Reid A. Berdanier and Nicole L. Key

Purdue University, West Lafayette, Indiana

An Experimental Investigation of the Flow Physics 

Associated With End Wall Losses and Large

Rotor Tip Clearances as Found in the Rear Stages

of a High Pressure Compressor

NASA/CR—2015-218868

September 2015
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Reid A. Berdanier and Nicole L. Key

Purdue University, West Lafayette, Indiana

An Experimental Investigation of the Flow Physics 

Associated With End Wall Losses and Large

Rotor Tip Clearances as Found in the Rear Stages

of a High Pressure Compressor

NASA/CR—2015-218868

September 2015

National Aeronautics and

Space Administration

Glenn Research Center

Cleveland, Ohio 44135

Prepared under Contract NNX11AI59A
Acknowledgments

This work was supported by the National Aeronautics and Space Administration (NASA) under the ROA-2010 NRA of the 

Subsonic Fixed Wing project. The technical monitor for this project was Dr. Mark Celestina. Additional support for this project

is gratefully acknowledged from Rolls-Royce, Mr. John Fabian, Dr. Aaron King, Dr. Roy Fulayter, Dr. Bronwyn Power, Dr. Tony 

Strazisar, Dr. John Adamczyk, Dr. Hyoun-Woo Shin, Dr. Anna Young, Dr. Ivor Day, Dr. John Brossman and all other people 

who have provided helpful insights throughout this project. Signifi cant thanks are also due to Mr. Robert McGuire and the other 

support staff members at Purdue University who helped to facilitate the success of this project.

Available from

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only. Their usage does not constitute an offi cial endorsement, 

either expressed or implied, by the National Aeronautics and 

Space Administration.

This work was sponsored by the Fundamental Aeronautics Program 

at the NASA Glenn Research Center.

Level of Review: This material has been technically reviewed by NASA technical management.

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TABLE OF CONTENTS 

Page 

LIST OF TABLES ...........................................................................................................................v

LIST OF FIGURES ....................................................................................................................... vi

LIST OF SYMBOLS ..................................................................................................................... xi

LIST OF ABBREVIATIONS ...................................................................................................... xiii

SUMMARY ................................................................................................................................. xiv

CHAPTER 1: INTRODUCTION ....................................................................................................1

Authors: Reid A. Berdanier and Nicole L. Key 

1.1. Background ...........................................................................................................................1

1.1.1. Survey of Literature .......................................................................................................1

1.1.2. Characteristics of Tip Leakage Flows ............................................................................2

1.1.3. Computational Validation Using High-Speed Experimental Data ................................3

1.1.4. Spanwise Mixing ...........................................................................................................4

1.1.5. Blockage and Stage Matching .......................................................................................5

1.2. Program Objectives ...............................................................................................................5

CHAPTER 2: EXPERIMENTAL METHODS ...............................................................................7

Authors: Reid A. Berdanier, Natalie R. Smith, and Nicole L. Key 

2.1. Research Facility ...................................................................................................................7

2.1.1. Integral Facility Components .........................................................................................7

2.1.2. Additional Geometry Considerations.............................................................................9

2.1.3. Variable Tip Clearance Hardware..................................................................................9

2.2. Steady Compressor Performance ........................................................................................11

2.2.1. Corrected Operating Conditions for Humidity Effects ................................................11

2.2.2. Holistic Performance Measurements ...........................................................................13

2.2.3. Isentropic Compressor Efficiency ................................................................................16

2.2.4. Detailed Steady Total Pressure Traverses ....................................................................16

2.3. Flow Visualization Techniques ..........................................................................................17

2.4. Tip Clearance Measurement System ..................................................................................18

2.4.1. Static Tip Clearance Measurements .............................................................................18

2.4.2. Operating Tip Clearance Instrumentation ....................................................................19

2.4.3. Operating Tip Clearance Measurement Uncertainty ...................................................21

2.5. Time-Resolved Flow Field Measurements .........................................................................21

2.5.1. Over-Rotor Static Pressure Measurements ..................................................................21

2.5.2. Total Pressure Measurements at Rotor Exit .................................................................23

2.5.3. Thermal Anemometry ..................................................................................................24

2.6. Particle Image Velocimetry ................................................................................................33

2.6.1. Experimental Setup ......................................................................................................33

2.6.2. Image Processing .........................................................................................................38

CHAPTER 3: STEADY COMPRESSOR PERFORMANCE ......................................................41

Authors: Reid A. Berdanier, Natalie R. Smith, and Nicole L. Key 

3.1. Compressor Performance Maps ..........................................................................................41

3.1.1. Total-to-Total Pressure Rise Characteristics ...............................................................41

3.1.2. Operating Tip Clearance Measurements ......................................................................42

NASA/CR—2015-218868 iii
Page 

3.1.3. Effect of Ambient Conditions on Tip Clearance and Total-to-Total Pressure Rise ....46

3.1.4. Stall Margin .................................................................................................................49

3.1.5. Total-to-Static Pressure Rise Characteristics ...............................................................50

3.1.6. Isentropic Efficiency ....................................................................................................52

3.1.7. Stator 3 Seal Leakage Flow Measurements .................................................................55

3.1.8. Casing Outer Diameter Surface Temperature Measurements......................................56

3.2. Stage Performance Characteristics .....................................................................................58

3.3. Inter-Stage Flow Measurements .........................................................................................60

3.3.1. Total Pressure ...............................................................................................................60

3.3.2. Total Temperature ........................................................................................................63

3.4. Flow Visualization ..............................................................................................................66

3.5. Stator Wake Measurements ................................................................................................74

3.5.1. Steady Pressure and Temperature Rake Data ..............................................................74

3.5.2. Detailed Steady Total Pressure Traverses ....................................................................79

CHAPTER 4: CHARACTERIAZATION OF COMPRESSOR STALL INCEPTION ................84

Authors: Reid A. Berdanier, Natalie R. Smith, Anna M. Young, and Nicole L. Key 

4.1. Measurement Technique .....................................................................................................84

4.2. Stall Inception Analysis ......................................................................................................84

CHAPTER 5: TIME-RESOLVED FLOW FIELD MEASUREMENTS ......................................90

Authors: Reid A. Berdanier and Nicole L. Key 

5.1. Over-Rotor Static Pressures ................................................................................................90

5.2. Influence of Stator Wakes on Leakage Flow ......................................................................95

5.2.1. Influence of Stator Wakes on Leakage Flow Trajectory .............................................98

5.2.2. Rotor Exit Total Pressures .........................................................................................107

5.2.3. Quantifying the Size of the Leakage Flow ................................................................117

5.3. Thermal Anemometry .......................................................................................................119

5.3.1. Flow Angles and Velocity..........................................................................................119

5.3.2. Blockage ....................................................................................................................123

5.3.3. Streamwise Vorticity .................................................................................................128

CHAPTER 6: PARTICLE IMAGE VELOCIMETRY ...............................................................132

Authors: Sayantan Bhattacharya and Pavlos P. Vlachos 

6.1. Stereo Results ...................................................................................................................132

6.2. Tomographic Results ........................................................................................................136

6.3. Summary ...........................................................................................................................137

CHAPTER 7: SUMMARY AND CONCLUSIONS ...................................................................138

Authors: Reid A. Berdanier and Nicole L. Key 

7.1. Overview of Methods and Findings ..................................................................................138

7.2. Recommendations for Future Research ............................................................................139

LIST OF PUBLICATIONS RESULTING FROM THIS PROJECT ..........................................140

LIST OF REFERENCES .............................................................................................................141

NASA/CR—2015-218868 iv
LIST OF TABLES 

Table .......................................................................................................................................... Page 

Table 2.1: Purdue three-stage compressor airfoil design parameters. ............................................ 7

Table 2.2: Tip clearance configurations and design intent clearance heights ................................. 9

Table 2.3: Radial distributions of rake measurement locations. ................................................... 13

Table 2.4: Capacitance probe measurements system specifications. ............................................ 21

Table 2.5: Flush-mounted sensor positions for each rotor as a percentage of axial chord. .......... 22

Table 2.6: Hot-wire calibration matrix. ........................................................................................ 27

Table 2.7: Average differences across one stator pitch for data processed with two calibration 

sets. Data collected at Stator 1 exit for TC1. ....................................................................... 31

Table 5.1: Summary of one-dimensional blockage parameters, listed in percent flow area. ..... 127

NASA/CR—2015-218868 v
LIST OF FIGURES 

Figure ......................................................................................................................................... Page 

Figure 2.1: Purdue three-stage axial compressor research facility. ................................................ 8

Figure 2.2: Compressor tip clearance casing geometry configurations. Sketch not to scale. ....... 10

Figure 2.3: Casing recess example photograph (TC2) highlighting 45 degree angles at both ends 

of recess. Flow is from left to right. ..................................................................................... 10

Figure 2.4: Missing 45 degree ramp upstream of Rotor 2, shown here for TC2. Flow is from left 

to right. ................................................................................................................................. 11

Figure 2.5: Humidity effects on measured compressor performance for a hot humid day in the 

context of the 100% corrected speedlines for the three tip clearance heights presented in 

this study. ............................................................................................................................. 12

Figure 2.6: Purdue three-stage compressor facility flowpath cross section. ................................. 13

Figure 2.7: Rake geometry including Kiel head details. ............................................................... 14

Figure 2.8: Rake positions for different tip clearance configurations. ......................................... 15

Figure 2.9: Schematic of flow visualization setup, forward looking aft. ...................................... 17

Figure 2.10: Blade-to-blade variability of tip clearance height with respect to “Blade 1” for all 

three rotor rows. ................................................................................................................... 19

Figure 2.11: Circumferential capacitance probe measurement locations. .................................... 19

Figure 2.12: Capacitance tip clearance measurement system components. (a) Probe; (b) 

Oscillator; (c) Demodulator. ................................................................................................ 20

Figure 2.13: Over-rotor static pressure measurement system. (a) Removable sensing block and 

frame; (b) Calibration chamber............................................................................................ 23

Figure 2.14: Fast-response total pressure probe. .......................................................................... 24

Figure 2.15: Purdue Compressor Research Laboratory hot-wire calibration facility schematic 

(Morrison, 2013). ................................................................................................................. 25

Figure 2.16: Hot-wire cailibration facility optional three-dimensional calibration feature. 

Adapted from Morrison (2013)............................................................................................ 25

Figure 2.17: Radial profile of inlet turbulence intensity, as measured at the AIP. ....................... 26

Figure 2.18: Angular definitions for slanted hot-wire sensor. The long prong is denoted by the 

large circle............................................................................................................................ 28

Figure 2.19: Example slanted hot-wire coefficient fit at 300 ft/s. ................................................ 29

Figure 2.20: Approximate accuracy of hot-wire processing algorithm for test-representative 

conditions at several known flow angles. ............................................................................ 30

Figure 2.21: Processed hot-wire data at Stator 1 exit for TC1 using two separate calibrations. .. 32

Figure 2.22: Schematic of PIV setup. ........................................................................................... 34

Figure 2.23: Camera, lens, and tilt adapter setup. ......................................................................... 35

Figure 2.24: Timing diagram for synchronized image capture..................................................... 36

Figure 2.25: Calibration plate mounted between blade passages. ................................................ 36

Figure 2.26: Calibration plate images from each of the four cameras. ......................................... 37

Figure 2.27: Laser reflection, lens filter characteristics, and fluorescent dye spectrum. .............. 37

Figure 2.28: Laskin nozzle used for seeding particles into upstream flow. .................................. 38

Figure 2.29: Top two camera overlap region after calibration. .................................................... 39

Figure 3.1: Compressor total pressure ratio map. ......................................................................... 42

Figure 3.2: Operating tip clearances measured using capacitance probes for TC1. ..................... 43

Figure 3.3: Operating tip clearances measured using capacitance probes for TC2. ..................... 45

NASA/CR—2015-218868 vi
Figure ......................................................................................................................................... Page 

Figure 3.4: Operating tip clearances measured using capacitance probes for TC3. ..................... 45

Figure 3.5: Overall compressor total pressure ratio discrepancy between hot and cold test days at 

a 100% corrected operating speed for TC2 and TC3. .......................................................... 46

Figure 3.6: Variations of total pressure ratio with ambient temperature for each of the three 

nominal tip clearance conditions (TC1-TC3). ..................................................................... 47

Figure 3.7: Overall total pressure ratio trends with rotor tip clearance for three tip clearance 

configurations (TC1-TC3), each for two separate ambient temperatures at NS. ................ 48

Figure 3.8: Radial total pressure profile shapes for the three tip clearance configurations with 

“hot” and “cold” ambient operating conditions at NS. ........................................................ 49

Figure 3.9: Stall margin and stalling flow coefficient effects with tip clearance height for 

specified loading conditions on the 100% corrected speedline. .......................................... 50

Figure 3.10: Total-to-static pressure rise characteristics for three tip clearance configurations. . 51

Figure 3.11: Peak total-to-static pressure rise coefficient as a function of tip clearance height. .. 52

Figure 3.12: Overall isentropic compressor efficiency at four corrected rotational speeds. ........ 53

Figure 3.13: Isentropic compressor efficiency for specified 100% speedline points as a 

function of tip clearance height. .......................................................................................... 54

Figure 3.14: Stator 3 seal leakage flow path and manifold tubes from (a) the rear bearing plate to 

(b) an orifice plate run (Brossman, 2012). ........................................................................... 55

Figure 3.15: Stator 3 seal leakage mass flow rate as a percentage of inlet corrected mass flow rate 

for each of the traversed data points in Figure 3.1............................................................... 56

Figure 3.16: Normalized surface temperatures for three loading conditions on the 100% 

corrected speedline. ............................................................................................................. 57

Figure 3.17: Normalized surface temperatures for the final “near-stall” traversed operating point 

at all four operating speeds. ................................................................................................. 58

Figure 3.18: Stage total pressure rise characteristics for the traversed map points. ..................... 59

Figure 3.19: Radial total pressure profiles at LL. ......................................................................... 60

Figure 3.20: Radial total pressure profiles at NL. ......................................................................... 61

Figure 3.21: Radial total pressure profiles at PE. ......................................................................... 61

Figure 3.22: Radial total pressure profiles at HL. ......................................................................... 62

Figure 3.23: Radial total pressure profiles at NS. ......................................................................... 63

Figure 3.24: Radial total temperature profiles at LL. ................................................................... 64

Figure 3.25: Radial total temperature profiles at NL. ................................................................... 64

Figure 3.26: Radial total temperature profiles at PE..................................................................... 65

Figure 3.27: Radial total temperature profiles at HL. ................................................................... 66

Figure 3.28: Radial total temperature profiles at NS. ................................................................... 66

Figure 3.29: Basic surface flow topology for Stator 1 at the high loading condition (a) image, (b) 

corner separations, and (c) drawn flow topology with saddle points (S) and nodes (N) 

labeled. ................................................................................................................................. 67

Figure 3.30: Surface flow visualization of corner separations for Stator 1 (left), Stator 2 (middle), 

and Stator 3 (right) at four loading conditions: NL (a-c), PE (d-f), HL (g-i), and NS (j-l) for 

TC1. ..................................................................................................................................... 69

Figure 3.31: Stator 1 at NS conditions viewed from the leading edge looking downstream. ...... 70

Figure 3.32: Flow visualization of vane corner separations at NL for TC1 (top), TC2 (middle), 

and TC3 (bottom)................................................................................................................. 71

NASA/CR—2015-218868 vii
Figure ......................................................................................................................................... Page 

Figure 3.33: Flow visualization of vane corner separations at HL for TC1 (top), TC2 (middle), 

and TC3 (bottom)................................................................................................................. 73

Figure 3.34: Stator 1 total pressure wakes and pitchwise total temperature distribution. ............ 75

Figure 3.35: Stator 2 total pressure wakes and pitchwise total temperature distribution. ............ 77

Figure 3.36: Stator 3 total pressure wakes and pitchwise total temperature distribution. ............ 78

Figure 3.37: Detailed total pressure traverses at Stator 1 exit, NL. .............................................. 79

Figure 3.38: Detailed total pressure traverses at Stator 1 exit, HL. .............................................. 80

Figure 3.39: Detailed total pressure traverses at Stator 2 exit, NL. .............................................. 81

Figure 3.40: Detailed total pressure traverses at Stator 2 exit, HL. .............................................. 81

Figure 3.41: Detailed total pressure traverses at Stator 3 exit, NL. .............................................. 82

Figure 3.42: Detailed total pressure traverses at Stator 3 exit, HL. .............................................. 83

Figure 4.1: Dynamic pressure traces at the inception of stall for TC1 at 100% . .................... 85

Figure 4.2: Dynamic pressure traces at the inception of stall for TC2 at 100% . .................... 86

Figure 4.3: Static pressure field unsteadiness over Rotor 1, presented as an RMS with respect to 

the EA, for each tip clearance configuration at NS. ............................................................ 87

Figure 4.4: Dynamic pressure traces at the inception of stall for TC1 at 90% . ...................... 88

Figure 4.5: Dynamic pressure traces at the inception of stall for TC1 at 68% . ...................... 88

Figure 4.6: Summary of stall inception trends for all tip clearance configurations at all 

investigated operating speeds. ............................................................................................. 89

Figure 5.1: Ensemble-average static pressure field over Rotor 1 for each tip clearance 

configuration at NL. ............................................................................................................. 90

Figure 5.2: Static pressure field unsteadiness over Rotor 1, presented as an RMS with respect to 

the EA, for each tip clearance configuration at NL. ............................................................ 91

Figure 5.3: Ensemble-average static pressure field over Rotor 1 for each tip clearance 

configuration at HL. ............................................................................................................. 92

Figure 5.4: Static pressure field unsteadiness over Rotor 1, presented as an RMS with respect to 

the EA, for each tip clearance configuration at HL. ............................................................ 92

Figure 5.5: Rotor 1 tip leakage trajectories for all three tip clearance configurations at three 

loading conditions on the 100% corrected speedline........................................................... 93

Figure 5.6: Static pressure field unsteadiness, presented as an RMS, over each rotor for TC2 at 

NL. ....................................................................................................................................... 94

Figure 5.7: Static pressure field unsteadiness, presented as an RMS, over each rotor for TC2 at 

HL. ....................................................................................................................................... 95

Figure 5.8: Schematic of the interaction between rotor tip leakage flow and upstream stator 

wake. (a) R2 passing in the middle of the S1 passage, (b) R2 passing through the S1 wake.

 ............................................................................................................................................. 96

Figure 5.9: Flow visualization on the casing endwall highlighting a modulation of the tip 

leakage flow for TC1. .......................................................................................................... 97

Figure 5.10: Schematic of leakage flow measurements over the rotor for two vane wake-leakage 

flow interaction positions, (a) and (b). A boxed region identifies a representative snapshot 

of the leakage flow identified by the static pressure measurements. ................................... 98

Figure 5.11: Static pressure field over Rotor 1, averaged for all 36 blades, in terms of RMS static 

pressure, at each pitchwise position across one vane passage (clockwise) for TC1 at NL.

 ........................................................................................................................................... 100

NASA/CR—2015-218868 viii
Figure ......................................................................................................................................... Page 

Figure 5.12: Static pressure unsteadiness (RMS) over Rotor 1 for vpA (a) and vpB (b) from 

Figure 5.11, and the trajectories (c). .................................................................................. 100

Figure 5.13: Static pressure field over Rotor 2, averaged for all 33 blades, in terms of RMS static 

pressure, at each pitchwise position across one vane passage (clockwise) for TC1 at NL. 

Only IGV and S1 are moved, S2 and S3 are fixed. ........................................................... 102

Figure 5.14: Static pressure unsteadiness (RMS) over Rotor 2 for vpA (a) and vpB (b) from 

Figure 5.13, and the trajectories (c). .................................................................................. 102

Figure 5.15: Static pressure field over Rotor 2, averaged for all 33 blades, in terms of RMS static 

pressure, at each pitchwise position across one vane passage (clockwise) for TC1 at NL. 

