST AND REDESIGN

lames If. Davis liwni I. Dellen

February 1970

U. S. ARMY AVIATION MATERIEL LABORATORIES FORT EUSTIS, VIRGINIA

CONTRACT TA 44-177.AMC-296(T)

CONTINENTAL AVIATION AND ENGINEERING CORPORATION

DETROIT, MICHIGAN

This document is lubject to ■pccial export control!, and each tranamittal to foreign governmenta or foreign oatlonala may be made only with prior approval of US Army Aviation Materici Laboratories, Fort Euatii, Virginia 23604.

D D C

MM 25 ,970

113

..

·•·

THIS DOCUMENT IS BEST QUALITY AVAILABLE. THE COPY

FURNISHED TO DTIC CONTAINED

A SIGNIFICANT NUMBER OF

PAGES WHICH DO NOT

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ACCiSSiM \t

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The findi) i»»;u8iiin U'i, ment of the

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DISCLAIMERS

in|j8 in this report are not to be construed as an official Depart- Army position unless so designated by other authorized

Wh^n Govfcrnment drawings, specifications, or other data are used for »e other than in connection with a definitely related Government

procurement operation, the United States Government thereby incurs no responsibility nor any obligation whatsoever; and the fact that the Govern- ment may have formulated, furnished, or in any way supplied the said drawings, specifications, or other data is not to be regarded by impli- cation or otherwise as in any manner licensing the holder or any other person or corporation, or conveying any rights or permission, to manu- facture, use, or sell any patented invention that may in any way be related thereto.

DISPOSITION INSTRUCTIONS

Destroy this report when no longer needed. Do not return it to the originator.

DEPARTMENT OF THE ARMY U S ARMY AVIATION MATERIEL LABORATORIES

FORT EUSTIS VIRGINIA 2 3604

The research described herein was conducted by Continental Aviation and Engineering Corporation under U.S. Army Con- tract DA 44-177-AMC-296(T). The work was performed under the technical management of Mr David B. Cale, Propulsion Division, U.S. Army Aviation Materiel Laboratories.

Appropriate technical personnel of this Command have re- viewed this report and concur with the conclusions contained herein.

The findings and reconnendations outlined herein will be considered in the planning of future axial compressor programs .

This is the second volume of a two-volume report. Volume I, USAAVLABS Technical Report 69-10A, covers the analysis and design. This volume covers test and redesign. The aerodynamic redesign portion of this volume is published as a classified addendum under separate cover.

I

■■»■

Task 10162203014413 Contract DA 44-177-AMC-296 (T)

USAAVLABS Technical Report 69-10B February 1970

ADVANCEMENT OF SMALL UAS TURBINE COMPONENT TECHNOLOGY

ADVANCED SMALL AXIAL COMPRESSOR

VOLUME II - TEST AND REDESIGN r

Continental Report 1033

By

James V. Davis Edmund J. Dellert

Prepared By

Continental Aviation and Engineering Corporation Detroit, Michigan

for

U.S. ARMY AVIATION MATERIEL LABORATORIES FORT EUSTIS. VIRGINIA

This document is subject to special export controls, ( and each transmittal to foreign governments or foreign nationals may be made only with prior approval of US Army Aviation Materiel Laboratories, Fort Eustis, Virginia 23604.

SUMMARY

This report presents the preliminary design and analysis of an advanced axial-centrifugal compressor for small gas turbines, and the detail design of the axial stages.

The program objective was to advance and demonstrate efficient high-pressure-ratio axial compressor technology to a level where, when matched analytically with both the advanced centrifugal compressor technology supplied by U. S. Army Aviation Materiel Laboratories (USAAVLABS) and the conventional engine component characteristics, a potential for a 0. 460-pound-per-horsepower-hour specific fuel consumption (SFC) turboshaft engine at 2500oF turbine inlet gas temperature would be provided.

This volume discusses the fabrication, tests, and redesign of the axial compressor. The original axial compressor design (Volume 1) was fabricated and tested. The axial compressor performance was capable of providing a potential for a 0. 484-pound-per-horsepower-hour SFC turboshaft engine at 2500oF turbine inlet gas temperature. However, a low flow problem prevented the compressor from achieving the target efficiency.

The compressor was redesigned, fabricated, and tested. This compressor performance exceeded the contract objective by demonstrating 80 -percent efficiency at 3. 1:1 pressure ratio with a 4. 91 lb/sec airflow, thus providing a potential for a 0.457-pound-per-horsepower-hour specific fuel consumption turboshaft engine at 2500oF turbine inlet gas temperature.

m

FOREWORD

This program is sponsored by the United States Army Aviation Material Laboratories under Contract DA44-1 77-AMC-296{T), Task 1G162203D14413.

This report, prepared by Continental Aviation and Engineering Corporation, presents Phase II and Phase III of a small axial compressor program for the advancement of small gas turbine cc -nponent technology.

The detailed aerodynamic redesign of the compressor is presented in an addendum of Volume II under separate cover.

The details of the compressor concept definition and compressor mechanical design are included in Volume I. The original compressor design is published under separate cover as an addendum of Volume I.

TABLE OF CONTENTS

Page

vm

XIV

SUMMARY iii

FOREWORD v

LIST OF ILLUSTRATIONS

LIST OF TABLES

INTRODUCTION 1

DISCUSSION 2

Fabrication of Original Design 2 Apparatus and Procedures 4 First Pig Test of Axial Compressor 21 Second Rig Test of Axial Compressor 59 Third Rig Test of Axial Compressor 62 Compressor Redesign 62 Mechanical and Structural Compressor Redesign . 68

FABRICATION OF REDESIGNED COMPRESSOR 83

Rotor Assembly 83 Stator Assembly 83 Inlet Guide Vane Assembly 87

FINAL RIG TEST OF REDESIGNED COMPRESSOR .... 88

Aerodynamic Test Results 88 Mechanical Test Results 98

CONCLUSIONS 104

RECOMMENDATIONS 105

APPENDIXES

I. Description of Traverse Data Computer Output . 107 11. Computer Output of Traverse Data for Compressor -

Original Design I 18 III. Computer Output of Traverse Data for Compressor

Redesign 138

DISTRIBUTION 158

'."

Vll

i


Figure

1

2

3

4

5

6

7

8

9

10

11

12

13

14

15

16

17

18

Page

Advanced Axial Compressor Test Riu Design Layout 2

Rotor Assembly - Electron Beam Weldment .... 3

First-Stage Stator Assembly 5

Second-Stage Stator Assembly 5

Compressor Test Cell '400 No. 1 Layout 7

Compressor Installation - 1400 No. 1 Test Cell . . 9

Compressor Test Cell 1400 No. 2 Layout 11

Compressor Installation - 1400 No. 2 Test Cell . . 13

Inlet Duct Probe Calibration !£,

Exit Duct Probe Calibration 16

Survey Probe Calibration 20

Survey Probe Calibration - Fully Shielded 20

Axial Compressor Map - First Rig Test 22

Axial Compressor First Rig Test - Transition Duct Exit Temperature and Pressure Ratio Data 24

Axial Compressor First Rig Test- Transition Duct Exit Mach Number and Efficiency Data 25

Axial Compressor First Rig Test - Static Pressure Distribution Along Compressor Tip 26

Axial Compressor First Rig Test - Transition Duct Static Pressure Distribution 27

Axial Compressor Rotor One - Deviation and Incidence Angle Along Blade Radial Height 28

vui


Figure Page

19 Axial Compressor Rotor One - Exit and Inh t Relative Mach Number Along Blade Radial Height 29

20 Axial Compressor Rotor One - Exit Mach Number and Diffusion Factor Along Blade Radial Height. . . 30

21 Axial Compressor Rotor One - Adiabatic Efficiency and Pressure Loss Coefficient Along Blade Radial Height 31

22 Axial Compressor Rotor One - Temperature and Pressure Ratio Along Blade Radial Height 32

23 Axial Compressor Stator One - Incidence and Abso- lute Inlet Air Angle Along Blade Radial Height ... 33

24 Axial Compressor Stator One - Inlet and Exit Mach Number Along Blade Radial Height 34

■

25 Axial Compressor Stator One - Pressure Loss and Diffusion Factor Along Blade Radial Height .... 35

26 Axial Compressor Stator One - Adiabatic Efficiency and Pressure Loss Coefficient Along Blade Radial Height 3o

27 Axial Compressor Stator One - Static Temperature and Pressure Ratio Along Blade Radial Height ... 37

28 Axial Compressor Rotor Two - Deviation and Inci- dence Angle Along Blade Radial Height 38

29 Axial Compressor Rotor Two - Inlet and Exit Mach Number Along Blade Radial Height 39

30 Axial Compressor Rotor Two - Exit Axial Mach Number and Diffusion Factor Along Blade Radial Height 40

»•

LX


Figure Page

*' Axial Compressor Rotor Two - Adiabatic Efficiency and Pressure Loss Coefficient Along Blade Radial Height 41

32 Axial Compressor Rotor Two - Temperature and Pressure Ratio Along Blade Radial Height 42

33 Axial Compressor Stator Two - Incidence and Air Inlet Angle Along Blade Radial Height 43

34 ALxial Compressor Stator Two - Exit and Inlet Mach Number Along Blade Radial Height 44

35 Axial Compressor Stator Two - Pressure Loss and Diffusion Factor Along Blade Radial Height 45

36 Axial Compressor Stator Two - Adiabatic Efficiency and Pressure Loss Coefficient Along Blade Radial Height 46

37 Axial Compressor Stator Two - Temperature and Pressure Ratio Along Blade Radial Height 47

38 First-and Second-Stage Axial Compressor Head Co- efficient Versus Flow Coefficient 49

39 Erosion of Feltmetal Rotor Shroud Material in First Stage of USAAVLABS Two-Stage 3:1 Compressor Test Rig 50

40 Typical Section, USAAVLABS First-Stage Rotor Passage Area 53

41 USAAVLABS First-Stage Rotor Passage Area Dis- tribution 55

42 USAAVLABS First-Stage Rotor - Comparison of Hub Contours 58

43 Advanced Two-Stage Axial Compressor Rig Test - Hub Relief 60

x

•


Figure Page

44 First-Stage Stator Assembly Showing Deep Groove on Labyrinth Seal Rub Ring 61

45 Advanced Two-Stage Axial Compressor Rig Test - Blade Twist 63

46 Two-Stage Axial Compressor - Original Design. . . 64

47 The Effect of Solidity on Shock Location for Super- sonic Cascades 66

40 Comparison of USAAVLABS First-Stage Flow Paths 69

49 Phase III Compressor - First-Stage Rotor Disc Radial and Centrifugal Stress 72

50 Phase III Compressor - First-Stage Rotor Blades Gas Force Intensity 73

51 Phase III Compressor - First-Stage Rotor Blade Centrifugal Stress 71

52 Phase III Compressor - First-Stage Rotor Blade Cen- trifugal Untwist Stress 74

53 Phase III Compressor - First-Stage Rotor Blade Gas Bending Stress 75

54 Phase III Compressor - First-Stage Rotor Blade Modi- fied Goodman Diagram 76

55 USAAVLABS Phase III Compressor- First-Stage Rotor Blade Interference Diagram 77

56 USAAVLABS Phase III Compressor - Second Stage Rotor Disc Radial and Tangential Stress 73

57 USAAVLABS Phase III Compressor- Second^Stage Rotor Blade Gas Force Intensity 78

XI


Figure Page

58 USAAVLABS Phase III Compressor- Second^Stage Rotor Blade Centrifugal Stress 7 9

59 USAAVLABS Phase III Compressor - Second-Stage Rotor Blade Centrifugal Untwist Stress 7 9

60 USAAVLABS Phase III Compressor - Second-Stage Rotor Blade Gas Bending Stress 80

61 USAAVLABS Phase III Compressor - Second-Stage Rotor Blade Modified Goodman Diagram 81

62 USAA VI ABS Phase III Compressor - Second-Stage Rotor Blade Interference Diagram 82

63 Redesigned Welded Rotor Assembly 84

64 Redesigned First-Stage Stator and Compressor Housings 85

65 Redesigned First-Stage Stator Showing EDM Cut in Vane Tip 86

66 Redesigned Variable Inlet Guide Vane Assembly . . 87

67 The Effect of Overall AxieJ-Centrifugal Compressor Pressure Ratio and Axial Compressor Performance on Brake Specific Fuel Consumption 89