Only S2 and S3 are moved, IGV and S1 are fixed. ........................................................... 103

Figure 5.16: Static pressure unsteadiness (RMS) over Rotor 2 for vpA (a) and vpB (b) from 

Figure 5.15, and the trajectories (c). .................................................................................. 103

Figure 5.17: Static pressure field over Rotor 3, averaged for all 30 blades, in terms of RMS static 

pressure, at each pitchwise position across one vane passage (clockwise) for TC1 at NL 

using two separate vane movement techniques. ................................................................ 105

Figure 5.18: Static pressure unsteadiness over Rotor 3 for vpA (a) and vpB (b) from Figure 5.17, 

and the trajectories (c) (TC1, NL). .................................................................................... 105

Figure 5.19: Static pressure field over Rotor 3, averaged for all 30 blades, in terms of RMS static 

pressure, at each pitchwise position across one vane passage (clockwise) for TC1 at HL.

 ........................................................................................................................................... 106

Figure 5.20: Static pressure unsteadiness over Rotor 3 for vpA (a) and vpB (b) from Figure 5.19, 

and the trajectories (c) (TC1, HL). .................................................................................... 106

Figure 5.21: Schematic of leakage flow measurements at rotor exit planes for two vane wakeleakage flow interaction positions, (a) and (b). ................................................................. 107

Figure 5.22: Average Rotor 1 exit flow field, in terms of RMS total pressure, at each 

circumferential location across one vane passage (clockwise) at NL. .............................. 109

Figure 5.23: Average Rotor 1 exit flow field, in terms of RMS total pressure, at each 

circumferential location across one vane passage (clockwise) at HL. .............................. 110

Figure 5.24: Average Rotor 2 exit flow field, in terms of RMS total pressure, at each 

circumferential location across one vane passage (clockwise) at NL. .............................. 112

Figure 5.25: Average Rotor 2 exit flow field, in terms of RMS total pressure, at each 

circumferential location across one vane passage (clockwise) at HL. .............................. 113

Figure 5.26: Average Rotor 3 exit flow field, in terms of RMS total pressure, at each 

circumferential location across one vane passage (clockwise) at NL. .............................. 115

Figure 5.27: Average Rotor 3 exit flow field, in terms of RMS total pressure, at each 

circumferential location across one vane passage (clockwise) at HL. .............................. 116

Figure 5.28: Identification of defect region associated with leakage flow unsteadiness. ........... 117

Figure 5.29: Percentage of rotor passage affected by tip leakage flow: passage-averaged value 

and range. ........................................................................................................................... 118

Figure 5.30: Radial profiles of absolute yaw angle at NL. ......................................................... 120

Figure 5.31: Radial profiles of absolute yaw angle at HL. ......................................................... 121

Figure 5.32: Radial profiles of pitch angle at NL. ...................................................................... 122

Figure 5.33: Radial profiles of pitch angle at HL. ...................................................................... 122

Figure 5.34: Radial profiles of normalized axial velocity at Rotor 2 and Rotor 3 exit. ............. 123

Figure 5.35: Example of defect region identification for Rotor 2 exit (TC1, NL). .................... 125

NASA/CR—2015-218868 ix
Figure ......................................................................................................................................... Page 

Figure 5.36: Example of defect region identification for Stator 1 exit (TC1, NL). .................... 125

Figure 5.37: Radial profiles of blockage at NL. ......................................................................... 127

Figure 5.38: Radial profiles of blockage at HL. ......................................................................... 127

Figure 5.39: Contours of streamwise vorticity at Rotor 1 exit for TC1 and TC3. ...................... 129

Figure 5.40: Contours of streamwise vorticity at for TC1 at Rotor 2 and Rotor 3 exit. ............. 130

Figure 5.41: Passage-averaged maximum streamwise vorticity strength. .................................. 131

Figure 6.1: Schematic of the PIV measurement plane and corresponding coordinate system. .. 132

Figure 6.2: Spanwise volume slices of radial velocity for stereo reconstructed velocity field at 

two loading conditions. ...................................................................................................... 133

Figure 6.3: Contours of radial velocity at different span-wise location for NL. A vector field 

representing the projected velocity in the  −  plane is superimposed over the contours. 

Flow is from right to left. ................................................................................................... 134

Figure 6.4: Contours of radial velocity at different span-wise location for HL. A vector field 

representing the projected velocity in the  −  plane is superimposed over the contours. 

Flow is from right to left. ................................................................................................... 135

Figure 6.5: Spanwise volume slices of radial velocity for tomographic reconstructed velocity 

field at two loading conditions........................................................................................... 136

NASA/CR—2015-218868 x
LIST OF SYMBOLS 

Symbol Description 

 Geometric Flow Area 

 Blockage 

Chord 

 Diameter 

	 Wire Voltage 

h Enthalpy 

 Blade Span/Annulus Height 


 Mass Flow Rate 

 Rotational Speed 

 Pressure 

 Radial Coordinate 

 Resistance 

Re Reynolds Number 

 Blade Thickness 

 Temperature 

 Blade Tip Speed 

 Absolute Velocity 

 Relative Velocity 

 Axial Coordinate 

Greek 

 Absolute Flow (Yaw) Angle 

 Relative Flow (Yaw) Angle 

 Ratio of Specific Heats 



∗

 Velocity Density Deficit 

 Isentropic Efficiency 

 Vorticity 

Ω Blade Angular Velocity 

 Pitch Angle 

Φ Flow Coefficient 

Ψ Total-to-Static Pressure Rise Coefficient 

" Density 

# Wire Slanting Angle 

$ Tip Clearance Height 

 Tangential Coordinate 

Subscripts 

AA Radial and Circumferential Area Average 

act Actual Condition 

Corrected Condition 

CA Circumferential Average 

eff Effective 

leak Leakage 

NASA/CR—2015-218868 xi
Symbol Description 

Subscripts 

LN Normalized Leakage 

 Meridional Direction 

mech Mechanical 

% Stagnation Condition 

 Radial Direction 

RA Radial Area Average 

ref Reference Condition 

& Isentropic Condition 

st Streamwise Direction 

stall Stall Condition 

 Axial Direction 

) Wire 

 Tangential Direction 

NASA/CR—2015-218868 xii
LIST OF ABBREVIATIONS 

Abbreviation Description 

AIP Aerodynamic Interface Plane 

BPP Blade Passing Period 

CFD Computational Fluid Dynamics 

DCA Double Circular Arc 

EA Ensemble Average 

HL High Loading 

HPC High Pressure Compressor 

IBR Integrally Bladed Rotor 

IGV Inlet Guide Vane 

LDV Laser Doppler Velocimetry 

LL Low Loading 

LSRC Low-Speed Research Compressor 

NB Number of Blades 

NL Nominal Loading 

NS Near Stall 

OPR Overall Pressure Ratio 

PE Peak Efficiency 

PIV Particle Image Velocimetry 

RMS Root-Mean-Square 

SFW Subsonic Fixed Wing 

SLE Special Limits of Error 

SM Stall Margin 

TC Tip Clearance 

TPR Total Pressure Ratio 

TTL Transistor-Transistor-Logic 

TTR Total Temperature Ratio 

NASA/CR—2015-218868 xiii
SUMMARY 

The focus of this work was to characterize the fundamental flow physics and the overall 

performance effects due to increased rotor tip clearance heights in axial compressors. Data have 

been collected in the three-stage axial research compressor at Purdue University with a specific 

focus on analyzing the multistage effects resulting from the tip leakage flow. Three separate 

rotor tip clearance heights were studied with nominal tip clearance heights of 1.5%, 3.0%, and 

4.0% based on a constant annulus height. 

Overall compressor performance was investigated at four corrected speedlines (100%, 90%, 

80%, and 68%) for each of the three tip clearance configurations using total pressure and total 

temperature rakes distributed throughout the compressor. The results have confirmed results 

from previous authors showing a decrease of total pressure rise, isentropic efficiency, and stall 

margin which is approximately linear with increasing tip clearance height. The stall inception 

mechanisms have also been evaluated at the same corrected speeds for each of the tip clearance 

configurations. 

Detailed flow field measurements have been collected at two loading conditions, nominal 

loading (NL) and high loading (HL), on the 100% corrected speedline for the smallest and 

largest tip clearance heights (1.5% and 4.0%). Steady detailed radial traverses of total pressure at 

the exit of each stator row have been supported by flow visualization techniques to identify 

regions of flow recirculation and separation. Furthermore, detailed radial traverses of timeresolved total pressures at the exit of each rotor row have been measured with a fast-response 

pressure probe. These data have helped to quantify the size of the leakage flow at the exit of each 

rotor. Thermal anemometry has also been implemented to evaluate the time-resolved threedimensional components of velocity throughout the compressor and calculate blockage due to 

the rotor tip leakage flow throughout the compressor. These measurements have also been used 

to calculate streamwise vorticity. 

Time-resolved static pressure measurements have been collected over the rotor tips for all 

rotors with each of the three tip clearance configurations for up to five loading conditions along 

the 100% corrected speedline using fast-response piezoresistive pressure sensors. These timeresolved static pressure measurements, as well as the time-resolved total pressures and velocities 

have helped to reveal a profound influence of the upstream stator vane on the size and shape of 

the rotor tip leakage flow. 

Finally, a novel particle image velocimetry (PIV) technique has been developed as a proofof-concept. In contrast to PIV methods that have been typically been utilized for turbomachinery 

applications in the past, the method used for this study introduced the laser light through the 

same access window that was also used to image the flow. This new method addresses potential 

concerns related to the intrusive laser-introducing techniques that have typically been utilized by 

other authors in the past. 

Ultimately, the data collected for this project represent a unique data set which contributes 

to build a better understanding of the tip leakage flow field and its associated loss mechanisms. 

These data will facilitate future engine design goals leading to small blade heights in the rear 

stages of high pressure compressors and aid in the development of new blade designs which are 

desensitized to the performance penalties attributed to rotor tip leakage flows. 

NASA/CR—2015-218868 xiv
CHAPTER 1: INTRODUCTION 

Reid A. Berdanier and Nicole L. Key 

Rising fuel costs and environmental concerns are continuing to drive gas turbine engine 

development toward increased overall efficiency and decreased fuel burn. However, achieving 

these goals will likely lead to smaller gas turbine engine cores for two reasons. First, 

improvements in materials and cooling schemes allow increased turbine inlet temperatures and, 

thus, increased overall pressure ratios (OPR) that will allow for higher thermal efficiencies. For a 

given engine size, an increased OPR is most easily accommodated through added stages. Second, 

larger bypass ratios lead to increased propulsive efficiencies in turbofan engines. For large 

aircraft, current under-wing engine mounting arrangements are limited in their ability to 

accommodate larger fan diameters. For small aircraft, the nacelle is already a significant 

contributor to overall aircraft drag and an increased fan diameter could have an overall negative 

effect. As these limits for the outer diameter of a turbofan are encountered, designers are turning 

their attention to reducing the size of the core as a means for increasing the bypass ratio. Both of 

these driving factors result in smaller blading, especially in the rear stages of a high-pressure 

compressor (HPC). 

As engine cores become smaller, the flow path area decreases, and the losses in the endwall 

regions become more significant as they comprise a larger percentage of the total flow field. 

Additionally, tip clearances will not scale with the decreased blade size because of 

manufacturing tolerances and the margin needed for transient operation and thermal growth. As a 

result, a smaller core leads to rotor tip clearance heights which are large relative to the blade size. 

A larger relative tip clearance height leads to increased blockage due to the associated tip 

leakage flows, as well as potential issues with stage matching at off-design operating conditions. 

Previous research has suggested increased tip clearance-to-blade-height ratios causes decreased 

efficiency and pressure rise capability. Similarly, the ratio of clearance-to-chord may also 

increase, leading to a decrease in stall margin. Thus, small blade heights with increased relative 

clearances are expected to have a detrimental impact on HPC efficiency and operability. 

However, these concerns with large tip clearance heights relative to blade size are not 

specific to new engine designs. Existing engine designs are also subject to these potential 

performance detriments as in-service deterioration results in larger clearances in compressors. 

The development of designs that are de-sensitized to large tip clearances would provide the 

ability to avoid the penalties of reduced stall margin and reduced efficiency typically associated 

with in-service deterioration. However, the flow physics of these tip leakage flows must be better 

understood in a multistage environment to generate these new designs – an understanding which 

must extend beyond the design point to also improve off-design compressor performance. 

1.1. Background 

1.1.1. Survey of Literature 

Over the years, several dedicated literature reviews have summarized the body of work 

focused on tip clearance effects in compressors: Reeder (1968) and Peacock (1982, 1983). 

Experimental campaigns investigating several rotor tip clearances have typically been limited to 

NASA/CR—2015-218868 1
isolated rotors or single stage machines operating in a low-speed regime. In particular, the results 

published by Hunter and Cumpsty (1982), Inoue et al. (1986), McDougall (1990), and Goto 

(1992) have helped to understand the overall performance effect of increased rotor tip clearance 

for idealized environments experiencing a clean inlet flow, free from the effects of an upstream 

stage. Multistage low-speed four-stage research compressors such as the facility used by Wisler 

(1985), and a similar facility by Tschirner et al. (2006), provide the important multistage flow 

environment which is expected to affect the rear stages of an HPC. However, these low-speed 

measurements often do not accurately represent the Reynolds numbers and Mach numbers 

present in the rear stages of an HPC, nor do they incorporate appreciable compressibility effects. 

Using data collected from multistage compressors of varying designs, Freeman (1985) 

gives a thorough overview of rotor tip clearance effects in axial compressors. This information 

provides a valuable summary of overall performance effects for varying rotor tip clearance 

heights, but it does not address the underlying flow physics and stage-to-stage impact of tip 

leakage flows. 

Of the existing studies considering several rotor tip clearance heights, few have studied the 

effects of rotor tip clearance heights larger than 3% blade span. Also, many published results 

have focused efforts toward building a detailed understanding of the fundamental leakage flow 

physics at a single rotor tip clearance height, instead of combining knowledge gained from 

several clearance heights: e.g., Lakshminarayana et al. (1982a, 1982b), Shin et al. (2008), and 

Sans et al. (2013). 

Experimental results from cascades have also been beneficial to the growth of knowledge 

relating to tip clearance flows. Of note, Williams et al. (2010) recently used a cascade to 

investigate very large tip clearances up to 6% of annulus height, which is representative for the 

HPC stages of industrial gas turbines. As expected, losses increased with increasing tip clearance, 

however, losses leveled off for clearances larger than 4% annulus height. The authors 

hypothesized that for large tip clearances, the effect of the endwall became less important, and 

the flow through the tip gap behaved more like that of a wing tip vortex since the casing had less 

impact on the leakage flow. 

1.1.2. Characteristics of Tip Leakage Flows 

The static pressure difference across the rotor tip clearance is the primary mechanism 

driving the flow through the tip gap. In addition to the pressure-driven flow, the relative motion 

between the rotor blade and the stationary casing endwall also creates a shear-driven flow. 

Without a tip clearance, there exists a separation in the corner between the suction surface of the 

blade and the endwall as the low-velocity fluid near-wall flow is unable to negotiate the required 

pressure gradient as the radius of curvature decreases. The presence of a small tip clearance can 

be beneficial since it will energize the separation-prone fluid. Although some authors have 

suggested the existence of this optimal clearance height which could lead to a peak efficiency 

condition, this clearance is typically smaller than what can be achieved for mechanical reasons. 

In most applications, the tip leakage flows found in compressor rotors are responsible for entropy 

generation, loss in peak pressure rise, and decreased stall margin (Cumpsty, 2004). 

In a canonical study, Wisler (1985) measured a change in overall compressor performance 

with increased tip gap in a four-stage low-speed research compressor (LSRC) at GE. When the 

tip clearance was increased from 1.36% to 2.8% of blade height, the overall compressor 

efficiency dropped by 1.5 points, and peak pressure rise was reduced by 9.7%. Wisler also 

showed a decrease of operability range by an 11% increase of stalling flow coefficient. 

NASA/CR—2015-218868 2
Some authors have discussed the formation of a shear layer and a vortex sheet as a result of 

the leakage fluid passing through the rotor tip gap interacting with the adjacent blade passage. 

Storer and Cumpsty (1991) used cascade measurements to show that this so-called tip leakage 

vortex increased in size and strength as the clearance was increased. Once the leakage vortex 

separates from the blade surface, it may “burst,” or break down into a high-loss region creating 

significant blockage to the main throughflow. The diffusive nature of the flow field contributes 

to this breakdown as the leakage vortex grows circumferentially and radially during its travel 

through the rotor passage. While passing through the tip gap, however, the clearance flow 

experiences little loss. Instead, the primary loss mechanism is due to the diffusive mixing that 

occurs when the jet-like clearance flow issues into the main passage flow. Because the exact 

location of the vortex breakdown may be disputable, the tip leakage “vortex” will usually be 

referred to as a tip leakage “flow disturbance” throughout this report. 

Using measurements acquired on the same LSRC used by Wisler (1985), Yoon et al. (2006) 

showed that the tip leakage vortex strengthens and moves upstream as flow coefficient is reduced 

and also as tip clearance is increased. New findings from this experiment include that the tip 

vortex was weakened and moved downstream as stagger was increased. At very large clearances 

(5.5% chord), the vortex exhibited a change in trajectory because of a sudden increase in flow 

between mid-chord and trailing edge. This pushed the tip vortex toward the neighboring blade. 

This circumferential flow trajectory increases the potential existence of a double leakage 

flow condition, as introduced by Khalsa (1996). Double leakage refers to the case when the tip 

leakage flow convects across the blade passage at a high trajectory angle and passes through tip 

clearance of the adjacent blade. In this case, the low-pressure fluid contained in the tip clearance 

flow faces a potentially compounding loss as it passes through the second clearance gap. The 

presence or absence of this effect is dependent upon the loading of the blade row through its 

effect on the trajectory of the vortex and the solidity of the blade row (Dickens and Day, 2011). 

The role of the upstream stator is important for the development of the tip leakage flow. In 

particular, Mailach et al. (2008) used Laser Doppler Anemometry to investigate the tip region of 

a four-stage low-speed research compressor. The upstream stator wakes had a profound 

influence on the trajectory of the tip leakage vortex, the intensity of the tip leakage vortex, and 

the position of maximum tip leakage flow along the chord. This same interaction has also been 

addressed recently by other authors (Smith et al., 2015b; Krug et al., 2015). 

Sirakov and Tan (2003) used time-accurate Reynolds-averaged Navier-Stokes simulations 

to understand the role of unsteadiness on rotor tip leakage flows by replacing the effect of an 

upstream stator with representative wake deficits. Two situations were investigated for their 

study: a steady wake deficit and a pulsing wake deficit – the second representing what would be 

experienced by a rotating blade row. The authors observed a beneficial unsteady interaction 

between the rotor and (simulated) stator, which was attributed to the suppression of the double 

leakage phenomenon. 

1.1.3. Computational Validation Using High-Speed Experimental Data 

Detailed experimental datasets are critical for validating computational models of tip 

leakage flow effects on compressor performance. Suder and Celestina (1996) compared 

computations to experimental data obtained in the single stage transonic compressor (Stage 37) 

facility at NASA Glenn Research Center. Two loading conditions (peak efficiency and near stall) 

were investigated at two operating speeds: 60% speed with subsonic flow and 100% speed where 

the shock structure could be studied. The tip clearance at design speed was 0.5% span (0.7% 

NASA/CR—2015-218868 3
chord). Laser anemometer measurements captured the interactions between the tip leakage 

vortex and the shock structure at high speeds. These authors showed that the radial influence of 

the tip leakage flows extended over a range up to twenty times the physical clearance at 

operating speeds for which shocks were present. At the subsonic conditions, however, the radial 

influence of the leakage flow was reduced to only five times the physical clearance. 

These experimental data for Stage 37 allowed Van Zante et al. (1999) to investigate 

computational models of the tip clearance flow field focusing on the impact of grid topology, the 

numerical treatment of the tip gap, and their effects on the solution. The authors found that 

gridding the tip gap did not provide any advantages over using a simple tip clearance model, but 

the wall-bounded shear layer was an important component to the tip clearance flow, especially 

when the difference between the relative velocities of the leakage jet and casing was large. Gupta 

et al. (2003) agreed that sufficient clustering of grid points near the casing, as a method for 

capturing the shear layer, was critical to a successfully modeling the tip leakage flows. However, 

these authors were able to achieve the best results by using a square tip and fully resolving the tip 

gap rather than using approximate methods. A summary of these and other difficulties affecting 

computational models for turbomachinery applications, including tip clearance gap modeling 

challenges, is addressed by Denton (2010). 