68 USAAVLABS Centrifugal Compressor Technology Performance Along Given Operating Line 90

69 USAAVLABS Two-Stage Transonic Axial Compressor Performance 92

70 Two-Stage Axial Compressor Transition Duct Exit Performance at Compressor Design Speed - Effi- ciency and Pressure Ratio 93

xii


.

Figure Page

71 Two-Stage Axial Compressor Transition Duct Exit Performance at Compressor Design Speed - Mach Number and Absolute Velocity 94

72 Two-Stage Axial Compressor Transition Duct Exit Triangle for a Centrifugal Compressor Inducer Inlet at Design RPM 95

73 Parametric Results of a Family of Centrifugal Com- pressors Operating in Single-Spool Configuration Behind USAAVLABS Two-Stage Axial Compressor .... 96

74 Advanced Two-Stage Axial Compressor Transition Duct Total Pressure Loss 98

75 Static Pressure Distribution Along Compressor Tip 99

76 Two-Stage Compressor Transition Duct Exit Static Pressure Distribution 100

77 Transonic Two-St age Axial Compressor - The Effect of Variable Inlet Guide Vane Setting Angle on 100- Percent Design Soe^d Compressor Performance . . 101

78 Transonic Two-Stage Axial Compressor - The Effect of Variable Inlet Guide Vane Setting Angle on 100- Percent Design Speed Compressor Performance . . 1 ü2

79 Advanced Two-Stage Compressor Comparison of Nominal and +20 Degree Inlet Guide Vane Stagger Angle on Compressor Performance 103

Xlll

LIST OF TABLES

Table Page

I First-Stage Rotor Modification Possibilities. ... 56

II Analysis of High-Pressure-Ratio Axial Rotors . . 67

III Summary of Phase III Component Redesign .... 70

IV Computer Output of Traverse Data for Compressor- Original Design - Test Number 19 118

V Computer Output of Traverse Data for Compressor - Original Design - Test Number 20 123

VI Computer Output of Traverse Data for Compressor- Original Design - Test Number 21 128

VII Computer Output of Traverse Data for Compressor - O ' <inal Design Test Number 22 133

VIII Computer Output of Traverse Data for Compressor - Redesign - Test Number 15 138

IX Computer Output of Traverse Data for Compressor- Redesign - Test Number 16 143

X Computer Output of Traverse Data for Compressor - Redesign - Test Number 17 148

XI Computer Output of Traverse Data for Compressor- Redesign - Test Number 18 153

xiv

INTRODUCTION

This report presents the work accomplished in Phases II and III of Contract DA44-1 77-AMC-296(T) for the United States Army Aviation Materiel Laboratories, Fort Eustis, Virginia.

The project objectives are to advance and demonstrate efficient high-pressure-ratio axial compressor technology to the level where, when matched analytically with the advanced centrifugal compressor technology supplied by USAAVLABS and the conventional engine component characteristics, a potential for a 0. 460-pound-per-horsepower-hour / SFC turboshaft engine at 2500 F turbine inlet gas temperature will be / provided.

The Phase I objectives are presented in Volume I.

The Phase II objectives were to fabricate and test the axial compressor to determine basic performance and to provide aerodynamic data for any necessary modification of the blade rows. An additional test of a modified compressor was to be conducted.

The Phase III objective was to redesign the axial compressor using the Phase II aerodynamic data as the basis for aerodynamic direction. The redesigned compressor was to be fabricated and tested to determine basic performance.

DISCUSSION

FABRICATION OF ORIGINAL DESIGN

General Fabrication Techniques

The Continental-designed compressor rig, Figure 1. is primarily an aerodynamic research vehicle, with the structural design emphasizing mechanical integrity, ease of assembly and instrumentation, and reason- able cost, wherever possible. Consequently, the majority of the stationary hardware fabrication was straightforward utilizing weldments and other common techniques. However, the manulacturing techniques used in the rotor and stator assemblies were somewhat more complex and are discussed in the following paragraphs.

Variable Inlet Guide Vane a

Long Transition Duct

Figure 1 Advanced Axial Compressor Test Rig Design Layout.

Rotor Assembly

The integrally bladed rotors and the rear shaft, each machined from a solid AMS 5616 forging, were joined into a unitized assembly by electron-beam welding. The weld joints on the rotor assembly are shown in Figure 2. Two rotor assemblies were fabricated in this manner.

Electron-Beam Weld Joints

Figure 2. Rotor Assembly - Electron Beam Weldment.

After welding of the first rotor assembly, radiographic (X-Ray) inspection revealed extensive porosity throughout the weld. The short span between the first-and second-stage rotors prevented the use of X-Ray techniques and porosity depth repair procedures. Therefore, prior to welding the second rotor assembly, an extensive weld qualification technique was developed and is summarized as follows:

1. Establish basic weld parameters by using a sample bead weld on plate.

2. Evaluate joint configuration using flat plates by:

a. X-Ray examination

b. Macroscopic examination of the weld.

3. Evaluate joint configuration using circular components in the same manner as above.

L

4. Weld rotor assembly

a. Examine by fluorescent penetrant

b. Examine by X-Ray

During the course of establishing the qualification technique, it was determined thatthe operator's skill in focusing the electron beam and subtle changes in the operation of the beam filament were major factors in producing a sound weld-

Use of this technique resulted in an acceptable weld for the second rotor assembly. Maximum runout after welding was 0.003 inch; the total Indicator reeding after stress relief was 0.007 inch. The excessive runout condition after stress relief was corrected by heat straight- ening.

Final machining of the rotor assembly was accomplished following acceptance of the welds. Excess stock was provided in critical areas, such as bearing journals and labyrinth seals, so that any misalignment resulting from the welding operation could be readily corrected.

Stator Assembly

The individual contours of both the first- and second-stage stator vanes were machined by conventional methods from AIS1 410 stainless steel.

The assembly of both stators was similar. The electron discharge machining (EDM) of the vane slots, in the inner and outer shrouds, v/as aocomj, lished by using the vane section as an electrode. The vanes were then placed in the slots and brazed in place (Figures 3 and 4).

Variations in vane thickness made it necessary to hand rework and fit each vane in a particular slot. In addition, holding fixtures were needed to maintain vane positioning during the brazing process.

APPARATUS AND PROCEDURES

Test Cell Installation

The compressor, with the integrally mounted speed increaser was adapted as a package to the test cell and driving facilities. Input torque for this package is restricted by the integral speed increaser

,

Figure 3. First-Stage Stator Assembly.

'.

Figure 4. Second-Stage Stator Assembly.

.

I

to approximately 360 horsepower at the compressor design speed of 59,600 rpm. Consequently, in order to decrease the horsepower require- ments, the compressor had to be operated at reduced inlet pressures up to 24 inches of mercury absolute at design point.

Two test cells can accommodate the compressor: the 1400 No. 1 or the 1400 No. 2. These test cells are in adjacent areas arranged back- to-back such that two reversible, 700-horsepower, electrical dynamome- ters can '-je coupled in series for 1400 horsepower to either cell or operated independently for 700 horsepower in eacii cell. Each cell provides lubrication services, inlet air temperature regulation from 300oF to -650F, inlet air pressure regulation from 40 inches of mercury absolute to highly depressed conditions, and exhaust services ranging from atmos- pheric to hi^h vacuum conditions.

The original compressor design was installed and tested in 1400 No. 1 test cell (Figures 5 and 6). This cell provides airflow measurement by means of an ASME nozzle station with the cell inlet plenum. Drive provision was from the intermediate shaft of a two-stage, 2500- horsepower, 42,000-rpm, 21:1 ratio gearbox. The first-stage ratio of this gearbox is 6.93:1 (1386 rpm) so that operation ranging to ic JS than half speed adequately acconunodate the input to the compressor/integral speed increaser package.

The redesigned compressor was Installed and tested in 1400 No. 2 test cell (Figures 7 and 8). Installation of the No. 2 test cell required airflow measurement by means of a previously calibrated compressor inlet station because no 'M measurement station was provided. Drive provision was from an ii nediate shaft of a two-stage, 1400-horsepower, 42,000-rpm, Z1 atio gearbox. The first-stage ratio of this gearbox is 3. 263:1 (6i .:n); due to its design of twin load-charing intermediate shafts, it > limited to 700 horsepower at 6500 rpm. Opera- tion ranging to 80 percent speed adequately provided the input to the compressor/integral speed increaser package.

Adjacent to each test cell is its respective control room, which provides instrumentation read-out equipment for both aerodynamic and mechanical units, control of service equipment, and control of the drive systems. Inlet air temperature and pressure, drive speeds, discharge air pressure, and required service equipment such as flowpath traversing instrumertation, oil pressure, and oil temperature are regulated remotely from the control rooms to provide the desired ranges for test operation.

>

^ -a

\

Figure 5. Compressor Test Cell 1400 No. 1 Layout.

P-

kmt.

Figure 6. Compressor Installation - 1400 No. 1 Test Cell.

Test Procedure

Both the original design and the redesigned compressors were run at low speed to check mechanical integrity and then gradually ac- celerated from 50 to 100 percent of design speed in increments of 10 percent with the exit throttle valve open. The abradable shrouds were inspected at intervals of 50,70, 90, and 100 percent of desigr speed. The compressor map was defined by setting a specific speed and gradually closing the throttle valve to determine the speed line. Approximate- ly five data points, at different flow rates, were obtained for each speed line. Surge was determined by observing fluctuations in the exit manometers and the inlet flow manometers. Traverse data were obtained at selected points after the compressor map was defined.