The observations by Van Zante et al. (1999) and Gupta et al. (2003) provide an excellent 

example of how a quality experimental dataset can be used to calibrate CFD models. However, 

the isolated rotor and single stage environment does not capture all of the important flow physics 

present in a multistage compressor. In particular, spanwise mixing and endwall boundary layers 

are influenced by upstream stages. These, in addition to the tip leakage flow, are major 

contributors to blockage in the endwall region and can lead to stage matching issues (Cumpsty, 

2004). 

1.1.4. Spanwise Mixing 

Spanwise mixing is responsible for distributing the high losses associated with the endwall 

flows toward mid-span. For geometry with low aspect ratios, this spanwise mixing component 

must be accurately modeled to capture the flow field correctly. Adkins and Smith (1982) 

developed a spanwise mixing model which suggested secondary flow generated spanwise and 

cross-passage velocities to enhanced mixing, and it agreed well with several test cases. However, 

because this model was based on inviscid secondary flows, it was not adequate for cases with 

significant regions of separated flow, including high loading or other off-design cases. 

A different mechanism responsible for spanwise mixing was proposed by Gallimore and 

Cumpsty (1986). Using an ethylene tracer gas technique, these authors concluded that turbulent 

diffusion, rather than the radial velocity associated with secondary flows, was the main 

mechanism for spanwise mixing. For this analysis, the experiments were performed near a peakefficiency point, but off-design conditions with highly separated flows were not investigated. 

In fact, both of the mechanisms identified by Adkins and Smith (1982) and Gallimore and 

Cumpsty (1986) can be important. Wisler et al. (1987) used the ethylene tracer gas technique and 

hotwire anemometry to investigate mixing at design loading and high loading operating 

conditions. In the “free stream region,” they found the primary mechanism for mixing was 

turbulent diffusion. In the endwall regions and along the blade surfaces, though, secondary flow 

effects were equally responsible, if not slightly more important than turbulent diffusion. A 

follow-on computational study by Leylek and Wisler (1991) confirmed the importance of both 

turbulent diffusion and secondary flows in the spanwise mixing process. However, these authors 

NASA/CR—2015-218868 4
noted important considerations to accurately recreate spanwise mixing representative of real 

engine operating environments. Of note, the turbulence intensity associated with upstream blade 

and vane wakes must be appropriately introduced, and the blade row inlet endwall boundary 

layer must be adequately represented to achieve the secondary flow field. 

The single stage low-speed study conducted by Goto (1992) at the Whittle Laboratory 

investigated the effect of different rotor tip clearances (up to 3% chord) on the flow field in the 

downstream stator, with an emphasis on spanwise mixing. As the tip clearance increased, the 

blockage in the tip region increased and led to a decrease in wake size near the stator hub and a 

suppression of the unsteadiness in the hub region. 

1.1.5. Blockage and Stage Matching 

In compressors, blockage represents the reduction of flow area due to the existence of low 

momentum fluid. Blockage will increase the velocity in the core flow region, thereby reducing 

the work done by the rotor on the core flow. As endwall boundary layers grow at off-design 

operating conditions, stage matching problems may be introduced. However, a lack of 

understanding for this endwall boundary layer growth in multistage axial compressors also exists 

at design conditions and typically forces designers to rely on correlations, as explained by 

Cumpsty (2004). 

The most significant body of data on compressor endwall boundary layers was acquired by 

Smith (1970) in the four-stage LSRC at GE. These data showed that blockage was essentially 

unchanged across a repeating stage, and blockage was a function of tip clearance and loading 

condition. However, the repeating-stage model is valid for multistage compressors with large 

aspect ratios and relatively small clearances. The repeating stage condition was not achieved for 

an aspect ratio of 1.96, the lowest aspect ratio tested by Smith. The rear stages of high pressure 

core compressors typically have aspect ratios on the order of unity. Thus, further information 

must be collected under these conditions to supplement the data published by Smith. 

Cumpsty (1986) presented measurements from a four-stage compressor at lower Reynolds 

numbers. The data indicated that the rotor tip clearance (or stator hub clearance) was a critical 

parameter for determining the endwall boundary layer thickness. Using a model developed to 

relate the flow through the tip clearance to the downstream blockage, Khalid et al. (1999) 

proposed that the total pressure loss in the endwall region resulted from the interaction of tip 

leakage flows and passage flows, and the vortical structure associated with the leakage vortex 

was deemed less important. Khalid et al. also correlated endwall blockage with aerodynamic 

loading on the blade. Data acquired by Suder (1998) on the isolated transonic rotor (Rotor 37) 

showed that the blockage in the endwall region increased with blade loading in agreement with 

Khalid’s correlation. Further, 3D Particle Image Velocimetry (PIV) measurements collected by 

Wernet et al. (2005) in a large-scale four-stage low-speed compressor at NASA Glenn Research 

Center showed that the radial velocity component was important in determining the radial extent 

of influence of the tip leakage flows and the blockage associated with these flows.

1.2. Program Objectives 

NASA led a Turbomachinery Technical Working Group that provided a list of propulsion 

improvements necessary to meet the Subsonic Fixed Wing (SFW) project goals, specifically 

allowing the project to meet its N+1 and N+2 reduced fuel burn goals. One of the main topics on 

NASA/CR—2015-218868 5
the list was tip leakage flows in high pressure ratio cores. The group anticipates that the smaller 

cores will result in blade tip clearances of 2% to 4% of blade height, and the resulting 

compressor efficiency loss due to the leakage flow through the tip gap could be as high as 5 

points, corresponding to a specific fuel consumption (SFC) effect on the order of 3% (Heidmann, 

2009). Thus, reductions in these endwall and tip leakage losses are absolutely necessary to 

realize SFC improvements. 

The research presented in this report supports NASA’s strategic goals of reduced fuel burn 

in the SFW area by contributing to the understanding of core compressor losses associated with 

tip leakage flows and other endwall losses. Once the flow physics associated with these losses 

are understood, CFD models can be generated to accurately predict off-design performance. 

These data will be instrumental in calibrating predictive tools for multistage compressor 

performance. Before designs attempting to mitigate or desensitize compressor performance to 

large tip clearances can be successfully achieved, the flow physics associated with these large tip 

clearance flows must be well understood and accurately predicted. The data acquired in this 

project help to illuminate such flow physics. 

The objective of this research project was to provide detailed flow field measurements in a 

three-stage axial compressor typical of the rear stages in an HPC. The availability of these data 

provides an opportunity to calibrate existing CFD tools for off-design prediction and stage 

matching. The experimental campaign has measured the flow field for three different rotor tip 

clearances: a baseline case with a 1.5% tip clearance (with respect to annulus height), a 3% tip 

clearance, and a large tip clearance of 4%. A comparison of three different tip clearance 

configurations allows the performance deltas between configurations to be captured. Overall 

compressor performance maps have been acquired, as well as detailed flow field measurements. 

A combination of traditional measurement techniques such as Kiel head total pressure probe 

traverses and hotwire anemometry have been used, in addition to a novel PIV technique, which 

has proven the ability to non-intrusively interrogate 3D flow field structures in the embedded 

stage of a multistage compressor. This project represents one of the few complete databases 

available in the open literature, and it is the only database focused on multistage flow in the rear 

stages of an HPC with large relative tip clearances. 

NASA/CR—2015-218868 6
CHAPTER 2: EXPERIMENTAL METHODS 

Reid A. Berdanier, Natalie R. Smith, and Nicole L. Key 

2.1. Research Facility 

2.1.1. Integral Facility Components 

The Purdue University three-stage axial compressor research facility is a geometricallyscaled design which models the rear stages of a high-pressure core compressor. In particular, the 

facility features engine-representative Reynolds numbers and Mach numbers, which makes it a 

useful vehicle for research at a high technology readiness level. Furthermore, the intermediate 

rotational speed regime and the multistage design of the facility create unique opportunities to 

analyze an “embedded” stage and complex multistage flow effects. 

The Purdue three-stage compressor facility features a 6061 aluminum casing and all 

blading is machined from 17-4 stainless steel. The compressor is comprised of an inlet guide 

vane (IGV) followed by three stages. The inlet guide vane and the three rotor rows are doublecircular arc (DCA) designs, and the three stator vane rows are NACA 65-series airfoils. All three 

of the rotor rows are integrated bladed rotor (IBR) “blisk” designs, such that each of the three 

rotor rows represents one solid piece of material. Further, each of the stationary vane rows has 

been created as a 180-degree monolithic segment, with vanes which are fully-shrouded at the 

inner and outer diameters. Specific geometric parameters for the blading in the facility are 

outlined in Table 2.1. 

Each of the stationary vane rows are individually circumferentially indexable for up to two 

full stator vane passages, allowing circumferential flow variations to be measured with stationary 

instrumentation. Each of the vane rows is moved with a pair of linear actuators during 

compressor operation, and precision string potentiometers provide feedback control of the vane 

positions. This two-vane circumferential measurement capability was an expansion beyond 

previous capabilities which was made possible through the new compressor casing hardware 

required for this project. 

Table 2.1: Purdue three-stage compressor airfoil design parameters. 

Blade 

Row 

Camber [o] Stagger [o] Chord [in.] 

/

Airfoils +, (avg) -.

101 Hub Tip Hub Tip Hub Tip N Type Rel Abs 

IGV -12.3 -14.0 8.2 9.0 2.00 2.00 0.065 44 DCA 0.26 0.26 3.0 

R1 44.7 25.8 32.6 47.8 2.46 2.81 0.045 36 DCA 0.46 0.32 7.0 

S1 49.5 48.3 25.6 24.0 2.11 2.11 0.065 44 65 0.35 0.35 4.4 

R2 42.4 24.6 35.4 49.8 2.60 2.96 0.049 33 DCA 0.45 0.31 7.7 

S2 49.7 48.4 26.1 24.6 2.22 2.22 0.065 44 65 0.33 0.33 4.6 

R3 39.5 22.9 38.2 51.8 2.75 3.13 0.051 30 DCA 0.44 0.31 8.4 

S3 62.4 61.2 19.7 18.1 2.35 2.35 0.065 50 65 0.31 0.31 4.8 

NASA/CR—2015-218868 7
The research compressor is driven from the rear by a 1400 hp AC motor through a 5:1 

speed-increasing gearbox to provide a nominal rotational speed of 5000 rpm at the design point. 

A liquid-cooled variable frequency drive system paired with an encoder on the motor drive shaft 

provides the capability to maintain a compressor operating speed within 0.01% of the desired 

mechanical speed. This capability creates an especially unique opportunity to ensure confidence 

in the repeatability of measurements as they relate to the rotational speed of the machine. In 

addition to the motor encoder, an optical laser tachometer aimed at the high-speed output shaft of 

the gearbox creates a transistor-transistor logic (TTL) signal which is used as a once-perrevolution (1/rev) trigger for the high frequency response data acquisition systems. 

The compressor operates with ambient, unconditioned air as the working fluid. As shown 

in Figure 2.1, the air is first drawn into a large settling chamber before passing through a 

reducing bellmouth into a two foot diameter duct. A long-form Venturi flow meter designed in 

accordance with ASME PTC-19.5 (2004) flow measurement standards accurately measures the 

mass flow rate passing through the compressor. 

Figure 2.1: Purdue three-stage axial compressor research facility. 

Two Rosemount 3051C high-accuracy pressure transmitters are installed for measuring the 

differential pressure across the Venturi. Because the relative uncertainty of the differential 

pressure measurement is dependent upon the calibration range, the two 3051C transmitters were 

calibrated in different pressure ranges (0 to 5 inH2O and 0 to 8 inH2O) and are connected via a 

manifold. These ranges were selected to ensure adequate coverage of the entire range of 

operating conditions (0 to 8 inH2O), while maintaining a measurement near the peak efficiency 

point which utilizes the smallest possible measurement uncertainty. Moreover, in the low flow 

rate measurement range, the pair of devices provides a redundant measurement for added 

confidence. As a complement to the differential pressure measurements, a pitot-static pressure 

probe with an integrated T-type thermocouple is installed just upstream of the Venturi meter, and 

the measurements from this probe are used to define the density and other thermodynamic 

properties required for the mass flow rate calculation equations. 

NASA/CR—2015-218868 8
After passing through the Venturi meter, the air continues through an insulated duct section 

that is 11 duct diameters in length. A bullet-nose cone upstream of the compressor directs the 

airflow to the constant annulus flowpath, which is defined by a two-inch passage height with a 

hub-to-tip ratio of 5:6. At the exit of the compressor, a sliding-annulus throttle controls the back 

pressure of the machine, and a scroll collector directs the air to an exhaust plane at ambient 

pressure. Additional specifics defining the existing facility layout can be found in Ball (2013), 

and a thorough discussion of facility health monitoring systems is provided by Talalayev (2011). 

2.1.2. Additional Geometry Considerations 

In addition to the typical geometry parameters prescribed for the Purdue three-stage 

compressor flowpath, other pertinent information may be required to accurately build a 

computational model of the machine. Of note, the labyrinth (knife) seal clearances under the 

stator platforms have been measured as “cold” (static) values: 0.022 in. for the IGV, 0.035 in. for 

Stator 1, 0.037 in. for Stator 2, and 0.024 in. for Stator 3. Operating “hot” clearances for these 

seals have not been measured. Further, all of the stator vanes have fillet radii at both inner and 

outer diameter endwalls of 3/32 in., and the rotor hub fillet radii are all 0.150 in. The leading 

edges of the rotor blades have elliptical shape, and the stators have circular leading edges. An 

average surface roughness of 30-40 in. has been prescribed for all of the internal flowpath 

surfaces, including the blades. A past analysis using an optical scan compared the design intent 

geometry of the rotor blisks with the as-manufactured geometry and confirmed that all measured 

dimensions are within the specified design tolerances of the parts. The rotor tip clearances for 

each defined operating condition have been measured, and the data are presented in this report. 

2.1.3. Variable Tip Clearance Hardware 

A series of compressor casing hardware components were required to support the tip 

clearance analyses desired for this project. Three tip clearance configurations were used for this 

study, as outlined by the information in Table 2.2. In each configuration, one discrete tip 

clearance height exists for all three rotor rows (i.e., no combinations of tip clearance heights 

were considered). The first tip clearance configuration (TC1) represents the baseline compressor 

configuration that has been presented by previous investigators with this facility. This baseline 

configuration represents a nominal 1.5% tip clearance height as a fraction of overall annulus 

height. 

Table 2.2: Tip clearance configurations and design intent clearance heights 

Configuration Nominal Clearance 

Height, $ [in] 

Normalized Clearance 

Height, $/ [%] 

Tip Clearance 

Representation

TC1 0.030 1.5 Smooth Wall 

TC2 0.060 3.0 Casing Recess 

TC3 0.080 4.0 Casing Recess 

As shown in the cartoon of Figure 2.2, the two larger tip clearance configurations (TC2 and 

TC3) were achieved through the use of a casing recess over the rotors. These compressor 

configurations nominally represent 3.0% and 4.0% tip clearance height. Although the differences 

NASA/CR—2015-218868 9
between casing recesses and traditional cropped rotors have been documented in the past (Wisler 

and Beacher, 1989), the parametric analysis discussed by Brossman (2012) showed through 

computational simulation of the Purdue three-stage axial compressor that the overall 

performance and the individual stage performances of a compressor with these types of casing 

recesses – which are introduced as far as possible upstream and downstream of the rotor – 

sufficiently represent the trends of the same machine with a smooth wall and a cropped rotor. 

This, combined with the budget constraints involved with a cropped rotor study, led to the choice 

of using recessed casings to achieve the two larger clearance configurations. 

Figure 2.2: Compressor tip clearance casing geometry configurations. Sketch not to scale. 

Figure 2.3: Casing recess example photograph (TC2) highlighting 45 degree angles at both ends 

of recess. Flow is from left to right. 

Because the emphasis of this study was not on the effects of trenched casings as discussed 

by Wisler and Beecher (1989), the axial extent of the trenches was increased as much as 

allowable to the adjacent stator vanes (approximately 0.4 axial chords upstream and downstream 

of the rotor blades). Furthermore, 45 degree angles were introduced to ease the transition from 

the 2-inch annulus height of the baseline compressor defined by the stator vanes and the 

increased space over the rotors. The introduction of these angle features facilitates grid topology 

continuity for computational comparison studies. In addition to the cartoon of the casing 

configurations in Figure 2.2, a photographic example of the casing recess for TC2 is shown in 

Figure 2.3. It is important to note here that a manufacturing error led to the removal of the 45 

degree ramp from the recess at the upstream side for Rotor 2, as shown in Figure 2.4, for both 

TC2 and TC3. 

NASA/CR—2015-218868 10
Figure 2.4: Missing 45 degree ramp upstream of Rotor 2, shown here for TC2. Flow is from left 

to right. 

To ensure that the three separate compressor casings would provide the best possible 

vehicle for tip clearance performance comparisons, the casing designs incorporated tight 

manufacturing tolerances. In addition to preventing the introduction of tip clearance nonuniformities for a given casing configuration, these tolerances also helped to ensure consistency 

between the three casings. The most important of these tolerances pertained to the inner diameter 

of the casing over the rotor tips, for which the diameter dimension was specified with a tolerance 

of +0.001/0.000 in., as well as geometric tolerances of 0.005 in. for the overall runout and 0.002 

in. for the profile shape. These tolerances, as well as all of the other pertinent dimensions, were 

verified by coordinate measuring machine (CMM) inspection processes. 

2.2. Steady Compressor Performance 

2.2.1. Corrected Operating Conditions for Humidity Effects 

The definition of a compressor operating point is typically “corrected” to standard day 

conditions to ensure consistent measurements with changing inlet conditions and comparability 

with computational simulations. To this end, many facilities operate according to classical 

corrected speed and corrected mass flow rate definitions defined by turbomachinery textbooks. 

However, careful examination of these equations shows that they do not appropriate account for 

the presence of humidity in the air. 

Existing compressor test codes and full engine analyses discuss the importance of humidity, 

and some suggest the use of conditioned or refrigerated air to avoid the potentially negative 

effects (NACA TN-1138, 1950; Erwin, 1964). Because the Purdue three-stage compressor 

facility does not have the luxury of a conditioned air supply, a method of appropriately 

accounting for the presence of humidity is required. 

Analysis of the derivation process leading to the classical equations defining corrected 

speed and corrected mass flow rate for a compressor highlights the assumptions of a thermallyand calorically-perfect gas which are not necessarily appropriate for a humid air mixture. 

Ultimately, an alternate presentation of these same equations has been derived to circumvent the 

need for perfect gas assumptions. These alternate equations utilize stagnation speed of sound and 

stagnation density in place of stagnation temperature and stagnation pressure in the classical 

equations. By this process, the corrected rotational speed is determined by 

NASA/CR—2015-218868 11
. =

mech

,3,act⁄,3,ref

, (2.1)

and the corrected mass flow rate is determined by 


 . = 
 act 7

"8,ref ,3,ref

"8,act ,3,act9 . (2.2)

A formal derivation of these conditions for humid air considerations is available in 

Berdanier et al. (2015). The results of this study showed that air conditions which are realizable 

in the hot and humid midwest United States summers can lead to a miscalculation of required 

mechanical speed on the order of 0.5% and actual mass flow rate on the order of 1.0%. Even for 

the intermediate-speed class of compressors encompassing the Purdue three-stage compressor, 

these effects are well within the measurement capabilities of existing instrumentation. 

Specifically, a comparison of the results using the classical equations for corrected conditions 

and the alternate representations of the equations is shown by the humidity zone identified in 

Figure 2.5. 

Figure 2.5: Humidity effects on measured compressor performance for a hot humid day in the 

context of the 100% corrected speedlines for the three tip clearance heights presented in this 

study. 

Figure 2.5 reveals that research facilities which operate with unconditioned ambient air 

(such as the Purdue three-stage compressor) must carefully consider humidity effects for 

correcting operating conditions, particularly when measuring small-scale performance changes 

or attempting to compare results at the same operating condition across several seasons of the 

year. The desired deliverables of this project match both of these requirements, and therefore, 

further emphasize the importance of humid air considerations for this project. 

0.7 0.8 0.9 1.0 1.1 1.2 

1.16 

1.20 

1.24 

1.28 

1.32 

1.36 

Normalized Inlet Corrected Mass Flow Rate, mc[−]

Total Pressure Ratio [−]

TC1

TC2

TC3

Humidity

NASA/CR—2015-218868 12
2.2.2. Holistic Performance Measurements 

For the Purdue three-stage compressor, the conditions used to calculate the corrected speed 

and corrected mass flow rate are measured at the Aerodynamic Interface Plane (AIP), denoted by 

axial position zero in the flowpath shown in Figure 2.6. At this upstream inlet plane of the 

compressor, the 50% span measurements from a seven-element total pressure rake and a sevenelement total temperature rake are used to represent the bulk one-dimensional flow. The amount 

of moisture in the air is determined by an Omega HX94 thermo-hygrometer located near the inlet 

of the compressor, capable of measuring relative humidity with an uncertainty of ±2%. These 

measured parameters are used as inputs with the REFPROP thermodynamic equation program 

(Lemmon et al., 2013) to calculate the pertinent stagnation speed of sound and stagnation density 

required for Equations (2.1) and (2.2) above. 

Figure 2.6: Purdue three-stage compressor facility flowpath cross section. 

In addition to the seven-element total pressure and total temperature rakes installed at the 

AIP, 13 circumferentially-distributed static pressure taps evaluate the static pressure for this 

upstream inlet plane at the outer diameter of the flowpath. Great care has been incorporated to 

ensure circumferential uniformity of the pressures and temperatures at the inlet of the 

compressor, as measured by Ball (2013). 

Referring again to the flowpath in Figure 2.6, the exit plane of the compressor (axial 

position 9) is also defined by a set of seven-element total pressure and total temperature rakes 

and one static pressure measurement. Separate from these upstream and downstream locations, 

additional seven-element total pressure and total temperature rakes are available at each of the 

axial positions labeled 1 through 8. Static pressures at these axial positions are evaluated by four 

circumferentially-distributed static pressure taps on the casing. The combination of these 

measurements at positions 1 through 9 provide the opportunity to define overall compressor 

performance and relative performance of each individual blade row or each stage. The radial 

position of the measurement locations for the seven-element pressure and temperature rakes at 

each axial position 0 through 9 are provided in Table 2.3. 

Table 2.3: Radial distributions of rake measurement locations. 

Ax. Pos. Total Pressure [% annulus height] Total Temperature [% annulus height] 

0 12 20 30 40 60 80 88 12 20 35 50 70 80 88 

1 – 8 12 20 35 50 70 80 88 12 20 35 50 70 80 88 

9 12 20 35 50 65 80 88 12 20 30 50 70 80 88 

NASA/CR—2015-218868 13
The axial location of the stations labeled in Figure 2.6 is halfway between the blade rows 

based on the mid-span geometry. Figure 2.7 provides the details of the Kiel head geometries for 

the total temperature and total pressure rakes. The centerline of the rake coincides with the 

location of the axial measurement plane. The recovery factor for both the total pressure and total 

temperature rakes (as provided by the manufacturer) is equal to one, with three significant digits, 

for an angular acceptance range of ±30 degrees. The rakes are aligned with a representative mean 

flow angle, but this angular acceptance facilitates accurate measurements at off-design operating 

conditions. 

Figure 2.7: Rake geometry including Kiel head details. 

Because the recessed casing geometry design (Figure 2.2) extends to the farthest axial 

extent, when the rakes are inserted with the TC2 or TC3 casing hardware, they are positioned 

within the tip clearance recess area. For this study, the rakes were radially positioned such that 

the measurement positions defined in Table 2.3 are consistent with the annulus height of the 

stator vanes (which are geometrically unaffected by the casing recesses). These rake positions 

are shown graphically in Figure 2.8, such that the outer edge of the rake was aligned with the 100% 

annulus height location, regardless of the tip clearance configuration (i.e., whether or not a 

casing recess was present). By this method, the detailed measurements collected at these same 

axial positions extend to 101.5% or 102.5% annulus height for the TC2 and TC3 tip clearance 

configurations, respectively. 

NASA/CR—2015-218868 14
Figure 2.8: Rake positions for different tip clearance configurations. 

The new variable tip clearance hardware for this project expanded the capabilities of the 

facility to accommodate measurement rake insertion at three or five circumferential positions 

(specifically for measurement plane 6, at the exit of the embedded second stage) for each of the 

positions 1 through 8 in Figure 2.6. These same rake insertion positions also function as 

locations for the installation of L.C. Smith precision probe positioners for detailed measurement 

traverses with a variety of instrumentation technologies. 

In addition to the flow measurements, a series of surface-mounted thermocouples are 

installed on the outer part of the casing over each blade row to yield an axial temperature 

distribution. These measurements facilitate CFD analyses implementing alternate boundary 

condition methods, such as the isothermal boundary condition presented by Bruna and Turner 

(2013) for Rotor 37, instead of the more common, but significantly simplified, adiabatic 

boundary condition. Because the data collected from this experimental study are expected to aid 

in the development and calibration of computational models, it is important that these 

temperatures are collected with the aerodynamic performance data to appropriately define the 

system boundary conditions. 

The total and static pressures for the facility are measured using Scanivalve 3217 and 3016 

Digital Sensor Array (DSA) pressure scanner modules, each containing 16 temperaturecompensated piezoresistive pressure sensors. Depending on the requirements of the specific 

measurement, the DSA channels are rated for differential pressure ranges of 1 psid, 2.5 psid, 5 

psid, or 15 psid. These devices have been selected for their long-term accuracy ratings of 0.12% 

FS, 0.12% FS, 0.05% FS, and 0.05%FS, respectively. Thermal measurement drifts from these 

devices are prevented through regular procedures for an on-line zero calibration prior to and 

during compressor operation. The differential pressures are referred to an atmospheric reference 

pressure which is separately measured using a high-accuracy Setra 270 barometric pressure 

sensor with an accuracy of 0.03% FS. 

The temperatures for the facility are evaluated using T-type thermocouples paired with 

special limits of error (SLE) extension wire. A Keysight Technologies 34980A integrating 

voltmeter paired with 34921A/T modules and terminal blocks accommodates the measurement 

NASA/CR—2015-218868 15
of these thermocouple channels. To achieve significantly reduced uncertainty contributions to the 

temperature measurements, Keysight Y1138A extension cables separate the terminal blocks 

from the measurement modules, and a 10 kthermistor with an interchangeability of ±0.18 oR 

provides the reference temperature for the isothermal terminal block. The temperature data 

reported for this study have been evaluated by an arithmetic mean of five independent 

measurements, each of which represents a thermocouple voltage measurement integrated over 

five power line cycles. 

Through the use of a liquid calibration bath, the measurement chain for each of the 

temperature rake measurement channels was evaluated. By incorporating a high-accuracy 

thermistor as a reference measurement in the same calibration bath, the overall uncertainty of 

each individual channel was determined. These calibration results show an overall uncertainty 

typically less than 0.45 oR for all temperature channels. However, the repeatability of the 

temperature measurements is an order of magnitude less than this value. 

2.2.3. Isentropic Compressor Efficiency 

For this study, the isentropic compressor efficiency will be evaluated using the formal 

definition, 

 =

h3;< − h3=

h3; − h3=

, (2.3)

as a function of stagnation enthalpies. Alternatively, a perfect gas assumption allows a 

simplification of isentropic efficiency, 

 =

TPR>>?@=A⁄?A − 1

TTR − 1

,

(2.4)

which appears to be preferable due to its definition as a function of measured properties – total 

pressures and temperatures appearing in the total pressure ratio (TPR) and total temperature ratio 

(TTR). However, following an effort to minimize perfect gas simplifications, there is an inherent 

complication with the use of Equation (2.4), as discussed by Lou et al. (2013) and continued by 

Berdanier et al. (2015). Specifically, Lou et al. (2013) showed that the isentropic efficiency is 

always over-predicted when using Equation (2.4). This over-prediction may be less than 1 

efficiency point for an overall TPR of 1.3, but it increases to over 5 efficiency points for a TPR 

on the order of 40 or more. The introduction of REFPROP as an integral component of the data 

processing suite for this study provides the opportunity to calculate isentropic efficiency via the 

definition of the parameter shown in Equation (2.3). As with the calculation of corrected 

parameters described above, the use of REFPROP also accommodates the inclusion of humidity 

in the calculation of the stagnation enthalpies required to evaluate efficiency.

2.2.4. Detailed Steady Total Pressure Traverses 

As described above, the steady compressor pressure rise is defined using a series of sevenelement total pressure rakes distributed throughout the machine. However, these measurements 

inherently lack spatial resolution and a more thorough distribution of measurements closer to the 

endwalls. Thus, in addition to the total pressure rake measurements, a miniature Kiel head total 

pressure probe (United Sensor KAC-type) was traversed in small radial increments downstream 

of each of the stator vane rows. This probe features a 0.063 in. outer diameter Kiel head and was 

NASA/CR—2015-218868 16
paired with a L.C. Smith precision probe positioner to incrementally achieve the desired plunge 

locations with position feedback. The probe accommodates a yaw and pitch acceptance cone of 

greater than ±40 degrees, but the probe was aligned with a representative mean flow angle. 

2.3. Flow Visualization Techniques 

In addition to the detailed total pressure traverses downstream of the stators, a flow 

visualization method was introduced to investigate the surface flow topologies on the stator 

vanes and evaluate regions of flow separation. The flow visualization technique was performed 

using a mixture of kerosene and powder paint. The mixture is inserted at the AIP (axial position 

0), as shown in Figure 2.9, while the compressor is running at a steady operating condition. The 

mixture flows through the compressor, and the kerosene evaporates leaving streaklines of paint 

on the flowpath surfaces, particularly in regions of flow recirculation. Following a series of 

preliminary tests, a mixture ratio of 5:2 kerosene-to-paint (based on volume) was selected to 

provide optimal paint coverage. The position of paint injection in Figure 2.9 was chosen 

specifically to allow full coverage of each individual vane row. Two colors of fluorescing 

powder paint were used, blue and orange, to provide high-contrast images. 

All flow visualization photographs were acquired with a Canon Rebel XSi DSLR camera 

with an EF-S 18-55mm f/3.5-5.6 SI lens. Because the paint is deposited in regions of 

recirculating flow, most of the paint is located near the trailing edge of the vanes. Thus, all 

pictures were taken aft of the vane, looking forward (unless otherwise noted). The photographs 

presented herein were all illuminated with a 100 Watt ultraviolet black light lamp, allowing the 

fluorescent power paint to be more clearly identified. Additional information about the setup 

procedure and photographic processing methods is available in Smith and Key (2015). 

Figure 2.9: Schematic of flow visualization setup, forward looking aft. 

NASA/CR—2015-218868 17
2.4. Tip Clearance Measurement System 

It is well known that static rotor tip clearances – sometimes referred to as “cold” 

clearances – do not typically represent the operating, or “hot” running, clearances. In general, 

thermal growth, pressure forces, and centrifugal effects due to high-speed rotor rotation 

contribute to blade growth and untwist which leads to difference rotor tip clearances at different 

operating conditions and for different ambient conditions. The influence of thermal effects on the 

compressor casing may also contribute significantly to changing rotor tip clearances. As a result, 

a series of measurements have been collected to evaluate the static and operating tip clearances 

of the Purdue three-stage compressor. 

2.4.1. Static Tip Clearance Measurements 

The static “cold” tip clearances have been evaluated for each of the three tip clearance 

configurations. These measurements were collected by measuring the blade-to-blade variability 

(i.e., the run-out of the rotor blade tips) using a fixed dial indicator with a rolling tip. The 

manufacturing tolerances of the rotor blisk show the blade-to-blade variability at any fixed 

location about the circumference of the compressor, as presented in Figure 2.10. These blade-toblade measurements were collected several times to calculate the mean values in Figure 2.10. As 

a result, the uncertainty of the mean value is on the order of 1×10-4 in. 

The concentricity of the rotor with respect to the casing has also been evaluated using static 

tip clearance measurements collected at several positions around the circumference of the 

compressor casing. These measurements have shown for each tip clearance configurations that 

the concentricity of the rotor with respect to the casing is on the order of 0.004-0.008 in. The 

smallest tip clearances are measured at a position of 345 degrees from the top of the compressor, 

in the direction of rotor rotation, as shown in Figure 2.11. The rotor concentricity can be 

controlled using a set of alignment bolts at both the front and rear bearing assemblies, as 

prescribed by Talaleyev (2011). This concentricity control is limited to within 0.008-0.012 in. 

based on bolt thread pitch. Thus, the measured concentricity of the rotor is confirmed to be 

within the adjustment capabilities of the system. 

NASA/CR—2015-218868 18
Figure 2.10: Blade-to-blade variability of tip clearance height with respect to “Blade 1” for all 

three rotor rows. 

2.4.2. Operating Tip Clearance Instrumentation 

For this study, a facility upgrade included the introduction of a CapaciSense 5-series 

frequency modulated (FM) capacitance probe tip clearance measurement system (produced by 

Pentair Thermal Management). The system features nine available channels, allowing three 

probes to be implemented for each of the three compressor rotors. The three probes are equallyspaced circumferentially, 120 degrees apart, at positions of 105 degrees, 225 degrees, and 345 

degrees – all measured from the top of the compressor in the direction of rotor rotation (see 

Figure 2.11). Each of the nine probes was individually calibrated using a custom-designed 

calibration disk which represents the tip geometry of the rotor blade through scaling techniques 

proven by the manufacturer. The probes were all calibrated for operation from a rub condition 

(0 in. clearance) to a maximum 0.2 in. clearance height. 

Figure 2.11: Circumferential capacitance probe measurement locations. 

1 12 24 36

−3

−2

−1

0

1

2

3

4

x 10−3

Blade−to−Blade Tip Clearance [in]

Rotor 1

1 11 22 33

−3

−2

−1

0

1

2

3

4

x 10−3

Blade Number

Rotor 2

1 10 20 30

−3

−2

−1

0

1

2

3

4

x 10−3 Rotor 3

NASA/CR—2015-218868 19
The electronics chain for the FM tip clearance measurement system has several key 

components. The probes, shown in Figure 2.12(a), were designed and built for application in this 

specific facility; a triaxial cable is permanently attached to the probe and provides enhanced 

rejection of electromagnetic interference. The use of a non-mineral-insulated cable limits the 

maximum operating temperature of this system to approximately 260 oC, well above the 

operating limits of the Purdue three-stage compressor. The oscillator, shown in Figure 2.12(b), 

drives the cable with an oscillating voltage (nominally 2 MHz). As the blade passes the probe, 

the measured capacitance modulates the driven frequency from the oscillator. This modulation is 

sensed by the carrier, and the demodulator (Figure 2.12(c)) converts the modulation frequency 

due to the blade passing event to a DC voltage. The DC voltage is correlated to a clearance 

height via the individual channel calibration. This proportionality between measured capacitance, 

frequency modulation, voltage, and tip clearance is the heart of the FM tip clearance 

measurement system. 

(a) (b) (c) 

Figure 2.12: Capacitance tip clearance measurement system components. 

(a) Probe; (b) Oscillator; (c) Demodulator. 

The clearance measurement system is controlled from a set of “Control and Processing 

Module” (CPM) computers operating Slackware 14 Linux. Each of these computers utilizes one 

Advantech PCI-1714UL data acquisition card to sample four of the nine channels (three total 

CPMs). The three CPMs are setup in a master-slave-slave configuration communicating via 

Ethernet, and the master CPM utilizes a proprietary graphical user interface (GUI) to configure 

and operate the entire probe measurement chain (including the oscillators and demodulators). 

The data acquisition card is capable of sampling the four analog channels at up to 10 MHz per 

channel, as well as an external trigger which is linked to the 1/rev signal from the laser 

tachometer for the compressor. For the measurements presented in this study, all data were 

collected using the full 10 MHz sampling capability of the data acquisition card. Other 

specifications for the capacitance probe measurement system are listed in Table 2.4. 

The clearance calculation process provides a blade-by-blade clearance output calculated 

from the peak-to-peak voltage for the typical blade pulse output signal (known as the Blade 

Passing Signal (BPS)). By synchronizing the measured clearances with the once-per-revolution 

trigger from the compressor, these calculated clearances from the BPS output can be traced to 

known blade numbers. As an alternative to this BPS method, the software applies a low-pass 

filter to the BPS output to create a DC voltage output signal which the manufacturer refers to as a 

“RMS” signal, although it should be noted that the low-pass filter mechanism does not represent 

a true root-mean-square calculation procedure. This RMS signal serves as a representative 

average measurement of the tip clearance for all blades by a particular probe. 

NASA/CR—2015-218868 20
Table 2.4: Capacitance probe measurements system specifications. 

Parameter Value 

Operating Frequency 2 MHz nominal 

Oscillator Sensitivity 100 kHz per pF 

Demodulator Sensitivity 500 mV per kHz typical 

Measurement Range 0.200 in. calibrated 

System Resolution < 4×10-5 in. (at 0.030 in. clearance) 

Signal-to-Noise Ratio 30-50 typical 

2.4.3. Operating Tip Clearance Measurement Uncertainty 

The manufacturer of the capacitance probe measurement system claims an umbrella 

uncertainty on the measurement system of less than 4×10-4 in. This value is based on historical 

comparisons with other measurement techniques, including laser measurements and rub sticks, 

as well as careful attention to the design, manufacture, and calibration processes to ensure 

minimal uncertainties. However, no formal uncertainty analysis had yet been performed to 

validate this claim, which represents more of a repeatability, or comparability, and does not 

consider uncertainty contributions from the electronics components. 

To overcome this limitation, an uncertainty analysis was performed based on known 

information about the electronic devices in the system and the calibration data. Following a 

method outlined by Müller et al. (1997) for a similar capacitance probe system, the uncertainty 

of the individual probes was evaluated to be on the order of 0.001 in. or less over the range of tip 

clearances investigated for this study. Additional information regarding the tip clearance system 

uncertainty analysis is available in Berdanier and Key (2015a). 

2.5. Time-Resolved Flow Field Measurements 

2.5.1. Over-Rotor Static Pressure Measurements 

Time-resolved static pressures over the rotor tips were measured using a custom-designed 

array implementing flush-mounted fast-response pressure transducers. This array incorporates 25 

Kulite XCS-062 subminiature pressure transducers (having an outer diameter of 0.066 in.) with a 

5 psig range. These XCS-series transducers feature high-sensitivity piezoresistive sensing 

elements, maximizing the resolution of the measurements. These sensors were installed with 

standard B-type screens for protection of the sensing elements, but the new sensors were filled 

with additional silicon material to reduce the cavity size behind the screen and allow for the 

greatest allowable frequency response. As a result, a natural frequency on the order of 30 kHz is 

expected for these new sensors. 

During the design process, a specific effort was made to accommodate all sensors in one 

axial row. Some previous authors have utilized a method of offsetting sensors in two or more 

NASA/CR—2015-218868 21
axial rows, separated by some angle in the pitch-wise direction, to accommodate more axial 

resolution (Yoon et al., 2006; Shin et al., 2008; Courtiade and Ottavy, 2012). However, such 

sensor orientations can introduce complications in the measurements for multistage machines, as 

the wakes propagating from upstream stators and potential fields from downstream stators may 

affect the measurements. Understandably, this effect can be accounted for if the sensors are 

moved relative to the stator vanes, but it introduces an additional complication to the data 

acquisition and processing procedures. As an alternative, other authors have achieved maximum 

axial resolution by implementing a series of sensors at different axial positions, spaced 

periodically in the circumferential direction with respect to the rotor pitch (Levis, 2006; Sans et 

al., 2013). However, this method could be also be negatively affected by blade-to-blade nonuniformities, rotor eccentricity, and any associated flow variability. 

The small diameter of the XCS-062 devices provided the ability to incorporate as many 

sensors as possible in one axial row, thereby increasing the spatial resolution of the output results. 

The sensors are permanently installed in a removable block which can be inserted into any one of 

nine frames. These nine frames represent one for each of the three rotor blade rows (R1-R3), and 

each of the three tip clearance configurations (TC1-TC3). The removable sensing block and one 

of the nine frames are shown in Figure 2.13(a). 

The instrumentation block was designed to maximize the number of sensors which can fit 

within the space defined by the instrumentation access location over Rotor 1, which features the 

least axial space. Because the same fixed sensor locations are implemented for Rotor 2 and Rotor 

3, whereas the axial chord of the rotors increases slightly moving from Rotor 1 to Rotor 3, the 

positions of the sensors change with respect to the leading and trailing edges of the blades for the 

three rotor rows. These sensor positions are presented for the three rotors in Table 1 as a 

percentage of axial chord. 