Aerodynamic Instrumentation

In general, the same aerodynamic instrumentation was used for both the original design and the redesigned compressors, except for the methods of airflow measurement and the presence of traverse probes in the original design compressor. The different methods of flow measurement, as described below, were necessitated by the lack of a nozzle station in the 1400 No. 2 test cell.

Specific instrumentation installations for both compressor configurations are discussed in the following paragraphs.

Compressor Inlet. The compressor inlet conditions were measured in the tapered inlet transition duct at a plane 4.0 inches forward of tie compressor inlet housing. The rake instrumentation was equally spaced at four points across the duct inner radius of 3.30 inches. The temperature was measured with four bare wire iron-cons tan tan temperature probes. The inlet total pressure was measured by four elements of 0.062-inch diameter with a 30-degree internal taper. One inlet duct wall static was provided at this plane for Mach number correction of the temperature rake.

Compressor Exit. The compressor exit incorporated one three-element total pressure rake, positioned 0.075 inch behind the trailing edge of the second-stage stator vane and centered between vanes. The three elements were equally spaced across the 0. 545 inch passage and fabricated from 0. 040-inch tubing. Two outside diameter wall static taps were installed 90 degrees apart, approximately 0. 020 inch behind the trailing edge of the stator and centered between vanes.

10

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-JBDJJOG&I.

yout.

ß

Figure 8. Compressor Installation - 1400 No. 2 Test Cell.

13

Transition Duct. Instrumentation at the exit of the transition duct consisted of three 3-element total pressure probes and three 3-element total'temperature probes. The temperature probes were bare wire iron-cons tan tan thermocouples, and the pressure probes were fabricated from 0.040-inch tubing. Elements were equally spaced across the 0.889-inch passage width. The rakes were circumferentially spaced 120 degrees apart. The sensing plane was located at approximately 0. 190 inch behind the mating surface of the rear bearing housing and the transition duct. Two outside diameter and one inside diameter wall static pressure taps were installed at this plane.

Wall static instrumentation through the transition duct consisted of three static pressure taps - two at the outside diameter, 90 degrees apart, and one at the blade diameter - at each of three locations evenly spaced axially between the duct inlet and the duct exit.

Interstage Stat Pressure Instrumentation

Instrumentation through this section consisted of outside diameter wall static pressure taps installed in pairs, 90 degrees apart, at the fol lowing locations:

1. At the center of the variable inlet guide vane housing flange, one between the vanes and another in line with a vane.

Z. At the first-stage rotor inlet, opposite the point of intersection of the rotor leading edge with the hub.

3. At the inlet to the first-stage stator, slightly ahead of the leading edge.

4. At the exit of the first-stage stator, slightly behind the trailing edge.

5. In line with the leading edge of the second-stage stator.

Inlet Airflow. With the redesigned compressor, the airflow was measured, using an airflow coefficient established by calibration of the inlet assembly against a 4.00- and a 5.00-inch ASME nozzle, over the flow range of 1.65 to 5.00 pounds per second. An average coefficient of 0. 1553 was established for use with a conventional airflow calibration curve. The inlet total pressure was measured by four wall static pressure taps manifolded together and located in the inlet plenum forward of the inlet transition duct. To establish the pressure drop, two wall static

14

pressure taps, locateJ in the inlet housing, were manifolded and teed to the inlet plenum pressure. On the original design compressor, airflow was measured with a 5-inch ASME nozzle mounted in the inlet plenum upstream of the compressor.

Traverse Probes. Actuated total pressure cobra probes and bare wire iron-constantan total temperature probes were installed axially behind each rotor on the original design compressor. The leading edge of one stator /ane was removed to enable the probes to traverse radially.

No traverse probes were utilized on the redesigned compressor.

Mechanical Instrumentation

The compressor assembly incorporated the usual rotating component rig instrumentation consisting of thermocouples on each of the bearing outer races, oil-in and -out temperatures and pressures, and vertical and horizontal accelerometers mounted on the front and rear of the compressor housing.

In addition, strain gages were mounted on two beams of the front bearing cage, at 90-degree spacing, to detect shaft oscillations trans- mitted through the cantilevered cage, thereby giving an indication of shaft motion relative to the housing.

The signals from the strain gages were fad into a dual beam os- cilloscope, where they could be displayed in sine wave or orbit form.

Data Reduction

Overall Data. The average inlet total temperature was established by arithmetically averaging the four thermocouple temperature readings at the inlet station. The average inlet total pressure was established by manifolding the four manometers connected to the four inlet pressure probes. The exit conditions were measured in a similar manner to the inlet conditions.

Since the Mach number at the inlet plane and at the exit plane is high enough to cause velocity effects on the total temperature probe, the inlet probes were calibrated against a kno vn source as a function of Mach number. The calibrations for each test are shown in Figures 9 and 10. The method of obtaining the actual temperature is summarized in the following:

15

,

12

Z

^ 2 8

8 r

5 I

IK

Calibration T, - 551 ^

X ^^

--'

0.1 0.2 0.3 0.4 0.5 0.6 MACH NUMBER

0.7 0.8

Figure 9. Inlet Duct Probe Calibration.

0«ok?Nc. 1, Calibration Tt'637 OR

O-RaU No. 2, Calibration Tt « 637 Oft

A-Kak« No. 3, Calibration Tt • 636 o«

I K 12

f I I o.i 0.2 0.3 0.4 0.S

MACH UUMKR

0.6 0.7 0.8 0.9

Figure 10. Exit Duct Probe Calibration.

16

1. Obtain static pressure and total pressure from manometer readings at the total temperature measuring station.

2. Calculate flow Mach number, using the following:

M =

where:

- 1

- 1

U.Pf (1)

M = Mach number

a = Specific heat ratio

P^ = Total pressure, psia

Pg = Static pressure, psia

Determine the temperature correction (AT/TJ) from the calibration curves (Figures 9 and 1C) and calculate the actual or true temperature as follows:

^T T true = Tt (1 + (2)

T true ■• True temperature, 0R

^T

T.

- Measured temperature by the probe, 0R

Percentage difference between actual temperature and probe temperature.

Overall Efficiency. The overall efficiency was computed using the standard efficiency formula shown below:

a - 1

H - i

TR - 1 (3)

17

m

wh ere:

r\ = Efficiency

PR = Exit total pressure divided by inlet total pressure

TR = Exit total temperature (corrected for Mach number effects) divided by inlet total temperature.

a = Ratio of specific heats determined by an integrated averaging process

Traverse Data Measurements. The probes were traversed radially inward, and the data were recorded at the following percentages of blade height locations: 5, ]0, 30, 50, 70, 90, and 95. The total pressure probe was adjusted to the proper flow angle by balancing the static pressures. The temperature probe data were recorded at the flow angles determined by the total pressure probe.

Both the total temperature and the total pressure probes were calibrated for Mach number effects. The method of obtaining the actual traverse temperature is equivalent to the method used for obtaining the actual overall temperature. The method of calculating the actual pressure is summarized below:

1. At each radial station, obtain the absolute Mach number from computer results. A definition of the computer output parameters is included in Appendix I.

2. Obtain pressure ratio (Pt probe/Pt true) from the probe calibration curve and calculate the actual pressure as defined below:

P. actual = P^ measured x Pt probe/Pt true

(4)

where:

Pt actual

Pt measured

Pt probe

= Actual pressure measured in calibration tunnel, psia

= Measured total pressure, psia

= Instrument pressure measurement in calibration tunnel, psia

18

Pf- probe/Pt true = Calibration total pressure ratio

The calibration curves for all of the traverse probes are presented in Figures. 11 and 12.

Traverse Data Reduction. At the inlet to the first-stage rotor, the test static pressure, total temperature, and total pressure were assumed constant from hub to tip. Flow conditions at this station were obtained by the continuity relationship.

At the exit of the rotors, traverse data total temperature and total pressure were used to establish flow conditions. The slope of static pressure with radius was determined by assuming radial equilibrium. The level of static pressure was calculated by assuming continuity. The tangential velocity was established from the total temperature rise and the Euler turbomachinery equation.

The flow conditions behind the first-stage stator were established by assuming design losses ô obtain the total pressure distribution. No change in total temperature was assumed from the exit of the first- stage rotor to the exit of the first-stage stator along a streamline. The flow direction w?.s considered to be axial, and radial equilibrium was used to calculate the static pressure slope at the stator exit. Continuity determined the level of static pressure.

The exit conditions from the second-stage stator were calculated in the same manner as those from the first-stage stator «xcept that the total pressure rake installed at the second stator exit was used to calculate the stator loss.

1Q

0.90

O Ficit-Stog« Roter Survoy Prob« D Socend-Stagt Rotor Survey Prob«

0.1 0.2 0.3 0.4 0.5 0.« MACH NUMBER

0.7 0.8 0.9 l.o

Figure 11. Survey Probe Calibration.

F 24

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2 I i

J r

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fr

/ ^

^

y y r

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r

i

^

/

/

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Survey Probe

tor Survey Prd

Calibration

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J /

r

m r

0.1 0.3 0.3 0.4 0.5 0.6

MACH NUMK«

0.8 0.9

Figure 12. Survey Probe Calibration - Fully Shielded.

20

FIRST RIG TEST OF AXIAL COMPRESSOR

The axial compressor assembled with the long transition duct and without the variable inlet guide vanes was tested to define basic performance and to provide data for any necessary aerodynamic modifications. The design pressure ratio objective was reached. However, because of a low-flow condition, the compressor did not obtain the design flow and efficiency. A comparison of the demonstrated performance with the design objectives at 59, 600 rpm is shown below:

Overall pressure ratio Efficiency, percent Corrected flow, lb/sec.

Design

3.0:1 82.3 5.00

Demonstrated

3.0:1 72.5 4.32

Mechanically, the rig functioned satisfactorily, except for the air erosion of the abradable feltmetal (metal fiber) shrouds.

Aerodynamic Test Results

Overall Performance Data. Sufficient test data. Figure 13, were obtained to define an overall compressor map. These data were measured from the inlet of the compressor (4.5 inches upstream of the inlet struts) to the transition duct exit. The test data indicated performance potential for obtaining the design objectives. The overall compressor characteristics obtained were more than adequate. An excellent flow range was demonstrated by the compressor at all speeds. The design speed stall margin attained at 3.0:1 pressure ratio was 10.5 percent. The definition of stall margin is shown below:

Stall Margin = (pR/WaVÖ^J s - fpR/Wa"V^| OP

|pR/waVo78j 0p x 100% (5)

where:

PR

vvaVo7s

S

OP

= Compressor total pressure ratio

- Inlet corrected airflow

= Surge

= Operating point

21

Z.i 3.0

CORRECTED AIRFLOW,

i. 0

Figure 13. Axial Compressor Map - First Rig Test.

22

Because of premature choking condition, the compressor did not obtain the design flow and efficiency. Some ;mprovement in flow was noticed with severe erosion of the first-stage rotor shroud. Figure 13.