Table 2.5: Flush-mounted sensor positions for each rotor as a percentage of axial chord. 

Blade Row Minimum Sensor

Location [%B] 

Maximum Sensor

Location [%B] 

Sensor Separation

[%B] 

Rotor 1 -14.0 114.0 5.33 

Rotor 2 -12.8 113.6 5.26 

Rotor 3 -11.9 113.2 5.21 

The excitation and amplification for these fast-response pressure sensors were provided by 

a Precision Filters 28000 chassis with four 28118 full bridge amplification cards, each of which 

is capable of managing eight channels for a total of 32 simultaneous operating sensors. Each of 

the channels in the Precision Filters system is independently managed through a manufacturerdesigned GUI to adjust the DC offset, gain, and on-board analog filter characteristics. The sensor 

signals were then digitized using an NI PXIe-1073 chassis utilizing two 16-channel PXIe-6358 

modules, each capable of simultaneous sampling of up to 1.25 MHz per channel, via NI BNC2110 connector blocks. 

Immediately prior to installation and operation in the compressor, the fast-response 

pressure sensors were calibrated in a custom-designed calibration chamber (Figure 2.13(b)) by 

prescribing a known pressure to the device. For each calibration, up to 15 data points were 

prescribed over the full sensing range of 0 to 5 psi. From these measurements, a least-squares 

NASA/CR—2015-218868 22
linear fit was applied to the calibration data. The channels were zeroed prior to each calibration 

and again prior to testing, although thermal compensation and a stable measurement system 

reduce the day-to-day drift of a representative channel to 1 mV or less over (after amplification), 

which corresponds to a pressure of less than 8×10-4 psi. 

For each rotor and each tip clearance configuration, data were acquired at several loading 

conditions on the 100% corrected speedline. In all cases, 500 revolutions of data were 

simultaneously sampled to allow phase-locked ensemble averaging and other pertinent statistics. 

(a) (b) 

Figure 2.13: Over-rotor static pressure measurement system. (a) Removable sensing block and 

frame; (b) Calibration chamber. 

2.5.2. Total Pressure Measurements at Rotor Exit 

In addition to the steady total pressures measured at positions downstream of the stator 

vane rows, time-resolved total pressure measurements were also acquired at the rotor exit planes. 

A custom-designed fast-response total pressure probe incorporates a Kulite LQ-062 subminiature 

pressure transducer embedded in a miniature Kiel head with a 0.083 in. outer diameter and a 

0.067 in. inner diameter (Figure 2.14). The sensor has a 5 psi maximum range with an 

electronically sealed gage reference condition. As with the surface-mounted pressure sensors 

described in Section 2.5.1 above, this transducer also includes a standard B-type protective 

screen, therefore reducing the natural frequency of the sensors to approximately 20 kHz. A 

separate pressure chamber designed for this probe provided the ability to calibrate the sensor at 

prescribed intervals over the 0-5 psi, range and a least-squares linear fit was applied to the 

calibration results. The sensor was balanced with the acquisition of a zero gage pressure reading 

prior to each calibration, and similarly before and after each test. As with the detailed steady total 

pressure measurements, these time-resolved rotor exit total pressures were acquired at two 

loading conditions on the 100% corrected speedline for the smallest and largest tip clearance 

configurations (TC1 and TC3). 

NASA/CR—2015-218868 23
Figure 2.14: Fast-response total pressure probe. 

2.5.3. Thermal Anemometry 

Hot-wire anemometry was used for this study to acquire time-resolved velocity and flow 

angle measurements throughout the compressor. For all measurements, a Dantec Dynamics 

StreamLine Pro frame was used with 91C10 constant temperature anemometer (CTA) bridges. A 

computer interfaces with the bridge frame through a PC operating the Dantec Dynamics 

StreamWare software. The sensor voltages output from the CTA bridges are digitized using an 

NI PXIe-1073 chassis through a NI BNC-2110 connector block and an eight-channel NI PXIe6356 module. A custom-designed NI LabVIEW GUI operates the L.C. Smith precision probe 

positioner holding the sensor, as well as the data acquisition procedures. 

The hot-wire sensors utilized for this project were calibrated using the hot-wire calibration 

facility in the Purdue Compressor Research Laboratory, which features a low-turbulence system 

with three-dimensional directional sensitivity and precision control of flow conditions. The 

facility, shown schematically in Figure 2.15, is fed from a reservoir of dry compressed air 

maintained at a nominal pressure of 300 psig. A digital pressure regulator (UP) adjusts the 

upstream pressure of the system to control the velocity via a proportional-integral-derivative 

(PID) control loop. The flow passes through two in-line flow heaters (H1, H2), and a separate 

PID control sequence maintains the jet temperature to within one degree Fahrenheit of the 

desired set point. The heaters ensure that the hot-wire calibration data are acquired at a constant 

temperature to remove calibration uncertainties due to temperature changes. After flowing 

through the two heaters, the air is diverted around a flow spreader and then passes through a 

series of honeycomb and screens for flow conditioning. A nozzle with a diameter of 1 in. 

accelerates the flow to the desired velocity, which is determined by a combination of stagnation 

plenum pressure and static pressure at the nozzle exit. The system maintains the jet exit velocity 

within 0.2 percent of the desired set point. An attachment to the calibration nozzle (Figure 2.16) 

provides the capability to vary the position of the probe with respect to the jet, effectively 

changing the pitch angle of the flow with respect to the probe. Additional information about the 

calibration jet facility is available from Morrison (2013). 

NASA/CR—2015-218868 24
Figure 2.15: Purdue Compressor Research Laboratory hot-wire calibration facility schematic 

(Morrison, 2013). 

Figure 2.16: Hot-wire cailibration facility optional three-dimensional calibration feature. 

Adapted from Morrison (2013). 

2.5.3.1. Inlet Turbulence Intensity 

The inlet turbulence intensity of the compressor was evaluated by traversing a Dantec 

Dynamics 55P12 miniature 45 degree slanting wire sensor radially across the flow field at the 

AIP (plane 0 in Figure 2.6). For these measurements, the sensor was positioned in a normal wire 

orientation perpendicular to the mean flow in a benign one-dimensional flow region where the 

slanting wire would not be affected by three-dimensional flows. For these data, the wire bridge 

was operated in a 1:1 mode to achieve increased frequency response with the CTA system on the 

order of 200 kHz. In the CTA control software, the overheat ratio (OHR) of the bridge was 

adjusted to the manufacturer-optimized value of 1.8, as defined the resistance of the wire, F, 

under heated (“hot”) and ambient (“cold”) conditions:

Po, To

Ps

H1

H2

UP

PRESSURE

SUPPLY

Flow

Spreader Honeycomb Mesh

Screens

S

NASA/CR—2015-218868 25
OHR =

F,hot

F,cold

. (2.5)

With sufficient frequency response from the hot-wire, the turbulence intensity can be 

measured as a scale by which to evaluate the amount of turbulent fluctuations in the flow. The 

turbulence intensity (TI) is evaluated as the ratio of the root-mean-square (RMS) of the velocity 

fluctuations to the mean velocity: 

TI =

N′

P

, (2.6)

where the velocity can be defined at any position as the sum of the average, P, and the random 

fluctuations, N

Q

 (RMS): 

 = P + N

Q

. (2.7)

The inlet turbulence intensity, Figure 2.17, has been evaluated for 16 radial measurement 

positions across the annulus height at the AIP. These results are presented for a loading condition 

near the peak efficiency point on the 100% corrected speedline. This figure shows the inlet TI is 

on the order of 1% up to approximately 75% annulus height, where it begins to increase toward 

the tip. This increase is likely associated to the increased fluctuations in the boundary layer, but 

the sharp increase close to the wall may be due to an influence from the hole in the outer casing 

through which the probe is inserted. These same data were collected at ten loading conditions 

from open throttle to a near stall operating condition, but the standard deviation of TI at each 

radial position is small (less than 0.05%) across a majority of the speedline, signifying an 

insensitivity of inlet TI with respect to loading condition. The largest changes are observed 

approaching the stall point on the 100% corrected speedline, for which the values shown in 

Figure 2.17 increase by approximately 0.20% across the entire annulus height. 

Figure 2.17: Radial profile of inlet turbulence intensity, as measured at the AIP. 

0 2 4 6 8 10 12 14 16

0

10

20

30

40

50

60

70

80

90

100

Turbulence Intensity [%]

Annulus Height [%]

NASA/CR—2015-218868 26
2.5.3.2. Velocity and Flow Angle Measurements 

The velocity and flow angles throughout the compressor were evaluated for axial 

measurement planes 2 through 8. These data were acquired at two loading conditions on the 100% 

corrected speedline for the two extreme tip clearance configurations (TC1 and TC3). These tests 

were conducted using a Dantec Dynamics 55R02 fiber-film sensor with a 45 degree slanting 

angle. For this sensor, the manufacturer specifies a sensing film length of approximately 0.05 in. 

and a nickel film diameter of approximately 5×10-5 in. The selection of the slanted probe for this 

project provides the opportunity to investigate three-dimensional flow fields, with a particular 

emphasis on the tip leakage flow features. For these measurements, the 55R02 sensor was 

operated in a standard 20:1 CTA bridge mode with a manufacturer-optimized overheat ratio of 

1.8, as previously defined by Equation (2.5). 

The fiber-film sensor technology is offered as an option which provides a more robust 

design than traditional thin wires, thereby reducing susceptibility to damage from the high-shear 

flows present in high-speed compressor applications, without the degradation of frequency 

response which is typically observed with traditional film sensors. Throughout this document, 

however, the term ‘wire’ will be used interchangeably to also describe the fiber-film sensor. 

The 55R02 sensor was calibrated over a range of velocities and flow angles, as prescribed 

in Table 2.6. The response of the fiber-film sensor was optimized near the upper end of the 

velocity calibration range (400 ft/s) using a square wave test signal. Using this method, the 

frequency response of the sensor was estimated to be approximately 85 kHz – nearly 30 times 

the blade passing frequency of the rotor blades in the desired compressor application. For each 

test campaign, a calibration was performed prior to the data collection, and a separate calibration 

was performed after the completion of the campaign for comparison of measurement drift. The 

jet temperature for these calibrations was always maintained at a constant 100 oF. 

Table 2.6: Hot-wire calibration matrix. 

Parameter Velocity, [ft/s] Yaw angle, [deg] Pitch angle, [deg]

Min. value 100 -90 -20 

Max. value 450 +90 +20 

Increment 50 5 5 

To evaluate three-dimensional phase-locked ensemble averaged velocity components, three 

separate wire positions are required with respect to the flow field. At each test condition and 

measurement location, the sensor was aligned with the approximate mean flow angle, and then 

yawed to the three desired measurement angles using the precision probe positioner. Following 

the procedure outlined by Shin and Hu (1986), the sensor angles used for this study were -60 

degrees, -20 degrees, and +20 degrees with respect to the null position, as defined by the 

coordinate system in Figure 2.18. 

NASA/CR—2015-218868 27
Figure 2.18: Angular definitions for slanted hot-wire sensor. The long prong is denoted by the 

large circle. 

The measured wire voltages were converted to known velocities and flow angles using 

calibration data with a non-linear solver technique for data reduction. In this case, the Joule 

heating principle and King’s law (1914) are applied to evaluate the voltage-velocity relationship. 

Next, the relationship between Nusselt number and Reynolds number proposed by Collis and 

Williams (1959) accounts for variations of flow properties between calibration and test 

conditions. When required, REFPROP was used to calculate thermodynamic properties for use in 

the non-dimensional Nusselt and Reynolds numbers. The effective Reynolds number, Reeff, is 

determined from the effective velocity, defined as any velocity which does not appear as a 

normal velocity with respect to the wire orientation (i.e., ±90 degrees yaw and 0 degrees pitch). 

Finally, the desired non-linear relationship: 

Reeff,S

Re

= TU + T=  + T;  + TV Re + TW   + T1  Re + TX  Re

+ TY 

; + TZ ; + T[ Re;

,

(2.8)

is developed for ten coefficients T\ determined from the calibration data, which must apply for 

all three wire orientations, ] = 1. .3. This equation is based on the method utilized by Schmidt 

and Okiishi (1976). 

An example comparison of the non-linear equation fit from Equation (2.8) with the discrete 

calibration points, Figure 2.19, represents this functional relationship for one of the calibration 

sets used in this study at a velocity of 300 ft/s. For this study, the slanted wire was found to be 

asymmetric about the null (zero degrees yaw) position. As a result, a separate coefficient set was 

defined for each half of the curves in Figure 2.19. The coefficient fit for the example data in 

Figure 2.19 has an R2 value greater than 0.98, where R2 represents the coefficient of 

determination as calculated from the residual sum of squares and the total sum of squares. 

Once wire voltages are measured in the compressor, the effective Reynolds numbers are 

calculated, and the three equations, Equation (2.8), are solved simultaneously using the predetermined coefficients to calculate the true Reynolds number and the flow angles, which are 

geometrically related to the pre-defined wire positions of -60, -20, and +20 degrees. 

#

+





+





NASA/CR—2015-218868 28
Figure 2.19: Example slanted hot-wire coefficient fit at 300 ft/s. 

2.5.3.3. Thermal Anemometry Uncertainty 

The uncertainty of the calibration data collected from the Purdue calibration jet facility has 

been evaluated previously by Morrison (2013). To evaluate the accuracy of the processing 

algorithm implemented for the velocity and flow angles measurements, a series of data were 

collected at known wire positions, velocities, and conditions in the calibration jet which 

represent the compressor operating conditions for each test campaign. Specifically, the 

temperature and density of the jet were adjusted to match the flow conditions determined from 

the steady total temperature and total pressure rakes (and wall static pressure), as measured 

simultaneously with the hot-wire. Furthermore, a representative velocity was also calculated 

from these rake measurements to select the operating velocity of the calibration jet. Using this 

method, the accuracy of the measurements at each axial measurement position (i.e., Rotor 1 exit, 

Stator 3 exit, etc.) and compressor loading condition was separately identified, Figure 2.20. 

The results in Figure 2.20 show that the calculated velocity from the hot-wire processing 

algorithm is a strong function of the operating condition in the compressor, whereas the 

calculated yaw and pitch angles are less dependent on operating conditions. An investigation 

with alternate processing methods (Berdanier and Key, 2015b) has shown that the velocity can 

be greatly affected if the conditions of the measured flow vary significantly from the calibration 

conditions. However, the calculated flow angles are much less sensitive to these temperature 

effects due to their greater dependence on the geometry of the probe itself. Indeed, analysis of 

the results in Figure 2.20 shows that the test conditions which yield the most significant velocity 

difference are the same conditions for which the temperature at the wire location was most 

different from the calibration temperature. In the same manner, the test conditions which show 

the least difference (closest to zero) are those for which the temperature at the location of the 

wire in the compressor was nearly identical to the calibration temperature. For reference, in this 

figure, the velocity difference on the order of -60 ft/s is representative of a temperature 40 oF 

lower than calibration conditions, and the velocity difference on the order of +30 ft/s is 

representative of a temperature 30 oF higher than calibration conditions. 

−90 −60 −30 0 30 60 90

0.1

0.3

0.5

0.7

0.9

1.1

Yaw Angle, α [deg]

Reeff/Re

φ = −20o

φ = +20o

Calibration

Coeff. Half 1

Coeff. Half 2

NASA/CR—2015-218868 29
Figure 2.20: Approximate accuracy of hot-wire processing algorithm for test-representative 

conditions at several known flow angles. 

The reason for this discrepancy in velocities is related to the correction proposed by Collis 

and Williams (1959). Previous studies have shown that the Collis and Williams correction 

relating the Nusselt number and the Reynolds number deviates for Mach numbers above 

approximately 0.3. In these cases, compressibility effects begin to become more significant for 

the hot-wire measurements. Referring again to the results in Figure 2.20, the rotor exit 

measurements are the locations at which the absolute flow velocity is expected to be the largest 

(due to the diffusion of flow in the relative reference frame). Furthermore, as the temperature 

decreases, the Mach number will increase. As a result, the measurements at Rotor 1 exit for TC1 

at NL with a velocity difference on the order of -60 ft/s represent a Mach number on the order of 

0.35. All of the tested conditions in Figure 2.20 with a velocity difference beyond approximately 

±25 ft/s represent flow conditions with an average Mach number greater than or equal to 0.30. 

Although the velocity differences in Figure 2.20 appear significant, this processing method 

of solving a set of non-linear equations represents the most accurate of two methods evaluated 

for these data. The other method implemented a look-up table approach, for which the 

temperature and density variations are independently accounted using known relationships, as 

discussed by Berdanier and Key (2015b). 

Aside from developing new relationships to accommodate temperature variations with 

compressibility considerations, future hot-wire measurements such as these may benefit from a 

more controlled selection of inlet conditions which closely resemble calibration conditions, or 

several calibration sets for different temperature ranges. However, neither of these 

recommendations is particularly advantageous: (i) A judicious selection of compressor operating 

conditions comparable to calibration conditions can only be achieved with temperaturecontrolled compressor inlet conditions. This is especially true for test campaigns which span 

several days in the midwest United States, where weather patterns can be particularly difficult to 

predict. (ii) Extended calibration time poses an increased potential to experience a broken sensor. 

The fragility of hot-wire sensors makes them particularly prone to breakage, and additional 

operating time (in a calibration jet, for example) greatly increases this potential for failure. 

−10 −5 0 5 10 15 20

−80

−60

−40

−20

0

20

40

Velocity Difference, 

∆V [ft/s]

Yaw Angle [deg]

−10 −5 0 5 10 15 20

−4

−3

−2

−1

0

1

2

Yaw Difference, ∆α [deg]

Yaw Angle [deg]

−10 −5 0 5 10 15 20

−4

−2

0

2

4

6

8

Pitch Difference, ∆φ [deg]

Yaw Angle [deg]

TC1 eR1 NL

TC1 eR1 HL

TC1 eS1 NL

TC1 eS1 HL

TC1 eR2 NL

TC1 eR2 HL

TC1 eR3 NL

TC3 eR1 HL

TC3 eS1 NL

TC3 eS1 HL

TC3 eR2 NL

TC3 eR2 HL

TC3 eS2 NL

TC3 eR3 NL

NASA/CR—2015-218868 30
To further evaluate the robustness of the hot-wire calibration and processing algorithm, two 

separate calibration data sets (one before the test campaign and another after the test campaign, 

each approximately one week apart) were used to process the same experimentally-collected data. 

Specifically, data collected at the exit of Stator 1 for TC1 at two loading conditions were 

processed using the two sets of coefficients (Equation (2.8)). This axial measurement position at 

Stator 1 exit was selected because it represents a data set which was collected with 

approximately equally time between the pre-calibration an the post-calibration data (about four 

days between the collected data and either calibration). The results from this comparison, Figure 

2.21 show the differences between the results processed with the two calibrations are much 

smaller than the potential differences with respect to operating temperature and density, as 

shown in Figure 2.20. For the nominal loading (NL) condition in Figure 2.21, the velocities show 

a constant offset between the two calibrations. The yaw angles are least affected in the wake 

region, but the pitch angles show the most difference in the wake region. Separately for the high 

loading (HL) operating condition, the velocities show little effect in the wake region, with a 

more noticeable change for the undisturbed mean flow. At this same HL condition, the yaw and 

pitch angles in the mean flow are largely unchanged, but the flow angles in the wake region are 

most different. The average differences across one vane pitch from Figure 2.21 are summarized 

in Table 2.7. 

Table 2.7: Average differences across one stator pitch for data processed with two calibration 

sets. Data collected at Stator 1 exit for TC1. 

Loading ∆>/

A [%] ∆ [ft/s] ∆ [deg] ∆ [deg] 

NL 2.0 10.3 0.9 0.4 

HL 1.7 9.0 0.7 0.1 

NASA/CR—2015-218868 31
Figure 2.21: Processed hot-wire data at Stator 1 exit for TC1 using two separate calibrations

0 20 40 60 80 100

0.1 0.2 0.3 0.4 0.5 0.6 

V / Ut [−]

NL

SS PS

0 20 40 60 80 100

−20−10

0

102030

Yaw Angle, α [deg]

0 20 40 60 80 100

−10

0

102030

Pitch Angle, φ [deg]

S1 Vane Passage [%]

50% Span

70% Span80% Span

pre

post

0 20 40 60 8

0.1 0.2 0.3 0.4 0.5 0.6 

V / Ut [−]

HL

SS P

0 20 40 60 8

−20−10

0

102030

Yaw Angle, α [deg]

0 20 40 60 8

−10

0

102030

Pitch Angle, φ [deg]

S1 Vane Passage [%]

50% Span

70% Span80% Span

NASA/CR—2015-218868 32
2.6. Particle Image Velocimetry 

While traditional probe traversing measurement techniques can be used to study 

compressor performance, their intrusive design will influence the structure of the tip leakage 

flow. Thus, a non-intrusive method of measuring the tip leakage flow is desirable. To resolve the 

flow inside the rotor blade passage non-invasive measurement techniques like Laser Doppler 

Velocimetry (LDV) has been used. Ma (2001) performed 3-component LDV measurements in 

the axial compressor to reveal the dominant effect of tip corner flow from an inlet guide vane 

suction surface on the rotor blade passage. Michon et al. (2005) revealed the unsteady flow 

structures in a single stage compressor rotor blade passage and computed the Reynolds stress 

distributions using 3-component LDV. 

Because LDV is a pointwise measurement, it can be very time-intensive to traverse the 

flow field with an LDV system. Thus, researchers have turned to Particle Image Velocimetry 

(PIV) to capture planes of data in the flow field. Balzani et al. (2000) successfully acquired PIV 

data in a compressor rotor stage. Sanders et al. (2002) first performed PIV measurements in a 

transonic Compressor stator stage. Stereo-PIV measurements in the rotor blade passage were 

carried out by Wernet et al. (2001, 2005), and they identified the regions with negative radial 

velocity and low axial velocity as the tip leakage flow. Voges et al. (2012) studied the flow in the 

tip clearance region of a transonic compressor rotor with casing treatment. Stereo-PIV 

measurements were typically performed in large-scale, low speed compressors by Liu et al. 

(2003). 

In all of these studies, periscopic optical probes were introduced to the flow field for light 

sheet delivery, which renders the measurement invasive. The size of the probe affects the flow, 

but a small probe results in difficulty in precisely aligning the laser beam to avoid streaking or 

flare of the beam. Seeding can also damage the probe or require a shutdown to clean the optics. 

Thus, capability to perform three-dimensional PIV on the embedded stage of a multistage by 

introducing the laser sheet through the same window used by the cameras to acquire the image 

has been developed for this project. A window over Rotor 2 was the only optically accessible 

region through which laser sheet was introduced and also particle images were captured. This 

work demonstrates, for the first time, the capability of doing three-component, three-dimensional 

PIV in a multistage compressor, without inserting any invasive imaging or light delivery probes 

inside the compressor. This section describes the design of the setup and the issues that were 

overcome to acquire the data in this manner. The following sections describe the experimental 

set-up and different measurement conditions, PIV processing and analysis of the velocity fields 

obtained in the blade passage. 

2.6.1. Experimental Setup 

PIV was performed in the embedded stage rotor passage (Rotor 2) to measure phaseaveraged velocity field between the blade passages. All PIV measurements were performed on 

the TC3 tip clearance configuration. The experimental setup is shown in Figure 2.22. A specially 

designed glass window with a MgF2 anti-reflection coating was used to minimize reflections at 

wavelengths larger than 425nm. This window was the only optically accessible region for PIV 

measurement. The effective field of view through the window was 2.71 inches in width and 5 

inches in height. The measurement plane was located at 2.56 inches from the top inner surface of 

NASA/CR—2015-218868 33

the window frame (i.e. at 4° circumferential location), which is a radial plane passing through the 

center of the window. 
Figure 2.22: Schematic of PIV setup. 

The Quantel Evergreen series double pulse NdYAG laser (532 nm) with 211mJ energy per 

pulse per head and a max repetition rate of 15Hz was used as the illumination source. The laser 

beam was guided by two mirrors and then passed through a cylindrical lens with f=-15 mm. The 

expanding beam was passed through a 500mm plano-convex lens placed at a distance of 1.57 

NASA/CR—2015-218868 34
inches from the previous cylindrical lens. The resulting light sheet was 4mm thick in the plane of 

measurement. 

Four Imperx 2M30L CCD cameras were used to acquire the images, Figure 2.23. The 

cameras were operated with external trigger in double exposure mode with a resolution of 1600 

by 1200 pixels. The images were recorded at 8 bit depth and at 9.8 frame pairs per second. The 

desired field of view with a magnification of 27.8 um/pixel was obtained using Nikon 105mm 

AF 1:2.8D lenses. Tilt adapters manufactured by LaVision were placed in between the lens and 

the cameras to meet Scheimpflug condition and achieve better focus in the plane of measurement. 

The lens angles were 70° with respect to the laser plane normal. Typically in stereo and 

tomographic PIV, lens angles of 30° to 45° are optimal, but due to optical access limitations, any 

further reduction in lens angle was not possible. At this steep viewing angle, the image was out 

of focus from 60% span towards the hub and thus good measurements were only obtained 

between the tip and 65% span. Thus, the effective measurement domain was 40mm in the axial 

direction and 17mm in the radial direction. 

Figure 2.23: Camera, lens, and tilt adapter setup. 

The timing and synchronization for the experiment were controlled by a Quantum 

Composer Plus pulse generator (model 9518). The pulse generator was operated in an external 

trigger mode, and the individual channels were operated in a duty cycle mode. The tachometer 

signal generated from the compressor rotor shaft was a square pulse train with a frequency equal 

to the compressor rotation frequency. This signal was used to trigger the eight channels in the 

pulse generator which in turn triggered the cameras and the laser heads, Figure 2.24. Thus phaselocked measurements were obtained using the tachometer signal, and data at different 

circumferential positions were acquired by varying the delay with respect to the tachometer 

signal. The delay between the two laser pulses was set to 2 s such that particles with the highest 

velocity (which was assumed to be the blade tip velocity) would only translate across the 

measurement plane by no more than a quarter of the laser sheet thickness. 

NASA/CR—2015-218868 35
Figure 2.24: Timing diagram for synchronized image capture. 

A multi-plane LaVision target was used for calibration, Figure 2.25. The 2.28 inch square 

target plate was mounted in between the staggered blade passage and the laser sheet was aligned 

with the top surface of the target. The calibration images for the four cameras are shown in 

Figure 2.26. The compressor blades blocked different portions of the field of view for each 

camera. 

Figure 2.25: Calibration plate mounted between blade passages. 

NASA/CR—2015-218868 36
Figure 2.26: Calibration plate images from each of the four cameras. 

The laser sheet was introduced through the casing window and initially reflections from 

blade surface and hub led to saturation of image pixels, and particle image intensities were not 

discernible, Figure 2.27(a). To overcome the reflection, fluorescent dye was used with the 

seeding fluid. Lens filters blocking wavelengths below 540nm were used to filter laser 

reflections, Figure 2.27(b). The fluorescent seeding particles were generated with Rhodamine B 

610 chloride (from Exciton Inc.) powder dye. The peak absorption and emission wavelength of 

this dye are sufficiently separated, Figure 2.27(c), and thus, the recorded images had very low 

background noise. 

 (a) (b) (c) 

Figure 2.27: Laser reflection, lens filter characteristics, and fluorescent dye spectrum. 

Flow seeding was one of the main challenges in this experiment. Typically, PIV 

experiments conducted in air implement particle seeding through the use of a fogger. However, 

when fluorescent dye was mixed with the fog fluid for this experiment, the resulting fluid 

NASA/CR—2015-218868 37
mixture properties changed, and the particles started depositing on the interior of the window 

instead of following the flow. As an alternative particle introduction method, a TSI 9307-06 sixjet Laskin nozzle was used to atomize a mixture of fluorescent dye and propylene glycol. The 

particle generation with Laskin Nozzle was very sensitive to the specific seeding fluid mixture 

concentration. Successful seeding was achieved for this study using three grams of the dye 

powder mixed with 1 gallon of propylene glycol and 1.69 oz. of ethanol. The small amount of 

ethanol reduced the surface tension of the fluid mixture to improve atomization. Any deviation 

from this mixture prevented the generation of good tracer particles. Air at 75-80 psi pressure was 

supplied through a one inch tube at the Laskin nozzle inlet to generate micron-sized tracer 

particles which were seeded in the upstream location through a 0.5 in. hole into the center of the 

compressor inlet duct, Figure 2.28. 

Figure 2.28: Laskin nozzle used for seeding particles into upstream flow. 

The output from the seeder was much lower than the flow rate through the compressor, and 

thus, seeding was sparse. To get sufficient signal for cross correlation in an interrogation window, 

1000 image pairs were recorded for each phase locked position, and reasonable particle density 

was achieved when summing up the phase-locked images. Experiments were carried out using 

the TC3 tip clearance configuration at two operating conditions on the 100% corrected speedline 

(nominal loading and high loading), and measurements were acquired at 20 phase-locked 

positions across one rotor blade passage. 

2.6.2. Image Processing 

The top two camera images were used to reconstruct planar 3-component velocity fields for 

each phase locked measurement location. In-house PIV software “Prana” was used for all 

calibration, cross-correlation image processing, and three-component velocity reconstruction. A 

NASA/CR—2015-218868 38
polynomial mapping function with cubic order in x and y and linear order in z was used to map 

the world coordinate system (x, y, z) in the measurement domain to the image coordinate system 

(X, Y) for each camera. The image overlap after applying the calibration is shown in Figure 2.29. 

The axial overlap between the cameras increases from 100% span to 70% span and then 

decreases because of the steep viewing angles of the cameras. 

Figure 2.29: Top two camera overlap region after calibration. 

The stereo PIV data processing is accomplished using a generalized reconstruction, as 

described by Soloff et al. (1999), in which the individual camera images are first correlated and 

then the planar velocity fields from individual cameras (=, = for camera 1 and ;, ; for 

camera 2) are combined with the gradients of the mapping function (`a, `b) to obtain the , , 

and  velocity components using a least squares fit: 

c

=

=

;

;

d =

e

f

f

f

g

`aB

= `ah= `ai=

`bB

= `bh= `bi=

`aB

; `ah; `ai;

`bB

; `bh; `bi;

j

k

k

k

l

m







n	. (2.9)

After the individual camera image pairs were summed to identify the region with particles, 

the rest of the image was masked to remove erroneous contributions from low-quality vectors. 

Next, the minimum background intensity was subtracted from the images. The summed-up 

images were then cross-correlated to get planar velocity fields. To obtain a more robust velocity 

estimate, the individual image pairs were correlated and the correlation planes were summed 

which results in a higher signal to noise ratio, as shown by Meinhart and Santiago (1999). The 

cross correlation was performed using two passes: the first pass utilized a window size of 128 

pixels, and the second pass utilized the same window size with a 50% Gaussian spatial filter, as 

recommended by Eckstein and Vlachos (2009a). The grid resolution in first pass was 32 by 32 

and in the second pass was 8 by 8. Robust Phase Cross Correlation (RPC), as described in 

NASA/CR—2015-218868 39
Eckstein et al. (2008) and Eckstein and Vlachos (2009b), was used to correlate the image pairs. 

The planar fields were validated using velocity threshold and Universal Outlier Detection (UOD) 

to remove erroneous vectors. The planar velocity vector fields were then dewarped onto physical 

coordinate space to determine overlapping region, and the , , and  components were then 

reconstructed using Equation (2.9). The reconstructed fields were median filtered to remove 

noisy vectors along the blade edges. 

Tomographic PIV analysis (Elsinga et al., 2006) was also performed by combining images 

collected from all four cameras. For this purpose, the commercially-available software package 

LaVision DaVis was used. First, a volumetric calibration was performed for all cameras using 

masked calibration images to remove the portions of the image blocked by the blade. Then, 

tomographic reconstruction was performed using the sum of the 1000 pairs of images. 

The individual camera images were pre-processed with a sliding minimum background 

subtraction from a local 3 by 3 pixel region. The particle images were then smoothed using a 

Gaussian filter and a subsequent sharpening filter to remove background noise. These images 

were used for tomographic reconstruction in which individual particles were back-projected from 

each camera, and their positions in three dimensional space were determined using a 

multiplicative line of sight algorithm (Atkinson and Soria, 2009). 

The reconstructed 3D particle field was cross correlated using an FFT-based volume cross 

correlation with 3 passes. A first correlation pass used 50% overlap and a 128 by 128 by 64 

voxel window. This was followed by two more passes with 75% overlap and window sizes of 

128 by 128 by 32 and 64 by 64 by 32 voxels, respectively. Using calibration information, the 

reconstruction was performed on a domain with 659 by 330 by 67 voxels. The final cross 

correlation pass yielded 8 vector planes in the volume of the laser sheet. As a post-processing 

step for validating the calculated vectors, velocity threshold of 16 pixels was applied, smoothing 

and UOD techniques were performed, and a software option was selected to interpolate and fill 

any void regions in the vector field. 

NASA/CR—2015-218868 40
CHAPTER 3: STEADY COMPRESSOR PERFORMANCE 

Reid A. Berdanier, Natalie R. Smith, and Nicole L. Key 

3.1. Compressor Performance Maps 

For this study, a series of compressor performance data were collected at four corrected 

operating speeds (100%, 90%, 80%, and 68%) for each of the three tip clearance configurations 

(TC1-TC3). The 68% corrected speedline was selected instead of the 70% corrected speed to 

avoid a Campbell diagram crossing corresponding to a rotor first torsion (1T) vibratory mode, as 

analyzed by Murray (2014). In all cases, the corrected speed and corrected mass flow rate were 

calculated according to the procedure outlined in Section 2.2.1. For each of the four speedlines, 

data from different tip clearance configurations were collected at equivalent corrected mass flow 

rates for direct comparison. 

3.1.1. Total-to-Total Pressure Rise Characteristics 

Benchmark compressor map data are presented in Figure 3.1. Each measurement point in 

Figure 3.1 is calculated from area-averaged rake measurements collected from the seven-element 

total pressure rakes at 20 equally-spaced vane positions across one vane pitch. In this figure, the 

overall total pressure ratio (TPR) across the compressor is calculated as the ratio of radially and 

circumferentially area-averaged total pressures from axial planes 1 and 9 in the Figure 2.6 

schematic: 

TPR =

3,[,AA

3,=,AA

. (3.1)

In Figure 3.1, five operating points from the 100% corrected speedlines have been designated for 

comparison throughout this document: a low loading (LL) condition which represents a high 

flow rate, negative incidence condition; a nominal loading (NL) condition representing an 

operating point at a mass flow rate slightly higher than the peak efficiency point; a peak efficient 

(PE) point; a high loading (HL) condition which represents a low flow rate, high incidence 

operating point; and a set of conditions near the stall point (NS) which have a stall margin of 

approximately 5%. For this study, the stall margin (SM) is defined as follows: 

SM =

o

TPR

p
 q

r

stall

− o

TPR

p
 q

r

o

TPR

p
 q

r

× 100% . (3.2)

On the abscissa of Figure 3.1, the corrected inlet mass flow rates have been normalized by 

the value at the nominal loading point. For reference, the calculated relative uncertainty of the 

overall total pressure ratio in Figure 3.1 is 0.16% (within the symbol size in the figure). The 

maximum relative uncertainty of the normalized inlet corrected mass flow rate using the ASMEstandard (ASME PTC 19.5, 2004) set of equations is 1.4% and occurs at the near-stall loading 

condition. Nearly 60% of this calculated measurement uncertainty is contributed by the ASMEstandard uncertainty for discharge coefficient, which will be reduced in the future upon 

calibration of the Venturi flow meter. However, the repeatability of the mass flow rate 

NASA/CR—2015-218868 41
measurements for a fixed throttle position at the same loading condition is typically on the order 

of 0.2%. 

In Figure 3.1, the final point at the top of each speedline (the lowest flow rate condition) is 

not a 20-point traversed point, but rather a representative location of the stall point for the 

compressor at that operating speed. Once the stall point is identified in a preliminary run, it is 

subsequently approached slowly through incremental throttle movements to determine the mass 

flow rate and pressure rise locating the stall point. An approximate representation of the stall 

point with respect to the traversed points on the rest of the speedline can be determined by a 

comparison with the last (NS) traversed point and by stalling the compressor in several stator 

vane positions with respect to the fixed rake positions. The dotted line connecting these points 

represents a stall line for the specific tip clearance configuration. Comparing these results, an 

increase of rotor tip clearance leads to a decrease of overall total pressure rise and moves the stall 

point to a higher flow rate (the stall margin is reduced). Additional discussion about the stall 

characteristics of the compressor is given in Chapter 4. 

Figure 3.1: Compressor total pressure ratio map. 

3.1.2. Operating Tip Clearance Measurements 

The operating points defined in Figure 3.1 can be characterized by their nominal (design 

intent) rotor tip clearance height, as listed in Table 2.2. However, several environmental factors 

can influence the physical components of the facility and, in turn, introduce variations of the 

operating tip clearance. These factors include, but are not limited to, centrifugal effects related to 

rotational speed changes and corresponding blade elongation, thermal effects due to temperature 

rise through the compressor, and pressure forces resulting from the pressure rise through the 

compressor. In a previous study guided toward modeling tip clearance changes in a multistage 

0.4 0.6 0.8 1.0 1.2 

1.05 

1.10 

1.15 

1.20 

1.25 

1.30 

1.35 

1.40 

LL

NL

PE

HL

Normalized Inlet Corrected Mass Flow Rate, mc[−]

Total Pressure Ratio, TPR [−]

68% Nc

80% Nc

90% Nc

100% Nc

TC1

TC2

TC3

NASA/CR—2015-218868 42
axial compressor, Dong et al. (2014) separated the potential contributions of operating tip 

clearance changes into several categories, as suggested above: 

$model = $static − u >∆$Av,thermal + >∆$Av,centrif w + u >∆$Ax,thermal + >∆$Ax,pressure w

− u >∆$A},thermal + >∆$A},centrif w

(3.3)

where the subscripts B, S, and D represent tip clearance change contributions from the blade, 

shroud (casing), and disk, respectively. This equation calculates the operating “model” 

clearances, $model , based on the static (cold) build clearance, $static . Equation (3.3) shows a 

thermal expansion or centrifugal growth of the blade or disk will decrease the clearance height, 

whereas a thermal expansion or pressure force acting on the shroud will increase the clearance 

height. 

This model has been applied specifically to the Purdue three-stage axial compressor to 

compare measured tip clearances with model predictions (Berdanier and Key, 2015a), and the 

outcome showed that the predicted model clearances are driven primarily by the thermal growth 

terms, which yield clearance change contributions several orders of magnitude larger than the 

pressure or centrifugal components. Ultimately, this previous study (Berdanier and Key, 2015a), 

in combination with the original model analysis presented by Dong et al. (2014), has confirmed 

that considerable changes in operating tip clearance height can be achieved as the loading 

condition is changed along a given speedline or between speeds. 

The operating tip clearances for this project were measured using the capacitance probe 

system outlined in Section 2.4. Beginning with the baseline tip clearance configuration (TC1), 

the measured operating tip clearances for the steady operating points defined in Figure 3.1 are 

presented in Figure 3.2 as a function of the normalized inlet corrected mass flow rate for each of 

the four identified speeds. These clearance data have been determined using the RMS calculation 

method, although an arithmetic mean of the blade-by-blade DNS data is similar. In Figure 3.2, 

the three circumferentially-distributed probes for each rotor row have been averaged to yield one 

representative tip clearance value for each rotor row. 

Figure 3.2: Operating tip clearances measured using capacitance probes for TC1. 