The transition duct overall data, Figuies 14 and 15, show a flow shift from tue hub to the tip. This is indicated by the increase in absolute Mach number from hub to tip. A decaying efficiency gradient from nub to tip is noticed.

Static Pressure Data. The statir pressure data, Figure 16, was normalized by the inlet total pressure to account for differences in the inlet total pressure. A complete range of data from choke to surge is also shown in Figure 16, for 100 percent of design speed.

The transition duct static pressure distribution, Figure 17, shows good agreement with the analytical prediction. The overall level of static pressure was high with respect to the predicted level because of the reduced flow rate.

Traverse Data. Traverse data were obtained for the compressor at the rotor exits in order to provide a basis for either a modification or a redesign. The interstage performance was recorded and reduced for the following test points:

Corrected Overall Corrected Speed - Pressure

Test Number Airflow-Lb/Sec Percent Ratio

19 4.514 105 2.898 20 4.315 100 3.095 21 4.359 100 3.032 22 4.372 100 2.903

Complete flow conditions for all blade rows obtained from the traverse data of the above test runs are presented in Appendix II. This appendix represents the computer program output data. The assumptions used to calculate the flow conditions are included in the discussion of traverse data reduction. A description of the computer program output symbols is included in Appendix I. Curves of the data from Continental Test 822B/run number 21 (plotted from Appendix II), are shown in Figures 18 through 37 and are v_ninpared to the design objectives. This particular test point (run 21) is cf particular interest because it was conducted near the design pressure ratio, 3.0:1, and the design speed, 59,600 rprn.

^3

40 60 80

RADIAL HEIGHT - PERCENT

Figure 14. Axial Compressor First Rig Test - Transition Duct Exit Temperature and Pressure Ratio Data.

24

!

RADIAL HEIGHT-P£RaNI

Figure 15, Axial Compr essor First Rig Test Transi

Number and Efficiency Data. tion Duct Exit Mach

25

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Figure 18, Axial Compressor Rotor One - Deviation and Incidence Angle Along Blade Radial Height.

28

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RADIAL HEIGHT -PERCENT 100

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Figure 20. Axial Compressor Rotor One - Exit Mach Number and Diffusion Factor Along Blade Radial Height.

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«A0IALH£IGMT - Pf«CCNT

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Figure 21. Axial Compreisor Rotor One - Adiabatic Efficiency and Pressure Loss Coefficient Along Blade Radial Height.

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Figure 22. Axial Compressor .Rotor One - Temperature and Pressure Ratio Along Blade Radial Height.

32

1/1 LU UJ

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Figure 23.

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Axial Compressor Stator One - Incidence and Absolute Inlet Air Angle Along Blade Radial Height.

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Figure 24. Axial Compressor Stator One - Inlet and Exit Mach Number Along Blade Radial Height,

34

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Figure 25. Axial Compressor Stator One - Pressure Loss and Diffusion factor Along Blade Radial Height.

35

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HUB RADIAL HEIGHT-PERCENT TIP

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Figure 26. Axial Compressor Stator One - Adiabatic Efficiency and Pressure Loss Coefficient Along Blade Radial Height.

36

f

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RADIAL HEIGHT-PERCENT

J 100

TIP

Figure 27. Axial Compressor Stator One - Static T^. Pressure Rat.o Aiong Blade Rad^He.gh"^^" ^

i7

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Test

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Figure 28. Axial Compressor Rotor Two - Deviation and Incidence Angle Along Blade Radial Height.

38

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HUB

20 40

RADIAL HEIGH1

60 80

PERCENT

100

TIP

Figure 29. Axial Compressor Rotor Two - Inlet and Exit Mach Number A' .ng Blade Radial Height.

39

m

Design

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TOR

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X UJ

^\ ^

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TIP

Figure 30. Axial Compressor Rotor Two - Exit Axial Mach Number and Diffusion Factor Along Blade Radial Height.

4U

Design

0.5 Z

y

o

8

•Ö 0 I

^r

90

M u It Hl

u

< 5

80

70

(sn

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HUB RADIAL HEIGHT -PERCENT TIP

Figure 31. Axial Compressor Rotor Two - Adiabatic Efficiency and Pressure Loss Coefficient Along Blade Radial Height.

il

^

1.75

LU Of D in UJ I

——— Design

Test

1.65

1.55

1.3

<

i 1.2 ^

I.I

^^ " -^

/

^m mm _, —^ i -

^^ '

0

HUB

20 40 60


80 100

TIP

Figure 32. Axial Compressor Rotor Two - Temperature and Pressure Ratio Along Blade Radial Height.

42

;

Design Incidence Angle is at Zero

40 60

»AOIAL WIGHT - PHCENT .

Figure 33. Axial Compressor Stator Two - Incidence and Air Inlet Angle Along Blade Radial Heignt.

43 .

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as

z X U

0.8

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0.2

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,

0 20 40 60


80 100

TIP

Figure 34. Axial Compressor Stator Two - Exit and Inlet Mach Number Along Blade Radial Height.

44

0.8

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§ 0.4

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on

iSt

• Design

• Test

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0 20 40 60


80 100

TIP

A

Figure 35. Axial Compressor Stator Two - Pressure Loss and Diffusion Factor Along Blade Radial Height.

4 5

•Design

0.3

y tt 0.2 LLI 5 u in

o § 0.1 1/1 on UJ S a.

90

> z uj 80 o

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80 100

TIP

Figure 36. Axial Compressor Stator Two - Adiabatic Efficiency and Pressure Loss Coefficient Along Blade Radial Height.

46

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1.7

-—— Design

O

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ÜJ B

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20 40 60 80

RADIAL HEIGHT-PERCENT

100

TIP

Figure 37. Axial Compressor Stator Two - Temperature and Pressure Ratio Along Blade Radial Height.

i 47

A head-flow analysis,using results of ^he traverse data, showed that both stages are operating at lower than design flow coefficient at near design point pressure ratio (Figure 38). A definition of the head flow parameters is shown below.

Definition of head and flow coefficient:

Head Coefficient =\h - GJ Cp Ttl ! (PR [

8 - l

] U^

(6)

Flow Coefficient t (V) u

where:

G

J

S Ttl

PR

h

U

Gravity Constant, it/sec

Mechanical heat equivalent, ft - lb/Btu

Specific heat at constant pressure, Btu/lb 0R

Stage inlet total temperature, 0R

Stage total pressure ratio

Ratio of specific heats

Mean radius wheel speed, ft/sec

Mean axial velocity, ft/sec

Mechanical Test Results. In completing the first test series, the advanced two-stage small axial compressor demonstrated excellent mechanical integrity v/ith only minor problems developing.

The primary problem, realized during the early phases of testing, involved the abradable shrouds used in providing minimum tip clearances on the rotors. Initially, both the first- and second-stage shrouds utilized feltmetal as the abradable material. During testing^ high-speed air erosion was experienced on the first-stage shroud, as shown in Figure 39.

48

0.7

0.6

, 0.5

I o 0.4

% S 0.3

v X 0.2

0. 1

"$ X

First Stage

0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 Flow Coefficient - p

0.7

0.6

0.5

0.4

0.3

0.2

0. 1

0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 Flow Coefficient - tf

Figure 38. First-and Second-Stage Axial Compressor Head Coefficient Versus Flow Coefficient.

-5- i

c v

I) 0 U

0»

X

X Design i r i

Point

Configuration

A Modified Compressor C onfigu ration

/ k IM /\ X

A 0

I s

Second Stage

49

Figure 39. Erosion of Feltmetal Rotor Shroud Material in First Stage of USAAVLABS Two Stage 3:1 Compressor Test Rig.

50

After considering the problems experienced with the feltmetal, a search of o'her abradable material candidates was made, resulting in thp decision to use flame-sprayed aluminum on the first-stage shroud. This shroud has proven to be very effective in use.

Aerodynamic Data Analysis

Comparison of Static Pressure Data and Traverse Data. The static pressure distribution shows an indication of choke in the vicinity of the inlet to the second-stage rotor. A choked condition exists when changes in downstream pressure do not affect the flow conditions upstream of the choked area; in this case, the flow is unaffected upstream of the second rotor inlet. Since all of the static pressure taps were located between blade rows (that is, there are no static pressure taps di- rectly over the rotor blades or stator vanes) an exact location of the actual choked position was undeterminable from these data.

In general, the traverse data indicate a severe loss at the first-stage rotor tip and a tip-to-hub flow shift as the flow passes through this blade row. The first-stage stator, as a consequence of this flow shift, is stalled at the tip and choked at the hub. Both the second-stage rotor and the second-stage stator appear to be in stall, as indicated by high blade and vane incidence values. This observation, which is con- trary to the results of the static pressure data, is discussed in detail below.

The analysis of the aerodynamic data was directed to determining the cause of the low-flow condition and to finding a solution. The anomaly in the data, a choked second-stage rotor based on static pressure measurements on the one hand, and a stalled second-stage rotor based on traverse data m ;asurements on the other hand, was most diffi- cult to decipher.

The conclusion reached from all the data was that the flow in the first rot«, r was stalled (or separated) at the tip, which caused a flow shift from the tip to the hub. This flow shift caused the first stator to be choked at the hub and stalled at the tip. The second-stage rotor, as a result of operating with a low inlet total pressure caused by the stator hub choke losses and the stator tip incidence losses, is choked and thus limits the overall compressor flow rate.

The apparent anomaly in the traverse data showing the second-stage rotor in stall can be explained by the assumptions used to xc- duce the data and to obtain the velocity triangles.

51

The first rotor exit traverses of total pressure and total temperature completely define the static conditijns from hub to tip between the first-stage rotor and the first-stage stator with addition of the folloving as:i imptions:

1. No .i>virl is assumed at the rotor inlet.

?,. The Euler turbomachinery equation defines the tangential swirl t ehind the rotor.

3. Design values of flow blockage are used.

4. Continuity was assumed.

5. Radial equilibrium was assumed.

These data and assumptions are considered adequate enough to describe the flow conditions behind the first-stage rotor. However, jince traverses were impracticable to obtain behind the first-stage stator, loss assumptif is had to be made in order to define the total pressure distribution in this location. In this case, the first-stage stator design losses were assumed. At the time thatthe data were being reduced, no blade row loss analysis was available at Continental to pre- dict the stator losses and accompanying flow shifts that result from the severe tip stall-hub choke phenomenon. Therefore, the traverse data assumptions used in this particular blade row were probably inadequate. The losses should be much higher than shown, and in turn, the total pressure behind the stator is probably much lower than shown. Thus, if the stator losses were much higher than the values assumed,the second- stage rotor would show indications of choking.

Examination of Various Rotor Modifications. In order to pin- point the caus^ of the first rotor tip stall condition, and in turn unchoke the second-stage rotor, an analytical study was conducted. This study consisted of determining why the first-stage rotor was stalled and of ex- amining various modifications to increase the flow and provide aerodynamic data for the redesign phase. These modifications were assessed for performance increase, practicality, length of time required to complete the modification, and cost.

An analysis of the first-stage rotor flow passage between adjacent blades revealed that the variation of area normal to the relative flow does not follow that of a typical rotor. This area, shown in Figure 40, is defined as being normal to the relative flow and bounded by:

52

TYPICAL SECTION. APPROXIMATELY NOPMAL TO RELATIVE FLOW

TOP VIEW TWO ADJACENT BLADES

SIDE VIEW

Figure -*U. Typical Section, USAAVLABS First-Stage Rotor Passage Area.

53

the suction side, surface of a blade

the pressure side, surface of an adjacent blade

the hub circumferential distance between blades

the tip circumferential distance between blades

The area distribution, shown on Figure 41 j minimizes at approximately 0. 1 inch upstream of the rotor stacking line. Using this minimum area as a basis, one-dimensional flow analyses, assuming both design aerodynamic and test data conditions, were performed. The flow analyses showed that, in both cases, the rotor was choked. In addition, further analyses showed that the inlet area to the flow passage or channel, Figure 41, was large enough to pass design flow. Based on these results, modifications to the first rotor were considered.

The modifications which were evaluated to increase the flow of the first rotor and in turn the overall efficiency are presented in Table I and include:

1. Open tip blade twist (or restagger), 5 degrees

2. Open leading edge twist, 7 degrees

3. Hub relief, 0. 110 inch

4. Extended tip, 0.060 inch

5. Variable inlet guide vanes, 21 degrees

6. Redesign, new stage one rotor and new inlet assembly

The flow and efficiency considerations for each of these candidate modifications are discussed below:

1. Blade Twist

The 5-degree open twist should increase the flow to the design value of 5 pounds per second, but at a first-stage efficiency penalty. Usually, when a rotor is twisted open, the throat is increased, but in turn, with an increase in incidence. For example, the first rotor incidence would increase from 6.6 to 7. 25 degrees at a design speed.

54

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U 5 w

s 8 < w OS

w Ü

<:

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0.56

0.54

0.52

0.50

0.5 -0.4-0.3 -0.2 -0.1 0 0.1 0.2 0.3 0.4 0.5

AXIAL DISTANCE FROM STACKING POINT - INCHES

TIP LEADING EDGE

TIP TRAILING EDGE

Figure 41. USAAVLABS First-Stage Rotor Passage Area Distribution.

55

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56

High incidence values significantly increase the rotor relative shock losses and cause severe flow separation.

2. Leading Edge Twist

The 7-degree leading edge tv/ist should also increase the flow to 5 pounds per second, but the incidence at design speed will be even higher. Since the minimum passage area is near the center (Figure 41), a higher twist angle is required for the leading edge twist than for the blade twist to obtain an equivalent amount of area increase. Thus, the rotor losses will be even higher than those of the blade twist.

3. Hub Relief

The hub relief shown on Figure 42 should increase the flow by the required amount and at the same time reduce the incidence to that of design, 3.0 degrees. This method opens the minimum area without changing the blade shape or stagger. The passage area, after relief, should approximate that of a typical rotor, as shown on Figure 41.

4. Extended Tip

The extended tip requires a tip radius increase of about 0.060 inch in order to pass 5 pounds-per-second flow. However, the incidence at design speed will not significantly change from the test value of 6. 6 degrees even though the flow would be increased; because the axial velocity and tip speed remain nearly constant, the relative air angle and, in turn, the incidence would not be substantially changed. The actual change in incidence is about 0. 5 degree and is not sufficient to substantially decrease the first-rotor losses.

5. Variable Inlet Guide Vanes

Use of the variable inlet guide vanes would not increase the flow to 5 pounds per second as shown in Table I, Since the airflow is lower than predicted, the relative air angles into the first rotor are much higher than the design angles. Therefore, the inlet velocity triangle is much "flatter" than originally in- tended, and a very large inlet guide vane turning angle is required to substantially increase the flow. As shown in Table I,

^7

■

•

ORIGINAL HUB CONTOUR

HUB RELIEF CONTOUR

'A^z/^z/z^m n

Figure 42. USAAVLABS First-Stage Rotor - Comparison of Hub Contours,

the relatively high value of variable inlet guide vane setting angle (21 degrees) increases the flow to only approximately 4.5 pounds per second.

6. Redesign

A redesign of the first rotor, using the first rig test results and the analytical study as a basis, should increase the flow and efficiency of the compressor to design values. The first stage would be redesigned to a slightly higher hub/tip ratio and different hub contour to ensure a satisfactory blade passage area distribution.

On the basis of the above flow and efficiency considerations for each modification, the hub relief and the redesign were recommended as being the most practical on the basis of time, cost, and risk.

58

L

SECOND RIG TEST OF AXIAL COMPRESSOR


The second rig test, with the hub relief first-stage rotor, was conducted and data were obtained at 60, 80, 90, and 100 percent of design speed. No significant change in compressor performance was observed compared to the first rig test. The flow at 100-percent design speed increased from 4.359 pounds per second to 4.428 pounds per second, an increase of about 1.5 percent (see Figure 43). No traverse data were obtained because of the small change in flow. Since the flow did not change, a check on the analysis of rotor flow passage area (candidate modification 3 (see Fabrication) was conducted. This analysis showed that the assumption of using the one-dimensional flow was inadequate. The examination should have used individual stream tube areas as the basis for analysis.

The decision was made to twist the rotor open 5 degrees at the tip (modification number 1 (see Fabrication). This is the only modification that had the possibility of providing a flow increase without a major change in hardware. As mentioned, the 5-degree open twist should increase the flow to the design value, but at a first-stage efficiency penalty. However, the overall compressor efficiency should increase because of the improved second-stage aerodynamic match. In addition, second- stage data should be obtained at near design inlet aerodynamic conditions, which is required to determine if the second-stage is performing satisfactorily and to provide a basis for a first-stage rotor redesign.

Mechanical Test Results

During the second rig test with the hub relief on the first-stage rotor, the compressor mechanical performance was again satisfactory. The only problems experienced were a result of a relatively strong, sus tained surge at 103-percent mechanical speed. Running at this speed (equivalent to 100-percent corrected speed with heated inlet) was necessary to avoid the speed range within which blade vibration had been experienced.

The only damage to the compressor resulting from the surge consisted of heavy rubs on the second-stage abradable shroud and on the first- stage labyrinth seal silver rub ring (Figure 44).

59

H

U 05

D 70 z ä u E u. u y60

< s a < 50

^ 100%

60% 80%

3.5

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DC H

£2-0

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RRECTED SPE ED)

1.5 2.0 2.5 3.0

CORRECTED AIRFLOW

3.5 4.0

WaVo" 4.5 5.0

lb/sec

Figure 43. Advanced Two-Stage Axial Compressor Rig Test - Hub Relief.

60

:

U >>

XI nj

a o

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§

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s u 1

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d CO .L U)

M ■iH

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ä ■H

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THIRD RIG TEST OF AXIAL COMPRESSOR


The third rig test, with the hub relief and 5-degree twisted-open first-stage rotor, was conducted and data were obtained at 60, 80, 90, and 100 percent of design speed. A reduction in flow from the previous test was noticed as shown in Figure 45, The flow at 100 percent oi design speed was reduced from 4. 428 to 4.36 pounds per second. Traverse data were obtained at the following pressure ratios at design speed: 2. 56:1, 2.97:1, and 3. 13:1, shown in Appendix III. An additional traverse data point was obtained at 80-percent design speed, also shown in Appendix III. The data, in general, showed no significant change except for a further reduction in efficiency from that of the first test. There- fore, these data were used only in overall content as the basis for the redesign.

Mechanical Test Results

During the third test, no mechanical problems were encountered.

COMPRESSOR REDESIGN

Aerodynamic Redesign

Data Analysis of First Three Rig Tests. Based on the data of the first three rig tests, it was theorized and concluded that the second- stage rotor was choked and was therefore limiting the compressor flow. The problem in interpreting this theory and applying it as a basis for a redesign was twofold. It had to be determined whether the choked rotor by itself was limiting the flow for the entire compressor, or whether some phenomenon upstream of the second stage was causing the second stage to be choked and, in turn, limiting the compressor flow.

It was concluded that the low-flow phenomenon noticed on the first three rig tests was caused by a stalled or separated first-stage rotor tip. The flow in the first-stage rotor is forced to shift down towards the hub, causing a choked hub area as a result of the stalled tip area (see Figure 46). This condition in turn chokes the first-stage stator hub and stalls the first-stage stator tip. The accompanying high losses due to these flow shifts cause the second-stage rotor to be operating in choke.

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Preliminary Aerodynamic Redesign. Before the aerodynamic details of the redesign were approachea, the cause of the first-stage rotor stalled tip had to be resolved. Therefore, a detailed analysis of a family of high-pêssure-ratio axial compressor rotors (seven Continental ro4ors and the NASA rotor 2E:::) was conducted to accomplish this task and to ^rovide direction for the redesign. All of the compressors included in tinj family have demonstrated near design pressure ratio and flow test performance with the exception of the USAAVLABS rotors, which did not demonstrate design flow.

The analysis included investigations and comparisons of actual design data, free vortex design data, and test data. The design conditions for the eight rotors were mainly compared on a free vortex basis to establish equivalent design diffusion and blade loading criteria. Since the data used in the analysis are proprietary, only the overall results are presented.

Many test correlations of tip loss coefficient, tip efficiency, tip relative Mach number, aspect ratio, and so forth were tried and related to the free vortex design criteria. The only successful correlation was a relationship between test data tip performance and tip solidity. A definite trend was established that showed a significant increase in tip performance with decreasing solidity. Since, in general, a high solidity positions the adjacent blade shock intersection towards the leading edge as shown in Figure 47, severe shock boundary layer interactions may occur and in turn cause high losses with a high tip Mach number - high tip solidity axial rotor design, such as the original USAAVLABS design.

A comparison of the range of design parameters for the family of axial compressor rotors with the USAAVLABS design parameters in Table II showed that the USAAVLABS rotor is within the range of aerodynamic parameters investigated with the exception of flow rate, aspect ratio, and solidity. Since there was an axial rotor very near the flow- rate and'aspect ratio of the USAAVLABS rotor (the Continental small 2.0.1 pressure ratio axial compressor, 7. 14-pounds-per-second flow rate and 0.67 aspect ratio), only the tip solidity stands out as a possible cause of the stalled tip condition, thus supporting the correlation developed with tip solidity and tip aerodynamic performance.

Detailed Aerodynamic Redesign. The detailed aerodynamic redesign of the compressor is presented in the classified Addendum to this report, published under separate cover.