0.75 0.95 1.15 

0.032 

0.033 

0.034 

0.035 

0.036 

0.037 

0.038 

0.039 

Normalized Inlet Corrected Mass Flow Rate, mc[−]

Average Operating Rotor Tip Clearance, τ [in]

100% Nc

0.65 0.85 1.05 

0.032 

0.033 

0.034 

0.035 

0.036 

0.037 

0.038 

0.039 

90% Nc

Rotor 1

Rotor 2

Rotor 3

0.55 0.75 0.95 

0.032 

0.033 

0.034 

0.035 

0.036 

0.037 

0.038 

0.039 

80% Nc

0.45 0.65 0.85 

0.032 

0.033 

0.034 

0.035 

0.036 

0.037 

0.038 

0.039 

68% Nc

NASA/CR—2015-218868 43
Considering Figure 3.2, these operating points were collected in immediate succession 

moving from left to right across the four parts of the figure. In this way, the results display the 

variation of rotor tip clearance height as the operating point changes with minimal point-to-point 

effect from ambient temperature changes. However, there are several noticeable effects of 

changes in ambient conditions. First, the final high flow rate condition for the 90% corrected 

speedline moving into the 80% corrected speed conditions represents a period of time at the 

beginning of a day when the ambient temperature was increasing; Second, there is a significant 

drop of measured tip clearances between the third and fourth points (from left to right) in the 80% 

speedline. A stretch of looming severe weather forced a stop of the experimental campaign after 

the completion of the third point in this line, and the data collection process resumed beginning 

with the fourth point on the following day. This discontinuity represents a change of 

approximately 2.5×10-3 in., which represents an 8% change of operating tip clearance for the 

nominal value of 0.030 in., and it is attributable to a change in ambient temperature of 

approximately 28 oF. It is important to note here that there is no measureable change of 

performance parameters at this speed and for this loading condition (away from the stall point at 

part-speed) due to this small measured tip clearance change, whereas a high loading point on the 

100% corrected speedline may be more likely to show a measurable performance change (see 

Section 3.1.3). Third, the 80% and 68% operating speed curves in Figure 3.2 show less 

relationship between measured tip clearance and loading condition (compared to 100%, e.g.) as 

the temperature rise through the compressor is significantly reduced in this part-speed operating 

regime. Thus, a majority of the observed decrease of tip clearance with loading conditions in the 

68% corrected speed range is due to a decrease in ambient temperature that occurred in the 

evening. 

Similar measurements are presented for the two larger tip clearance configurations, TC2 

and TC3 in Figure 3.3 and Figure 3.4. As with the TC1 results in Figure 3.2, these larger tip 

clearance measurements were also collected in succession, but without the discontinuity in the 80% 

speedline discussed for TC1. Although the mean weather conditions were extremely similar 

between measurements collected for the three different tip clearances, the TC2 and TC3 data 

collection processes began at a different time of day than TC1. As a result, a comparison of the 

results in Figure 3.3 and Figure 3.4 show a very similar trend which is slightly different than the 

observed trends for TC1. In particular, the 100% speedline data for TC2 and TC3 began in the 

early morning hours prior to sunrise and, therefore, the high-loading, low flow rate operating 

points which showed the highest tip clearance values for TC1 display a less discernable trend as 

the highest static temperature rise conditions are affected slightly by the slow temporal variation 

of temperature through the day. 

NASA/CR—2015-218868 44
Figure 3.3: Operating tip clearances measured using capacitance probes for TC2. 

Figure 3.4: Operating tip clearances measured using capacitance probes for TC3. 

The operating tip clearance measurements in Figure 3.2, Figure 3.3, and Figure 3.4 are a 

crucial component of this overall measurement campaign. In particular, these results have shown 

the potential extent to which even a small variation of ambient temperature can impact the tip 

clearance height. Furthermore, it is entirely insufficient to report the design intent tip clearance 

heights, as in Table 2.2, as a result of the potential variations in clearance height between 

operating conditions alone. For absolute consistency, these measured operating tip clearance 

values are used throughout the remainder of the document whenever clearance values at a 

specific operating condition are required. The following section explores the potential influence 

that these tip clearance height changes can pose for overall performance measurements. 

0.75 0.95 1.15 

0.063 

0.064 

0.065 

0.066 

0.067 

0.068 

0.069 

0.070 

Normalized Inlet Corrected Mass Flow Rate, mc[−]

Average Operating Rotor Tip Clearance, τ [in]

100% Nc

0.65 0.85 1.05 

0.063 

0.064 

0.065 

0.066 

0.067 

0.068 

0.069 

0.070 

90% Nc

Rotor 1

Rotor 2

Rotor 3

0.55 0.75 0.95 

0.063 

0.064 

0.065 

0.066 

0.067 

0.068 

0.069 

0.070 

80% Nc

0.45 0.65 0.85 

0.063 

0.064 

0.065 

0.066 

0.067 

0.068 

0.069 

0.070 

68% Nc

0.75 0.95 1.15 

0.085 

0.086 

0.087 

0.088 

0.089 

0.090 

0.091 

0.092 

Normalized Inlet Corrected Mass Flow Rate, mc[−]

Average Operating Rotor Tip Clearance, τ [in]

100% Nc

0.65 0.85 1.05 

0.085 

0.086 

0.087 

0.088 

0.089 

0.090 

0.091 

0.092 

90% Nc

Rotor 1

Rotor 2

Rotor 3

0.55 0.75 0.95 

0.085 

0.086 

0.087 

0.088 

0.089 

0.090 

0.091 

0.092 

80% Nc

0.45 0.65 0.85 

0.085 

0.086 

0.087 

0.088 

0.089 

0.090 

0.091 

0.092 

68% Nc

NASA/CR—2015-218868 45
3.1.3. Effect of Ambient Conditions on Tip Clearance and Total-to-Total Pressure Rise 

Changes of mechanical speed necessary to match corrected speed conditions, as well as 

thermal blade growth, can cause rotor tip clearance heights to change. The effect of changing tip 

clearances due to ambient conditions, as introduced in Section 3.1.2 may seem to be a negligible 

difference in the scale of the overall clearance height. In fact, Walsh and Fletcher (2008) explain 

that these corresponding differences are usually ignored. However, if great care is taken to 

reduce measurement uncertainty, and all known corrected conditions are appropriately accounted 

for, then the variations due to these very small tip clearance changes are measurable and 

repeatable. 

As part of this study, a series of corrected speedlines were collected for different tip 

clearance configurations (TC2 and TC3) for two different ambient temperature conditions. These 

results, Figure 3.5, show that measurable changes may exist. In this case, a “hot” day represents 

an average ambient temperature of 77 oF and a “cold” day represents an average ambient 

temperature of 24 oF for measurements collected with both TC2 and TC3. As with the data 

shown in Figure 3.1, these results represent area-averaged rake measurements collected at 20 

equally-spaced vane positions across one vane pitch. The performance discrepancies highlighted 

by Figure 3.5 are most significant at high loading (low flow rate) conditions approaching the stall 

point. At these high loading conditions, the two operating points are up to 0.42% different in 

overall total pressure ratio (nearly four times the uncertainty of the same quantity). 

Figure 3.5: Overall compressor total pressure ratio discrepancy between hot 

and cold test days at a 100% corrected operating speed for TC2 and TC3. 

The differences highlighted in Figure 3.5 are consistent with untraversed data (i.e., 

measurements collected at one discrete position with respect to the stationary vanes) acquired 

over the course of a full calendar year in the same facility, as presented in Figure 3.6 for all three 

tip clearance configurations (TC1-TC3). Over the same range of ambient temperatures, TC3 is 

less affected (approximately 0.15% different in overall total pressure ratio at the near stall 

operating conditions in Figure 3.5), compared to the TC1 and TC2 configurations. Specifically, 

0.8 0.9 1.0 1.1 1.2 

1.15 

1.20 

1.25 

1.30 

1.35 

Normalized Inlet Corrected Mass Flow Rate, mc[−]

Total Pressure Ratio, TPR [−]

TC2 Hot

TC2 Cold

TC3 Hot

TC3 Cold

NASA/CR—2015-218868 46

Figure 3.6 highlights a negatively-sloping trend of total pressure ratio with increasing inlet 

temperature, whereas TC3 does not display the same distinct trend. 
Figure 3.6: Variations of total pressure ratio with ambient temperature for each of the three 

nominal tip clearance conditions (TC1-TC3). 

The overall compressor total pressure rise is assessed for each of the three tip clearance 

configurations with hot and cold inlet conditions at a near stall (NS) operating condition in 

Figure 3.7 as a function of the average Rotor 1 tip clearance height. The results in Figure 3.7 

show that tip clearance changes with ambient temperature may not be negligible (e.g., 2.5×10-3

in. for a 25 oF change of ambient temperature) as some authors suggest. For the three-stage axial 

compressor facility utilized in this study, the aluminum compressor casing has a coefficient of 

thermal expansion which is more than two times larger than that of the stainless steel blades. As 

a result, changes of ambient inlet temperature introduce corresponding tip clearance changes 

which can be significant (on the order of 0.1% annulus height). These observations, along with 

the data presented from Section 3.1.2, suggest that it is a necessity to measure operating tip 

clearance for studies such as this which desire to clearly evaluate tip clearance effects on 

compressor performance. 

0 20 40 60 80 100

−0.7

−0.6

−0.5

−0.4

−0.3

−0.2

−0.1

0

Inlet Temperature [ oF ]

(TPR−TPRmax)/TPRmax

TC1

0 20 40 60 80 100

−0.7

−0.6

−0.5

−0.4

−0.3

−0.2

−0.1

0

Inlet Temperature [ oF ]

TC2

0 20 40 60 80 100

−0.7

−0.6

−0.5

−0.4

−0.3

−0.2

−0.1

0

Inlet Temperature [ oF ]

TC3

NASA/CR—2015-218868 47
Figure 3.7: Overall total pressure ratio trends with rotor tip clearance for three tip clearance 

configurations (TC1-TC3), each for two separate ambient temperatures at NS. 

To further investigate the observed performance deltas with ambient temperature changes 

due to clearance height, the normalized total pressure ( NRP3 ) is defined to represent the 

circumferential average at each radial position, normalized by the radial and circumferential area 

average at the same axial position: 

NRP3,\ =

3,\,~

3,\,AA

. (3.4)

Using this definition affords the ability to compare the relative shapes of the radial total pressure 

profiles without the inherent offset due to changes in overall pressure rise with increased rotor tip 

clearance height. Investigation of the radial total pressure ratio shapes for all three tip clearance 

configurations, Figure 3.8, shows an increased total pressure loss in the tip region for hot days, 

suggesting a change in the rotor tip leakage flow which leads to a measurable performance 

change. The uncertainty of the quantities presented in Figure 3.8 is within the symbol size of the 

figure. Only the results at Rotor 1 exit and Stator 2 exit are shown in Figure 3.8 because they 

represent the most significant differences between the two temperature conditions. As expected, 

the profiles shapes for TC2 fall between the results for TC1 and TC3. 

In agreement with the results in Figure 3.5, the cold data generally exhibit less total 

pressure loss at the tip (except for TC3). The less distinct differences observed for TC3 

compared to TC1 or TC2 are consistent with the less noticeable trend in Figure 3.6(c). This may 

be explained by considering that a 2.5×10-3 in. change in tip clearance is less significant for a 

0.080 in. nominal tip clearance than for a 0.030 or 0.060 in. nominal tip clearance height. The 

trends for TC2 are nearly as strong as those for TC1, an observation which could suggest TC2 

represents a critical point in the loss development with increased tip clearance. For example, TC2 

is near the point where the hub and tip losses switch trends in the Stator 2 exit profiles, as shown 

in Figure 3.8. Additional specifics regarding these observed performance differences with 

ambient conditions are discussed by Smith et al. (2015a). 

0.02 0.04 0.06 0.08 0.10 

1.32 

1.33 

1.34 

1.35 

1.36 

Average Rotor 1 Tip Clearance Height, τ [in]

Total Pressure Ratio, TPR [−]

TC1 Hot

TC1 Cold

TC2 Hot

TC2 Cold

TC3 Hot

TC3 Cold

NASA/CR—2015-218868 48
Figure 3.8: Radial total pressure profile shapes for the three tip clearance 

configurations with “hot” and “cold” ambient operating conditions at NS. 

3.1.4. Stall Margin 

Using Equation (3.2), the stall margin has been calculated for each of the primary identified 

points on the 100% corrected speedline. These results, Figure 3.9, are presented with the average 

measured tip clearance on the abscissa, as described in Section 3.1.2. Based on the results in 

Figure 3.9, the stall margin follows a non-linear trend with increasing tip clearance for each of 

the four chosen loading conditions: the decrease of stall margin is more significant moving from 

TC2 to TC3 than from TC1 to TC2 at the same corrected mass flow rate. This difference 

increases more noticeably at higher flow rates (lower loading conditions) than at the lower flow 

rates (higher loading conditions). This observed trend also exists for other definitions of stall 

margin (i.e., based solely on corrected mass flow rates or based on the total pressure ratios in 

Equation (3.1)). 

Previous studies of tip clearance effects on overall compressor performance have identified 

the loss of operability range using a percent loss of flow coefficient instead of stall margin 

(Wisler, 1985; McDougall et al., 1990), for which the flow coefficient is defined by the ratio of 

the axial flow velocity, B, to the rotor tip speed, : 

Φ =

B



. (3.5)

The same quantities can be compared with these data, using a percent difference of the flow 

coefficient at the stall point as a percent difference from the baseline tip clearance (TC1): 

ΔΦstall =

Φstall − Φstall,TC1

Φstall,TC1

× 100% . (3.6)

Using this definition with a four-stage low-speed research compressor, Wisler shows an 11% 

increase of stalling flow coefficient for a rotor tip clearance height increase of 1.4% to 2.8% 

0.99 1 1.01

12

20

35

50

70

80

88

NRPo,3 [−]

Annulus Height [%]

Rotor 1 Exit

0.99 1 1.01

12

20

35

50

70

80

88

NRPo,6 [−]

Stator 2 Exit

TC1 Cold

TC1 Hot

TC2 Cold

TC2 Hot

TC3 Cold

TC3 Hot

NASA/CR—2015-218868 49
based on annulus height; the data from McDougall et al. show approximately 10% increase for a 

tip clearance height increase of 0.9% to 2.2% based on annulus height in a single-stage lowspeed compressor. If the same method is applied to these data, the results in Figure 3.9 show a 

3.0% increase of stalling flow coefficient between TC1 and TC2 and a 4.7% increase between 

TC2 and TC3. These stalling flow coefficient data show that the performance of this three-stage 

axial compressor is less sensitive to tip clearance changes than has been previously reported for 

single-stages, isolated rotors, and repeating stage low-speed axial compressors. In this case, it is 

possible that the multistage compressor design with non-repeating stages provides the 

opportunity for stage matching adjustments which may postpone stall. 

Figure 3.9: Stall margin and stalling flow coefficient effects with tip clearance 

height for specified loading conditions on the 100% corrected speedline. 

3.1.5. Total-to-Static Pressure Rise Characteristics 

As an alternative to the total pressure ratio as a metric for achieved pressure rise, some 

authors (Hutton, 1956; Wisler, 1985; McDougall, 1990) have discussed the total-to-static 

pressure rise coefficient, defined as: 

Ψ =

[ − 3=

=

;

"

;

. (3.7)

Figure 3.10 shows this total-to-static pressure rise coefficient, presented as a function of the flow 

coefficient, Φ, for each of the three tip clearance configurations at all four corrected operating 

speeds. As with Figure 3.1, the point at the lowest inlet mass flow rate for each of the curves 

represents the stall point of the compressor. The results in Figure 3.10 show the expected result 

of decreasing total-to-static pressure rise coefficient as the tip clearance is increased. Another 

observations from Figure 3.10 is for the 100% corrected speedline of TC3, which has a location 

1.0 2.0 3.0 4.0 5.0 

10

20

30

40

50

60

70

80

Normalized Tip Clearance, τ / H [%]

Stall Margin, SM [%]

TC1

TC2

TC3

0

2

4

6

8

10

12

Stalling Flow Coefficient Difference, ∆Φstall

LL NL PE HL

[%]

NASA/CR—2015-218868 50
in the characteristic at a flow coefficient of approximately 0.5 where the slope of the curve 

changes sharply to be less negative (approaching zero). This location of the characteristic also 

corresponds to a region of the curve which has a slight positive slope in the speedline from 

Figure 3.1. This is the only speedline from the 12 presented lines which features a total pressure 

rise decrease prior to the stall point, which may explain the unique feature of this characteristic 

in Figure 3.10. Also in this figure, the results for different speeds with a given tip clearance 

configuration do not collapse to one single line, alluding to the significance of compressibility 

effects in the Purdue three-stage compressor. 

Figure 3.10: Total-to-static pressure rise characteristics for three tip clearance configurations. 

Using the definition from Equation (3.6), previous authors have shown a decrease of peak 

total-to-static pressure rise coefficient on the order of 9.7% (Wisler, 1985) and 12.5% 

(McDougall, 1990) for a tip clearance change on the order of TC1 to TC2 (nominally 1.5% to 

3.0%). Figure 3.11 shows this same comparison using a percent difference of the total-to-static 

pressure rise coefficient with respect to the 1.5% nominal tip clearance (TC1): 

ΔΨ =

Ψ − ΨTC1

ΨTC1

× 100% , (3.8)

presented as a function of average measured tip clearance. In contrast to the results discussed by 

previous authors, the data in Figure 3.11 show that the three-stage compressor studied here has a 

total-to-static pressure rise coefficient decrease of approximately 4.0% for a similar tip clearance 

change from TC1 to TC2. In fact, even a change from TC1 to TC3 (nearly twice the change 

introduced by Wisler or McDougall et al.) only results in an 8.5% decrease in total-to-static 

pressure rise coefficient compared to the baseline configuration TC1. 

However, the results of Figure 3.11 also show an important difference that a clearance 

change from TC2 to TC3 follows a markedly different trend that that between TC1 and TC2. 

0.4 0.5 0.6 0.7

0.6

0.8

1

1.2

1.4

1.6

1.8

2

2.2

Total−to−Static Pressure Rise Coefficient, 

Ψ [−]

TC1

100% Nc

90% Nc

80% Nc

68% Nc

0.4 0.5 0.6 0.7

0.6

0.8

1

1.2

1.4

1.6

1.8

2

2.2

Flow Coefficient, Φ [−]

TC2

0.4 0.5 0.6 0.7

0.6

0.8

1

1.2

1.4

1.6

1.8

2

2.2

TC3

NASA/CR—2015-218868 51
Thus, the assumed linearity suggested by many previous authors may not be valid for this 

compressor application or for large ranges of tip clearance heights. Despite this observation, it 

can be noted that the trends of ΔÅ appear to be relatively insensitive to rotational speed, as 

shown by the nearly collapsed results in Figure 3.11. 

Figure 3.11: Peak total-to-static pressure rise coefficient as a function of tip clearance height. 

3.1.6. Isentropic Efficiency 

As explained earlier, the isentropic efficiency of the compressor has been evaluated using 

the formal definition from Equation (2.3). The required stagnation enthalpies for this equation 

are calculated using the thermodynamic equation program REFPROP (Lemmon et al., 2013) 

through the use of measured conditions (stagnation temperature, stagnation pressure, static 

pressure, and relative humidity) at axial positions 1 and 9 – in agreement with the measurement 

planes used to define total pressure ratio in Equation (3.1). Using Equation (2.3), the isentropic 

compressor efficiency has been calculated for each of the measurement points shown in the 

overall compressor total pressure ratio map of Figure 3.1 (neglecting the stall points). These 

efficiencies, shown in Figure 3.12, are presented for each of the four corrected rotational speeds 

as a function of the normalized inlet corrected mass flow rate. 

The uncertainty of the efficiency values in Figure 3.12 can be influenced by uncertainty 

contributions from measured pressures, temperatures, and humidity. Great care has been taken to 

ensure that the respective uncertainty for each of these measured parameters is as small as 

possible (see Section 2.2). With this in mind, the efficiency uncertainty has been carefully 

evaluated using the T-s region technique implemented by Lou et al. (2013). The uncertainty 

values for the 100% corrected speedlines in Figure 3.12 are on the order of 5 points across the 

entire flow range for all tip clearances. Similarly, the 90% corrected speedline data have 

efficiency uncertainties of approximately 6 points across the entire flow range. At the lower 

1.0 2.0 3.0 4.0 5.0 

1.8

1.9

2.0

2.1

2.2

Normalized Tip Clearance, τ / H [%]

Peak Total−to−Static Pressure Rise, 

Ψpk [−]

TC1

TC2

TC3

−9

−6

−3

0

Percent Difference, ∆Ψpk [%]

100% Nc

90% Nc

80% Nc

68% Nc

NASA/CR—2015-218868 52
speeds of 80% and 68% corrected speed, however, the reduced pressure rise leads to a reduced 

temperature rise (although still measurable and appreciable), especially at high flow rate 

conditions. As a result, the uncertainties for the 80% corrected speedline data vary from 

approximately 6 points at low flow, high loading coditions to as much as ten points for the high 

flow, low loading conditions. At the 68% speed conditions, the uncertainties vary from 

approximately 8 points at low flow conditions to more than 15 points at high flow conditions. 