-Reference NASA Report CR-54583

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TABLE 11

i ANALYSIS OF HIGH-PRESSURE-RATIO AXIAL ROTORS |

Range of Design Parameters First Rotor Design Parameter

Covered in Analysis Design Parameter 1 Low High

Flow Rate, lb/sec 5.0 215.5 5.0 I

Pressure Ratio 1.70:1 2.05:1 1.86.1

Efficiency, percent 84 90 87

Inlet Hub/Tip Ratio 0.45 0.68 0.50 j

Tip Aspect Ratio 0.58 1.57 0.58 j

Tip Solidity 1.00 1.80 1.80

Tip Relative Inlet Air Angle*! degrees 63.9 68.3 65.5

Tip Air Turning Angle*, 1 degrees 0.9 17.3 8.0 1

Tip Pressure Ratio Head Coefficient* 0.219 0.394 0.30

Tip Relative Velocity Ratio* 0.61 0.76 0.70

Tip Axial Velocity Ratio 0.77 1.04 0.91

Inlet Tip Speed, ft/sec 1324.0 1480.0 1411.0

Tip Relative Inlet Mach Number* 1.37 1.52 1.43

Tip Diffusion Factor* 0.37 0.53 0.39 j

Static Pressure Rise/ 1 Inlet Velocity Head* 0.279 0.375 0.34

*Based on free vortex criteria.

67

MECHANICAL AND STRUCTURAL COMPRESSOR REDESIGN

The prime objective in the Phase III mechanical design was to provide a structurally sound vehicle for testing the redesign aerody- namics. The nature of the aerodynamic changes incorporated in the redesign made it feasible to use much of the Phase II hardware without modifications.

Mechanical Redesign

Figure 48 shows a comparison of the redesigned and original configurations of the compressor. The majority of the rig hardware is unchanged from the first design, consequently the assembly sequence, stack-ups, lubrication systems, and instrumentation also remain unchanged.

Table III lists the components which were redesigned to satisfy the new aerodynamic physical parameters. As seen from this table, four of the nine major components were manufactured through modification of original hardware. Since these components were some of the more complex ones in the compressor rig, substantial savings in cost and lead time were obtained. All of the other components, except those listed in Table III, were used without modification for the Phase III rig test.

Structural Redesign

The main area of structural investigation was in the rotating assembly. This included the rotor disc and blade stresses, vibratory characteristics, and shaft dynamics.

First-Stage Rotor and Blades. The integrally bladed first- stage rotor was machined from a Greek Ascoloy (AMS 5616) forging. The certified material exhibited the following physical properties at room temperature:

Ultimate tensile strength 142 ksi Yield strength at 0. 2 percent 114 ksi Endurance limit stress - smooth bar 63 ksi Endurance limit stress - notched

bar Kt = 3.6 22 ksi

Continental computer program^ based on the Manson Elastic Method, were utilized for the disc analysis.

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Rotor radial and tangential stresses are shown in Figure 49 as a function of distance from rotor centerline. The maximum radial and tangential bore stress of 38,500 psi at the design speed of 59,600 rpm is well within the allowable material limits. Average tangential stress of 35, 300 psi provided an ample burst margin of 1. 87 for the first- stage rotor.

The structural analysis of the first-stage blading was performed using Continental computer methods for the gas load conditions speci- fied in Figure 50 and a design speed of 59,600 rpm.

Figures 51, 52, and 53 show blade centrifugal stresses, centrifugal untwist, and gas bending stresses, respectively. All stresses are within the AMS 5616 physical material property limits. The relatively high compressive untwist stress combined with centrifugal and gas bending stresses results in a net moderate compressive stress at the blade leading and trailing edges. A maximum combined steady stress of 76,000 psi tension occurs at the blade midchord root location. This point is s.iown on the modified Goodman diagram. Figure 54. As indicated on the diagram, the vibratory margin for the blades is high.

The torsional and bending natural frequencies of the first- stage blade are given in Figure 55. The interference diagram indicates that at design speed, no resonance will occur in any of the modes.

Second-Stage Rotor and Blades. The second-stage rotor, like the first, is integrally bladed and machined from the same AMS 5616 forging. The only change made in this rotor was the number of blades. The disc, being identical to the Phase 11 disc but carrying a lower rim load, exhibits conservative stress levels.

Figure 56 shows the disc tangential and radial stresses as derived from the Continental computer programs. The maximum radial and tangential stress, at the bore, is 48, 500 psi at 59,600 rpm. An average tangential stress of 43, 500 psi results in a conservative burst margin of 1.69 for the rotor.

The second-stage blading is physically identical to the Phase II blades. However, the decrease in number of airfoils results in an increase in gas loading per blade (Figure 57) and consequent changes in stress levels.

Design speed analysis of the blades resulted in centrifugal untwist and in gas bending stresses as shown in Figures 58, 59, and 60 respectively. None of the stresses exceed the safe operating limits.

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Figure 59. UbAAVLABS Phase III Compressor - Second-Stage Rotor Blade Centrifugal Untwist Stress.

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At the blade root trailing edge location, a maximum combined steady stress of 47,000 psi tension exists. Figure 61 shows this maximum stress point on the modified Goodman diagram and indicates a satisfactory vibratory margin for the second stage blades.

The interference diagram in Figure 62, showing the natural frequencies of the blades, indicates that no resonance will occur at design speed.

Shaft Dynamics. The shaft dynamics of the Phase III USAAVLABS compressor configuration were analyzed to determine if the original main bearing spring cages could be utilized or if one or both of the supports would have to be redesigned to accommodate the change in mass and moment of inertia resulting from the new rotor geometry.

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Using the original bearing support spring rates and the new rotor mass parameters as input, the first three modes of lateral vibra- tior were calculated by computer techniques using the Prohl-Myklestad Hölzer type analysis. The results are tabulated below:

First Critical Second Critical Third Critical

6, 800 rpm 12,700 rpm

132, 000 rpm

With the operating range of the compressor between 20, 000 and 60,000 rpm, it was concluded that the existing main bearing spring supports would function satisfactorily with the new rotor configuration.

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FABRICATION OF REDESIGNED COMPRESSOR

The redesigned components for the Phase III compressor and the nature of their redesign are listed in Table III. As shown in Table UI, some parts were fabricated by modifying Phase II hardware in addition to total fabrication of items such as the rotors.

Fabrication of some of the more critical components is covered in the following paragraphs.

ROTOR ASSEMBLY

Because of the overall similarity of the redesigned rotor to the original rotor, the integrally bladed rotors were machined by the same methods as used previously. On the second-stage rotor, where the only change was the number of blades, it was possible to use much of the original tooling.

The two rotors and the shaft were electron-beam welded into an assembly, using the same procedure as was used in welding the first two rotor assemblies. The rotor weldment was acceptable, with no excessive runout (see Figure 63).

STATOR ASSEMBLY

Increasing the tip diameter of the first-stage rotor required that corresponding changes be made to the stationary abradable rub shroud, integral with the first-stage stator. In addition to machining the diameter of the shroud, the flame-sprayed aluminum insert was replaced to provide proper clearances (see Figure 64).

Also, for aerodynamic reasons, the top 20 percent of the first- stage stator vane leading edge was closed, varying from 8. 5 degrees at the tip to 0 degrees at 80-percent vane length. This was accomplished by making an EDM cut at the vane tip extending downstream approximately 50 percent of the chord length. Through the use of a fixture, the vane was then bent the appropriate amount (see Figure 65). The EDM slots were then filled with epoxy and blended to provide a smooth flowpath.

No modifications were made to the second-stage stator.

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INLET GUIDE VANE ASSEMBLY

The increased rotor tip diameter required that the outer flowpath diameter in the variable inlet guide vane (VIGV) assembly also be increased. It was possible to maintain the original vanes and actuation sys tern, although the vanes are cantilevered through the outer flowpath, by modifying or replacing the bushings and spacers in the vane retention system. Figure 66 shows the modified VIGV assembly as it will be installed for the final rig test.

Figure 66» Redesigned Variable Inlet Guide Vane Assembly.

87

FINAL RIG TEST OF REDESIGNED COMPRESSOR

AERODYNAMIC TEST RESULTS

The compressor was assembled with the long transition duct and with the variable inlet guide vanes. The compressor was tested with the instrumentation as defined in Volume I except for the traverse probes.

The redesigned compressor test demonstrated sufficient performance to provide a potential for a 0. 457-pound-per-horsepower-hour specific fuel consumption engine. Figure 67 represents the performance of an engine using the USAAVLABS centrifugal techrology, Figure 68, and the conventional engine component characteristics listed below. The calculation of specific fuel consumption is based on an unregenerated free shaft power turbine engine cycle which is operated as follows:

1. The centrifugal compressor, Figure 68, runs at 100-percent mechanical speed.

2. The total cycle pressure drop is 11.5 percent:

Inlet Combustor Exhaust

Turbine efficiencies are:

High Pressure Low Pressure Power Turbine

0.5 percent 4.5 percent 6.5 percent

83.9 percent 86.0 percent 88. 1 percent

4. Combustion efficiency is 98. 5 percent.

T. Output mechanical efficiency is 98 percent.

6. Turbine inlet gas temperature is 2500oF.

7. Air cooling bleed is 2 percent from centrifugal compressor discharge into the exhaust duct.

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The Effect of Overall Axial Centrifugal Compressor Pres sure Ratio and Axial Compressor Performance on Brake Specific Fuel Consumption.

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The compressor demonstrated the following performance at design speed (performance measured from compressor inlet to transition duct exit):

Pressure ratio 3.0:1 3.1:1 3.2:1 Efficiency, percent 79.0 80.0 81.0 Airflow, lb/sec 4.920 4.910 4.865

The overall compressor characteristics, Figure 69, are probably more than adequate for use in an engine. As can be seen in Figure 69 , a peak efficiency of 84 percent was obtained ata 2. 3:1 pressure ratio. A stall margin (defined in First Rig Test Section) of 10. 1 percent was obtained at a 3. 1:1 pressure ratio.

The transition duct exit pressure ratio and efficiency radial profiles at 100 percent of design speed are presented in Figure 70. As can be seen, the pressure ratio profile is reasonably flat while the efficiency profile falls off towards the shroud as a result of higher losses at the tip of the compressor. The pressure ratio and efficiency radial profiles translate to the velocity and Mach number profiles shown in Figure 71. These data were obtained by assuming a linear gradient of measured static pressure from hub to tip. The radial profiles at the transition duct exit are, in general, skewed from hub to tip. This condition can be improved through minor development, if necessary. However, these profiles do provide acceptable inlet conditions to a centrifugal compressor inducer, as shown in Figure 7 2. This figure presents inlet velocity triangles to a centrifugal compressor inducer running at the same mechanical speed (single spool) as the axial compressor. The inducer tip inlet relative Mach number is 0.88 and the tip inlet relative flow angle is 58.8 degrees, both well within conventional inducer design limits.

Since this axial compressor was designed for a specific engine application using a centrifugal compressor, the design rotational speed was maximized to provide for as high a centrifugal specific speed as practical for single spool application. Figure 73 shows the effect of the overall axial-centrifugal compressor pressure ratio and centrifugal specific speed. This figure is for a family of centrifugal compressors capable of match behind the axial compressor running at design speed and at a 3. 1:1 pressure ratio. If, for example, an overall axial-centrifugal pressure ratio of 15:1 is desired, the required centrifugal pressure ratio would be 4.85:1 and the specific speed would be 7 1. 2.

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Cx, Mx 1.945 Inch Radius

Cx, Mx 1.500 inch Radius

ß\ =54.0° Cx " 400 ft/sec

MIR -0.525 Mx -0.309

U " 550 ft/sec Cx, Mx 1.056 Inch Radi IUS

Figure 72. Two-Stage Axial Compressor Transition Duct Exit Triangle for a Centrifugal Compressor Inducer Inlet at Design RPM.

95

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Corrected Flow = 4.91 lb/$ec

Comcted Speed = 59,600 RPM

10 12 14 16 18

OVERALL AXIAL - CENTRIFUGAL PRESSURE RATIO

Figure 73. Parametric Results of a Family of Centrifugal Compressors Operating in Single-Spool Configuration Behind USAAVLABS Two-Stage Axial Compressor.

96

(

■■

An investigation of the transition duct loss using the averaged total pressure test data, Figure 74, showed that the duct loss at the 3.1:1 pressure test point was Z. 25 percent. It is believed that the actual loss is more in the order of 1 percent. The transition duct inlet total pressure test dat.i are based on a rake of three probes circumferentially located midway between adjacent stage two stator vanes, and therefore, records main passage pressure. Thus, the pressure rake does not meas ure the true mass averaged pressure. Additional probes on the suction surface and pressure surface of the vanes were not installed because of possible blockage effects. A compressor efficiency (measured from compressor inlet to second-stage stator exit) of 82. 5 percent at 3. 17 pressure ratio is calculated if the measured pressure is assumed.

The measured tip static pressure distribution throughout the compressor at 100 percent design speed is compared to the design static pressure distribution in Figure 75. The data for run number 14, at 3.015 pressure ratio, compare very well with the design tip static pressure values except for the exit of rotor one. The higher measured static pressure values at this location are possibly attributable to tip losses higher than those of design rotor one. However, in the absence of traverse data, no definite conclusion can be reached as to the cause of the higher pressure. There are indications that the second-stage rotor choking condition has been relieved. This is shown by the larger spread in static pressure at the inlet to the second-stage rotor as compared to the original design measured static pressure distribution at this location.

The transition duct measured static pressure values are compared to the design values in Figure 76. The minor differences between design and test static pressure are probably attributable to the skewed radial velocity gradient at the compressor exit.

The effect of the variable inlet guide vane on design speed compressor performance is shown in Figure 77. A peak efficiency of 81. 5 percent was recorded at +10 degrees stagger angle. As can be seen in Figure 78, a 30-degree change in stagger angle produced a 3.5-percent change in surge flow at design speed. A more dramatic change in flow was expected with the part-span guide vanes based on results of previous Continental testing of another transonic axial compressor. Those data may indicate that the first stage is not limiting flow and, therefore, making the compressor insensitive to inlet guide vane changes.

07

Figure 74.

Nominal Inlet Guide Vane Stagger Angle, 100% Detign Speed

2.9 3.0 3.1 3.2 3.3

PRESSURE RATIO (FROM INLET TO TRANSITION DUCT EXIT)

Advanced Two-Stage Axial Compressor Transition Duct Total Pressure Loss.

A compressor test at 50, 70, 90, and 100 percent of design speed was conducted with the variable inlet guide vanes set at +20 degrees stagger angle. The data from this test, shown in Figure 79. as compared to the nominal inlet guide vane data, showed a loss in efficient flow range. Therefore, the nominal guide vane setting angle provides the best compressor performance at both high and low speeds. The difference in performance shown on Figure 79, is due to the rematch between stages as a result of inlet guide vane swirl. An increase in part speed performance with inlet guide vane swirl is expected after the compressor is fully developed.

MECHANICAL TEST RESULTS

During the final test series, the compressor rig, incorporating the redesigned hardware, exhibited excellent mechanical integrity.

98

The only minor problem that developed during running was erratic vibration readings at 100 percent design speed. These readings were observed on the vertical accelerometer mounted on the outside of the compressor housing. The horizontal accelerometer showed no vibration. Also, the strain gages on the front bearing cage gave no indication that the rotor was vibrating relative to the housing. Therefore, it was concluded that the compressor assembly as a whole was being excited.

Although the amplitude was not excessive, refrigerated inlet was used at the higher speed lines to keep the mechanical speed below 9 5 percent and, consequently, out of this vibration range.

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Figure 77. Transonic Two-Stage Axial Compressor - The Effect of Variable Inlet Guide Vane Setting Angle on 100-Percent Design Speed Compressor Performance.

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VARIABLE INLET GUIDE VANE SETTING ANGLE - DEGREES

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Figure 78. Transonic Two-Stage Axial Compressor - The Effect of Variable Inlc*' Guide Vane Setting Angle on 100-Percent Design Speed Compressor Performance.

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CONCLUSIONS

1. The performance of the redesigned axial compressor has exceeded the contractural aerodynamic pressure ratio and efficiency goals. A potential for a 0. 457-pound-per-horsepower- hour specific fuel consumption turboshaft engine at 2500oF turbine inlet gas temperature using the projected USAAVLABS advanced centrifugal technology and conventional engine component characteristics was demonstrated. The feasibility of a 17:1 overall axial-centrifugal pressure ratio was shown.

Z. The redesigned compressor configuration showed a significant increase in design and part power performance level, over that of the original design compressor.

Original Design Redesign

Percent Corrected Speed 100 80 100 80

Airflow 4.36 3.29 4.91 3.72

Pressure Ratio 3.0:1 2. 1:1 3. 1:1 2. 1:1

Percent Efficiency 72. 5 76 80 H4

3. The performance increase was attributed to the reduced solidity in the first- and second-stage rotors of the redesigned compressor.

4. The mechanical design of the compressor and the rig proved to be extremely reliable over the entire 97 hours of rig running.

5. Through development, the compressor should be capable of even higher efficiency levels than those that were demonstrated.

6. The exact performance contribution lor each of the combined variations such as the solidity, aspect ratio, and blade shape, made to the redesigned compressor is not known at this time. Time and funding precluded an independent analysis of each of the variations.

104

RECOMMENDATIONS

The compressor should be developed to inc^ âse efficiency and part-speed flow range. Traverses of the redesigned compressor should be conducted to determine interstage performance and to provide direction for any modifications.

The following additional tasks are recommended to increase and more closely define the performance of the compressor;

1. A series of first-stage tests should be conducted with varying solidity of the first-stage rotor to extend the data used for redesign and thus determine the optimum solidity for this type of rotor configuration.

2. The blade shape of the first-stage rotor should be changed in an attempt to minimize shock losses.

3. A series of rotor tip clearances tests should be conducted to determine the optimum tip clearance.

-

. 4. A straight transition duct should be tested to evaluate the basic compressor performance and the long transition duct losses.

5. The short transition duct should be evaluated.

105

APPENDIX I

DESCRIPTION OF TRAVERSE DATA COMPUTOR OUTPUT

OUTPUT NOMENCLATURE AND UNITS

The following is a listing of output quantities in the sequence of their appearance in the program output. For each quantity, the sym- bol is given as it appears in output, and the quantity is defined as to its meaning and units.

The output appears under one of the following three row desig- nations: INLET, ROTOR, or STATOR. Under any of these readings, two types of output appear. The first is the output for each streamline; flow characteristics, properties, and geometry are described on a streamline or incremental basis within the flow field at particular radial stations. The second is the row output summary; interstage flow performance and geometric properties are summarized on an overall basis for the row. Appropriate quantities, given also on a streamline basis, are mass averaged in the summary.

As mentioned previously, the three types of rows, or axial stations, within a compressor are INLET, ROTOR, and STATOR. INLET is self-descriptive; it refers to the initial axial station considered, and all quantities given apply to the inlet stations. ROTOR refers to an axial station or row which, in terms of radial specification of streamlines, is considered to be at a rotor exit; however, both rotor inlet and exit quantities appear in ROTOR output. Similarly, STATOR refers to stator exit with respect to the definition of radial stations; as for a rotor, stator inlet and exit quantities are given under STATOR.

The three types of axial stations or rows lend themselves to a form of subscripting. Symbols for output quantities may contain a num- eric character for row designation according to the following convention:

1. Refers to compressor inlet row or inlet to first stator. 2. Refers to a rotor exit and stator inlet row. 3. Refers to a stator exit and/or rotor inlet row.

OUTPUT;

INLET TO COMPRESSOR:

FLOW Absolute Airflow lb/ sec

)07

OUTPUT: (Continued)

INLET TO COMPRESSOR: (Continued)

RPM Actual Speed

POl

TOl

PS

TS

ALI

EPS1

PERL

DW1

revolutions/minute

Inlet Total Pressure

Inlet Total Temperature

Inlet Static Pressure

Inlet Static Temperature

Inlet Air Angle

psia

0R

psia

0R

degrees

Streamline Angle With Respect to Axis at Inlet degrees

Percent Radial Height (From Hub) %

Rl Radius at Inlet inches

R/RT Streamline Radius/Actual Tip Radius -

CXI Inlet Axial Air Velocity ft/sec

CU1 Inlet Tangential Air Velo- city ft/sec

CR1 Inlet Radial Air Velocity ft/sec

CM1 Inlet Meridional Air Velocity ft/sec

Ul Blade Velocity Based on Radial at Inlet ft/sec

CA1 inlet Absolute Air Velocity ft/sec

MIA Inlet Absolute Mach No. .

Incremental Flow Rate Between Streamlines ibm/ sec

108

OUTPUT: (Continued)

•

INLET TO COMPRESSOR: (Continued)

RC Radius of Curvature of Streamline

WCR1 Corrected Inlet Flow Rate

NCRI Corrected Wheel Speed

WC/A1 Ratio of Corrected Inlet Flow to Actual Inlet Area

POA Mass Averaged Total Pressure

TOA Mass Averaged Total Temperature

PHI Ratio of Inlet Axial Air Velocity to Blade Velocity at Mean Radii

HUB/TIP Ratio of Hub Radius to Tip Radius Actual

AREA Inlet Annular Area, Actual

AREAE Inlet Annual Area, Effective

CP Constant Pressure Specific Heat

inches

Ihm/sec

rpm

Ihm/ sec

psia

JR

inches^

inches^

Btu/lbm0R

GAMMA Ratio of Specific Heats

ROTOR OUTPUT:

POIR Total Pressure at Rotor Inlet, Relative to Rotor psia

P02R Total Pressure at Rotor Exit, Relative to Rotor psia

109

OUTPUT: (Continued)

ROTOR OUTPUT: (Continued)

TOIR

T02R

PS2

ZR

Total Temperature at Rotor Inlet. Relative to Rotor 0R

Total Temperature at Rotor Exit, Relative to Rotor 0R

Static Pressure at Rotor Exit psia

Rotor Loss Coefficient POIR - POZR POIR - PS1

PERL2 Percent Radial Height at Rotor Exit %

R2 Radius at Rotor Exit

R/RT Radius at Streamline Divided by Tip Radius

Bl Inlet Air Angle, Relative to

POIR - PS1

DEQUIV Equivalent Diffusion Factor

inches

Rotor degrees

THETA Flow Turning Angle, (Bl - B2) degrees

B2 Exit Air Angle, Relative to Rotor degrees

DBx Incidence Angle. (Bl - Bl*) degrees

SLD Solidity, Ratio of Chord to Spacing -

DFACTR Rotor Diffusion Factor -

DP/QR PS2 * PS1 for Rotor

110

OUTPUT; (Continued)

ROTOR OUTPUT:(Continued)

DW2 Incremental Flow Rate Between Streamlines at Rotor Exit

Bl Angle Between Tangent to Blade Mean Camber Line and Axis at Inlet

THETA* Blade Turning Angle (Bl* - B2*)

B2 Angle Between Tangent to Blade Mean Camber Line and Axis at Exit

DEV Deviation Angle, (B2 - B2*)

EPS2 Streamline Angle With Respect to Axis, at Rotor Exit

RC2 Radius of Curvature of Stream- line at Exit

F-TANG Tangential Force on Blades

F-AXIAL Axial Force on Blades

R-STRESS Radius at WhichForces are Given

lb/ sec

degrees

degrees

degrees

degrees

degrees

degrees

Ibf

Ibf

inches

MIR

M2R

WIR

W2R

CX2

Inlet Relative Mach No.

Exit Relative Mach No.

Inlet Air Velocity, Relative to Blade

Exit Air Velocity, Relative to Blade

Exit Axial Air Velocity

1 11

ft/ sec

ft/sec

ft/sec

OUTPUT; (Continued)


WU2 Exit Tangential Air Velocity ft/sec

Exit Meridional Air Velocity ft/sec

Exit Radial Air Velocity ft/sec

Exit Blade Velocity ft/sec

Stage Total Pressure Ratio

Stage Total Temperature Ratio -

State Efficiency %

CM2

CR2

U2

PRS

TRS

EFFS

PRC

TRC

Cumulative Total Pressure Ratio

Cumulative Total Tempera- ture Ratio

EFFC Cumulative Efficiency %

MX2 Axial Mach No. at Rotor Exit

CX2/CX1 Ratio of Exit to Inlet Axial Velocity

WCR2 Corrected Flow Rate, at Rotor Exit Ibm/sec

NCR2 Corrected Wheel Speed, at Rotor Exit '/ rev/min

WC/A2 Ratio of Corrected Exit Flow Rate to Actual Rotor Exit Area Ibm/sec it*

PRSA Mass Averaged Stage Pressure Ratio

TRSA Mass Averaged Stage Tempera- ture Ratio

112

OUTPUT: (rontinued)


EFFSA Mass Averaged Stage Effi- ciency

PRCA Mass Averaged Cumulative Pressure Ratio

TRCA Mass Averaged Cumulative Temperature Ratio

EFFCA Mass Averaged Cumulative Efficiency

P02A Mass Averaged Total Pressure at Rotor Exit

T02A Mass Averaged Total Temperature at Rotor Exit

PHI2 Flow Coefficient - Ratio of Rotor Exit Axial Air Velocity to Rotor Exit Mean Blade Velocity

%

psia

PSI2 Pressure Coefficient -

AREA2 Actual Rotor Exit Area inches^

AREE2 Effective Rotor Exit Area inches^

HPS Stac,e Horsepower (Absorbed) hoi sepower

HPC Cumulative Stage Horsepover horsepower

CP Specific Heat Constant Pressure Btu/lbm 0R


STATOR OUTPUT:

i

P02A Absolute Total Pressure at Stator Inlet

113

psia

R3 Radius, at Stator Exit

R/RT Radius at Streamline Divided by Tip Radius

AL2 Stator Inlet Air Angle

THETA Stator Flow Turning Angle

AL3 Stator Exit Air Angle

DAL2 Stator Incidence Angle

SLD Solidity; Ratio of Vane Chord to Spacing

DFACTS Stater Diffusion Factor

DP/QS PS3 - PS2

psia

OUTPUT: (Continued)

STATOR OUTPUT: (Continued)

P03A Absolute Total Pressure at Stator Exit

T023A Absolute Total Temperature through Stator (at both Inlet and Exit)

PS3 Static Pressure at Stator Exit psia

ZS Stator .Loss Coefficient

DPO/P Ratio of Total Pressure Loss Across Stator to Stator Inlet Total Pressure

PERL3 Percent Length (from Hub) at Stator Exit

inches

degrees

degrees

degrees

degrees

PS 2A - PS2

DEQU1V Equivalent Diffusion Parameter -

1 14

OUTPUT; (Continued)


DW3

A LZ*

THETA:

AL3*

Incremental Flow Rate Between Streamlines, at Stator Exit Ibm/sec

Stator Inlet Metal Angle; Angle between Tangent to Vane Element Mean Line at Leading Edge and ^xis degrees

Vane Camber or Turning Angle degrees

Stator Exit Metal Angle; Angle between Tangent to Vane Element Mean Line at Trailing Edge and Axis degrees

DEV

EPS3

RC3

Stator Deviation Angle degrees

Streamline Angle With Respect to Axis, at Stator Exit degrees

Radius of Curvature of Stream- lines at Stator Exit inches

F-TANG Tangential Force on Blades lb

F-AXIAL Axial Force on Blades lb

R-STRESS Radius at WhichF-TANG and F-AXIAL Are Given

M2A

M3A

C2A

C3A

CX3

Stator Inlet Absolute Mach No.

Stator Exit Absolute Mach No.

Stator Inlet Absolute Air Velocity

Stator Exit Absolute Air Velocity

Stator Exit Axial Air Velocity

inches

ft/ sec

ft/sec

ft/sec

1 15

OUTPUT: (Continued)


CU3 Stator Exit Tangential Air Velocity

CM3 Stator Exit Meridional Air Velocity

CR3 Stator Exit Radial Air Velo- city

U3 Blade Velocity, Based on R 3, of Next Rotor

PRS Stage Total Pressure Ratio

TRS Stage Total Temperature Ratio

EFFS Stage Efficiency

PRC Cumulative Total Pressure Ratio

TRC Cumulative Total Tempera- ture Ratio

EFFC Cumulative Efficiency

MX3 Axial Mach No. at Stator Exit

CU2 Stator Inlet Tangential Veloc- ity

WCR3 Corrected Flow Rate, at Stator Exit

NCR2 Corrected Wheel Speed, at Stator Exit

WC/A3 Ratio of Corrected Exit Flow to Actual Stator Exit Area

1 16

ft/sec

ft/sec

ft/sec

ft/sec

ft/sec

Ibm/sec

rev/min

Ibm/sec ft

OUTPUT: (Continued)


PRSA Mass Averaged Stage- Pressure Ratio

TRSA Mass Averaged Stage Tempera ture Ratio

EFFSA Mass Averaged Stage Efficiency

PRCA Mass Averaged Cumulative Pressure Ratio

%

TRCA Mass Averaged Cumulative Temperature Ratio

EFFCA Mass Averaged Cumulative Efficiency

POiA Mass Averaged Total Pressure at Stator Exit

T03A Mass Averaged Total Temperature at Stator Edge

PHI3 Flow Coefficient - Ratio of Stator Exit Axial Air Velocity to Rotor Inlet Mean Blade Velocity

PSI3 Pressure Coefficient

AREA3 Stator Exit Area, Actual

AREE3 Stator Exit Area, Effective

psia

R

inches^

inches

CP Specific Heat Constant Press- ure Btu/lbm 0R


117

APPENDIX II

COMPUTOR OUTPUT OF TRAVERSE DATA FOR COMPRESSOR

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137

APPENDIX III

COMPUTOR OUTPUT OF TRAVERSE DATA FOR COMPRESSOR REDESIGN

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52

157

1.

Unclassified Security CU»»ific»tion

DOCUMENT CONTROL DATA R&D (Socurltr eU»atHcmliort at lilt», body of mbmtrmct mnd lnd»*fng mnnolmtlon muml b» mnffd whmn IM ovmimll rmpott I» ctmtmlltmd)

l ORI6INATINS ACIIVIIY (Corpormlm mulhor)

Continental Aviation and Engineering Corporation Detroit, Michigan

U. RtPORT tCCURITV CLA»IFICATION

Unclassified

i mfomr TITLC

lb GROUP

ADVANCEMENT OF SMALL GAS TURBINE COMPONENT TECHNOLOGY

Final Report I f 7Vp# of rmpofl and IneltMlvm dsMa)

t- AuTMOmiH (Pint namt, mlddlm Initial. Imtlnmm»)

James V. Davis Edmund J. Dellert

• RIPORT 0»TI

February 1970 •«. CONTHACT OA GRANT NO.

DA 44-177-AMC-296(T) » PROJtC T MO.

1G162Z03D14413

7«. TOTAL NO. OP PAGCS

171 7». NO OF R«F«

M. ORIGINATOR'« RCPOPT NUPrfBERISI

USAAVLABS Technical Report 69-10B

tb OTHER REPOR r NO( Jl (Any othat numban ttal may ba aaal0tad Ulla raport)

10 DISTRISUTION tTATKMCNT This document is subject to special export controls, and each transmittal to foreign governments or foreign nationals may be made only with pr'or approval of U. S. Army Aviation Materiel Laboratories, Fort Eustis, Virginia 23604,

Volume II of a 2-volume report

»rr

12. IPONSORINC MIL) T ARV ACTIVITY

U.S. Army Aviation Materiel Laboratories Fort Eustis, Virginia

This report presents the redesign analysis of a two-stage axial compressor program for thi advancement of small gas turbine component technology. The discussion covers aorication, test, and redesign of the axial compressor which was presented in Voume I.

The Continenta.-redesigned compressor demonstrated a potential for a 0. 457-pound- per-horsepower-hour SFC turboshaft engine at 2500oF turbine inlet temperature. It exceeded the contract objective by demonstrating 80 percent efficiency at 3. 1:1 pressure ratio with a 4. 91-lb/jec airflow.

f\M fUM 4 M ^1 «■»».*€■» DO roMM ft». ■ JAM U, , WHICH !■ Unclassified Socurity Claxincatioa

Unclassif ietl 8«curity CU»»lflc«tlon

KIV wonot

Axial Compressor Design Component Technology Small Gas Turbine

nOLC N r

Unclassified ••rurltT CUtainrallon

UNCLASSIFIED AD NUMBER LIMITATION CHANGEStion (SFC) turboshaft engine at 2500oF turbine inlet gas temperature would be provided. This volume discusses the fabrication, tests, and redesign

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