These relatively large uncertainty values are primarily due to the propagation of 

measurement uncertainty when calculating enthalpies, especially hU[<, as explained by Lou et al. 

(2013). Indeed, the data for this project also show that hU[< is the largest component of the 

efficiency uncertainty. Despite the large measurement uncertainties, however, the repeatability of 

the efficiencies shown in Figure 3.12 is typically on the order of 0.5 points – an order of 

magnitude improvement over the calculated uncertainty. 

As an additional consideration, measurement variability due to machining tolerances, blade 

row interactions, and more, can also influence calculated performance parameters such as 

isentropic efficiency. The measurement locations for data presented here were carefully selected 

to avoid these effects as much as possible based on prior knowledge of the machine. Nonetheless, 

a vane wake variability study characterizing the circumferential variation in the wakes shed from 

a particular stator row was recently conducted by Methel et al. (2015). The results from this 

study, combined with a similar investigation of passage-to-passage variations in total 

temperature, have shown that the circumferential variation in the flow field is small (on the order 

of measurement uncertainty) yet repeatable. Thus, efficiencies calculated by combining pressure 

and temperature measurements from different sectors of the machine can be influenced by this 

variation. The only way to avoid this error in the absolute value of efficiency would be to use a 

measurement rake with combined total pressure and total temperature sensors. However, the 

measured changes in efficiency between different configurations are valid and meaningful. 

Figure 3.12: Overall isentropic compressor efficiency at four corrected rotational speeds. 

Beginning with the data presented in Figure 3.13, the compressor efficiency is separately 

considered for each of the three tip clearance configurations at the four operating conditions of 

primary interest on the 100% corrected speedline: LL, NL, PE, and HL. Figure 3.13 shows the 

0.75 0.95 1.15 

4 pts

Isentropic Compressor Efficiency, [pts] η

Normalized Inlet Corrected Mass Flow Rate, mc[−]

100% Nc

TC1

TC2

TC3

0.65 0.85 1.05 

90% Nc

0.55 0.75 0.95 

80% Nc

0.45 0.65 0.85 

68% Nc

NASA/CR—2015-218868 53
expected result of decreased compressor efficiency with increased rotor tip clearance, as 

discussed in earlier studies for isolated rotors and multistage machines. At each of the four 

operating points, the efficiency decreases linearly with increasing tip clearance. Furthermore, the 

constant of linearity is nearly identical for all four operating points (certainly within the 

repeatability of the measurements). 

Figure 3.13 presents a decrease of approximately 2.5 points at the peak efficiency point 

between TC1 and TC2 as compared to Wisler’s (1985) canonical observation of 1.5 points for 

the same change in clearance size. In a four-stage low-speed compressor similar to the design 

used by Wisler, Tschirner et al. (2006) showed an efficiency decrease of 3.6 points between TC1 

and TC2 (also 1.5% and 3.0% with respect to annulus height) at the design point. However, 

because Tschirner et al. used cantilevered stators, and the stator hub clearances were increased at 

the same rate as the rotor tip clearances, the resulting decreases in pressure rise and efficiency 

are larger than a study which studies rotor tip clearances exclusively. 

The compressor performance data presented by Freeman (1985) suggest a rule-of-thumb of 

approximately 3% to 5% stall margin lost per 1% increase of tip clearance and approximately 1.4 

to 2 efficiency points lost per 1% increase of tip clearance height. These results are based on 

linear fits of data collected from several engine tests and research compressor studies. Applying 

Freeman’s suggestions to these tip clearance values would suggest an efficiency loss on the order 

of 2 to 3 points and a stall margin loss on the order of 5 to 8 percent comparing TC1 and TC2. 

The results presented herein are certainly in agreement with this rule of thumb, but a comparison 

of these data with results from other authors discussed so far also emphasize Freeman’s 

important observation that the appreciable scatter of data about a general trend shows discernable 

differences between different machines, airfoil families, levels of aerodynamic loading, and 

rotational speed regimes. 

Figure 3.13: Isentropic compressor efficiency for specified 100% .speedline points as a 

function of tip clearance height. 

1.0 2.0 3.0 4.0 5.0 

TC1

TC2

TC3

2 pts

Normalized Tip Clearance, τ / H [%]

Isentropic Compressor Efficiency, 

η [pts]

LL NL PE HL

NASA/CR—2015-218868 54
3.1.7. Stator 3 Seal Leakage Flow Measurements 

To adequately define boundary conditions for computational models, the leakage flow 

paths for the compressor have been carefully evaluated. At the inlet of the compressor, the air 

paths are all sealed, but a potential leakage path exists under the Stator 3 labyrinth (knife) seals 

to atmospheric pressure. Ball (2013) considered these leakage flow paths and the related knife 

seal geometry for computational models. Ball determined that whether the Stator 3 seal flow was 

allowed to exit to ambient pressure, or if the respective flow paths were plugged, the effect on 

overall compressor pressure rise was negligible. However, a measureable change of overall 

compressor efficiency was observed, likely due to a localized heating of the rear bearing and 

other components at the rear of the compressor in the absence of the Stator 3 leakage flow path. 

For this study, the leakage pathways remained open to the atmosphere, but the air was 

connected to a manifold through four tubes at the rear of the compressor, Figure 3.14, to capture 

the leakage flow. These tubes delivered the leakage flow to a Wyatt Engineering orifice plate 

with a discharge coefficient of 0.6042 and a diameter ratio of 2.294 (as provided by the 

manufacturer). The pressure drop across the orifice plate was measured by a Rosemount 3051C 

differential pressure transmitter with a calibrated range of 0-13 inH2O. The representative overall 

uncertainty of mass flow rate measurement for this orifice plate meter setup is less than 2.4% of 

the calculated value for mass flow rate. 

Figure 3.14: Stator 3 seal leakage flow path and manifold tubes from (a) the rear bearing plate to 

(b) an orifice plate run (Brossman, 2012). 

The measured Stator 3 seal leakage mass flow rates for each of the measured points in 

Figure 3.1 (seal leakage flow rate data for the stall point were not collected) are shown in Figure 

3.15, normalized as a percentage of the inlet corrected mass flow rate: 


 LN =




leak,.


 .

× 100% , (3.9)

for which the measured leakage mass flow rate is first corrected using the inlet conditions of the 

compressor using Equation (2.2). The increase of leakage flow rate at lower inlet mass flow rate 

conditions in Figure 3.15 suggests that the leakage mass flow rate is affected by the pressure rise 

NASA/CR—2015-218868 55
through the compressor (increased loading and pressure rise at low flow rates leads to increased 

leakage flow rate as the pressure differential with respect to ambient pressure increases). 

However, the curves in Figure 3.15 are independent of tip clearance configuration, showing that 

the leakage flow rate is influenced more by the flow rate through the compressor than the overall 

pressure rise through the machine (in which case, the leakage flow rates would change more 

significantly for different tip clearance configurations). The one potential exception to this is the 

final two to three points on the 100% speedline for TC3, which appear to roll over unlike the two 

smaller tip clearances TC1 and TC2. This observation aligns with the previous discussions about 

total pressure ratio and total-to-static pressure rise characteristics. 

Figure 3.15: Stator 3 seal leakage mass flow rate as a percentage of inlet corrected mass flow rate 

for each of the traversed data points in Figure 3.1. 

3.1.8. Casing Outer Diameter Surface Temperature Measurements 

Often, turbomachinery system analyses assume that the machine operates adiabatically. 

This assumption is based on the difference in order of magnitude between the work done by the 

rotor on the flow and that of the heat flux from the device. However, an analysis of the total 

temperature profiles through the compressor (see Section 3.3.2) show an increase of the fluid 

total temperature in the tip region near the compressor casing, especially for the front stages. In 

this case, the more significant temperature of the aft section of the compressor conducts through 

the aluminum casing and influences the flow at the front of the machine. As a result, alternative 

boundary condition techniques may be better suited for computational analyses, such as the 

isothermal wall boundary condition discussed by Bruna and Turner (2013) or a conjugate heat 

transfer analysis. 

0.55 0.75 0.95 1.15 

0.0 

0.2 

0.4 

0.6 

0.8 

1.0 

Normalized Inlet Corrected Mass Flow Rate, mc[−]

Normalized Leakage Mass Flow Rate, 

mLN [%]

68% 

Nc

80% 

Nc

90% 

Nc

100% 

Nc

TC1

TC2

TC3

NASA/CR—2015-218868 56
Several surface-mounted T-type thermocouples were installed on the outside of the 

compressor casing to isolate these temperature measurements. One thermocouple was installed 

over each of the seven blade rows at a position noted in Figure 2.11. A separate analysis of the 

circumferential uniformity of the casing temperatures showed the circumferential variations are 

less than 1.8 oR. Prior to collecting data at each operating point, an additional surface-mounted 

K-type thermocouple installed over Rotor 3 was allowed to reach a steady operating condition to 

prevent any transient temperature effects in the collected data. 

Figure 3.16: Normalized surface temperatures for three loading conditions on the 100% 

corrected speedline. 

Beginning with Figure 3.16, the casing outer diameter surface temperature trends are 

shown at three selected loading conditions (LL, NL, and HL) for all three tip clearance 

configurations. These temperatures are presented as a normalized temperature ratio with respect 

to the area-averaged temperature measured by the total temperature rakes at the AIP (plane 0 in 

Figure 2.6). As the work imparted by the rotor on the fluid increases with loading condition, the 

temperature of the outer compressor casing also increases through the compressor. Figure 3.17 

shows the same normalized temperature ratio, but for the final “near-stall” traversed operating 

point, at each of the four speeds and all three tip clearance configurations. Figure 3.16 and Figure 

3.17 both show that the TC1 data continue to experience temperature rise, even at the Stator 3 

measurement position, whereas the larger tip clearance configurations (TC2 and TC3) appear to 

level off more significantly comparing the surface temperatures over Rotor 3 and Stator 3. 

AIP IGV R1 S1 R2 S2 R3 S3

1.00 

1.02 

1.04 

1.06 

1.08 

1.10 

LL

NL

HL

Axial Measurement Position

Normalized Surface Temperature [−]

LL

NL

HL

TC1

TC2

TC3

NASA/CR—2015-218868 57
Figure 3.17: Normalized surface temperatures for the final “near-stall” traversed operating point 

at all four operating speeds. 

3.2. Stage Performance Characteristics 

In addition to the overall performance, the stage-by-stage compressor performance has also 

been considered for this study. The stage total pressure ratios are shown in Figure 3.18 as a 

function of the stage inlet corrected mass flow rate. For this figure, the stage inlet corrected mass 

flow rate uses the same equation for the inlet corrected mass flow rate, Equation (2.2), but the 

values for stagnation density and stagnation speed of sound are selected to represent the 

measured parameters at the inlet to the stage. Stage 1 is defined for axial positions 1 to 4, Stage 2 

is defined for axial positions 4 to 6, and Stage 3 is defined for axial positions 6 to 9. By these 

definitions, the combination all three stages results in the overall total pressure ratios from Figure 

3.1. 

Figure 3.18 allows several important observations regarding the performance of the threestage compressor, particularly relating to off-design performance and stage matching. Of note, 

each of the three stages has distinctly different characteristic shapes at all speeds and tip 

clearance configurations. The Stage 1 characteristics show the total pressure rise turns to a 

positive slope near the top of the curve as the compressor approaches stall. The Stage 2 curves all 

show a position where the slope of the curve changes sharply. This location appears at the peak 

efficiency point of the compressor for each curve – these characteristics could be approximated 

by two linear regions. Finally, the Stage 3 curves have a curvature which more closely resembles 

Stage 1 than Stage 2, but with less total pressure rise. 

AIP IGV R1 S1 R2 S2 R3 S3

1.00 

1.02 

1.04 

1.06 

1.08 

1.10 

68% Nc

80% Nc

90% Nc

100% Nc

Axial Measurement Position

Normalized Surface Temperature [−]

68% Nc

80% Nc

90% Nc

100% Nc

68% Nc

80% Nc

90% Nc

100% Nc

68% Nc

80% Nc

90% Nc

100% Nc TC1

TC2

TC3

NASA/CR—2015-218868 58
Figure 3.18: Stage total pressure rise characteristics for the traversed map points. 

For Stage 1 and Stage 3, the operating points closest to the stall point (the actual stall points 

are not shown in Figure 3.18) show a noticeable and consistent decrease of total pressure rise 

with increased tip clearance height. This same observation may be made for the 100% corrected 

speedline of Stage 2, but the part-speed operating conditions (90%, 80% Nc, and 68% Nc) for 

Stage 2 show a more consistent offset of results across the entire speedline. Figure 3.18 also 

shows that the individual stage total pressure rise may actually increase by a small amount for 

the larger tip clearance configurations (TC2 and TC3) at the highest flow rate position on the 100% 

speedline for Stage 1. An analysis of the individual rake pressure measurements which combine 

to create the area-averaged total pressure ratios shown in Figure 3.18 reveals that the lower 60% 

annulus height performs at a slightly higher total pressure exiting Stator 1 – at the expense of a 

more significant endwall blockage and total pressure loss in the tip region. Although the 

magnitudes of these differences are greater than the repeatability of the measurements, they are 

on the order of the uncertainty of the measurements, so no conclusions can be attributed to the 

observations. 

These stage pressure rise characteristics bring rise to a consideration of stage matching 

effects in the compressor. It is expected that the Purdue three-stage axial compressor is wellmatched at the design speed with the design tip clearance configuration (TC1). Comparing the 

TC1 performance with TC2 throughout Figure 3.18, the trends of the curves are very similar – 

many of the characteristics overlap for significant portions of the flow range. This suggests that 

the stage matching of the compressor may not be significantly altered by a change from 1.5% to 

3.0% rotor tip clearance height. However, more distinct differences exist for the largest tip 

clearance configuration (TC3).

0.4 0.6 0.8 1.0 1.2 

1.02 

1.04 

1.06 

1.08 

1.10 

1.12 

Normalized Stage 1 Inlet

Corrected Mass Flow Rate [−]

Stage 1

Stage Total Pressure Ratio [−]

0.4 0.6 0.8 1.0 1.2 

1.02 

1.04 

1.06 

1.08 

1.10 

1.12 

Normalized Stage 2 Inlet

Corrected Mass Flow Rate [−]

Stage 2

68% Nc

80% Nc

90% Nc

100% Nc

0.4 0.6 0.8 1.0 1.2 

1.02 

1.04 

1.06 

1.08 

1.10 

1.12 

Normalized Stage 3 Inlet

Corrected Mass Flow Rate [−]

Stage 3

TC1

TC2

TC3

NASA/CR—2015-218868 59
3.3. Inter-Stage Flow Measurements 

In addition to the performance metrics discussed thus far, the inter-stage total pressure and 

total temperature measurements collected from the seven-element rakes also provide valuable 

information regarding the radial distributions of measured flow properties. A direct comparison 

of these radial distributions for each of the three tip clearance configurations may be affected by 

the overall pressure rise (or temperature rise) through the compressor, therefore making it 

difficult to draw meaningful conclusions about the comparisons, especially near the rear of the 

compressor where discrepancies between tip clearance configurations may be large. As a result, 

these spanwise measurements must be normalized by some chosen value to accommodate 

comparison of the radial total pressure and total temperature profile shapes. The five selected 

loading conditions on the 100% corrected speedline have been chosen for comparison between 

the three tip clearance configurations. 

3.3.1. Total Pressure 

As defined in Section 3.1.3, the normalized radial total pressure, Equation (3.4), is again 

considered here for comparison of the pressure profiles through the compressor. Figure 3.19 

shows this normalized radial total pressure shape comparison at the LL operating condition for 

each of the three tip clearances, TC1-TC3, at axial measurement planes 3 through 8 (as denoted 

by the schematic in Figure 2.6). For reference, the uncertainty of measurements presented in 

Figure 3.19 is smaller than the symbol size in the figure. Using Equation (3.4), any reductions in 

pressure rise due to an increase of rotor tip clearance at the same axial measurement plane are 

avoided, and the radial stagnation pressure profile shapes remain for comparison. 

Figure 3.19: Radial total pressure profiles at LL. 

As the tip clearance height is increased in Figure 3.19, the tip leakage flow increases and 

the tip leakage flow disturbance grows in size and strength. This effect is not significant for the 

low loading (LL) condition, which is expected from the lack of distinct difference in compressor 

overall total pressure ratio between tip clearance heights for high flow rate conditions in Figure 

0.99 1 1.01

12

20

35

50

70

80

88

NRPo,3 [−]

Rotor 1 Exit

Annulus Height [%]

0.99 1 1.01

12

20

35

50

70

80

88

NRPo,4 [−]

Stator 1 Exit

0.99 1 1.01

12

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35

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70

80

88

NRPo,5 [−]

Rotor 2 Exit

0.99 1 1.01

12

20

35

50

70

80

88

NRPo,6 [−]

Stator 2 Exit

TC1

TC2

TC3

0.99 1 1.01

12

20

35

50

70

80

88

NRPo,7 [−]

Rotor 3 Exit

0.99 1 1.01

12

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35

50

70

80

88

NRPo,8 [−]

Stator 3 Exit

NASA/CR—2015-218868 60
3.1. However, this effect begins to become more apparent as the loading condition is increased 

up the speedline to the nominal loading (NL) and peak efficiency (PE) points in Figure 3.20 and 

Figure 3.21, respectively. In these figures, the increased blockage created by the tip leakage flow 

disturbance growth redistributes a portion of the flow from the tip region toward the hub region, 

as distinguished by the reduction of NRP3 at the tip and increase near the hub for all axial 

measurement positions. Also in these figures, it is possible to distinguish a set of similar profile 

shapes at the three rotor exit planes (axial positions 3, 5, and 7), which differ from a separate set 

of similar profiles shapes at the stator exit planes (axial positions 4, 6, and 8). 

Figure 3.20: Radial total pressure profiles at NL. 

Figure 3.21: Radial total pressure profiles at PE. 

0.99 1 1.01

12

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50

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88

NRPo,3 [−]

Rotor 1 Exit

Annulus Height [%]

0.99 1 1.01

12

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35

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70

80

88

NRPo,4 [−]

Stator 1 Exit

0.99 1 1.01

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88

NRPo,5 [−]

Rotor 2 Exit

0.99 1 1.01

12

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35

50

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88

NRPo,6 [−]

Stator 2 Exit

TC1

TC2

TC3

0.99 1 1.01

12

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50

70

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88

NRPo,7 [−]

Rotor 3 Exit

0.99 1 1.01

12

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88

NRPo,8 [−]

Stator 3 Exit

0.99 1 1.01

12

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88

NRPo,3 [−]

Rotor 1 Exit

Annulus Height [%]

0.99 1 1.01

12

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35

50

70

80

88

NRPo,4 [−]

Stator 1 Exit

0.99 1 1.01

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88

NRPo,5 [−]

Rotor 2 Exit

0.99 1 1.01

12

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88

NRPo,6 [−]

Stator 2 Exit

TC1

TC2

TC3

0.99 1 1.01

12

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35

50

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80

88

NRPo,7 [−]

Rotor 3 Exit

0.99 1 1.01

12

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88

NRPo,8 [−]

Stator 3 Exit

NASA/CR—2015-218868 61
As the loading condition is increased further up the speedline toward the stall point, these 

trends become more distinct. In particular, Figure 3.22 shows the same normalized radial 

pressure profile shapes at a high loading (HL) condition for the three tip clearance configurations. 

The same general trends of redistributed flow from the tip toward the hub with increased tip 

clearance persist in Figure 3.22, but there are more distinguishable characteristics between the 

axial positions. Of note, the difference between radial pressure profiles for the TC1 and TC3 

configurations at Stator 2 Exit are more pronounced than for Stator 1 Exit or Stator 2 Exit. 

Specifically, a corner separation condition exists at the hub of Stator 2, which is energized by the 

redistributed flow from the tip region toward the hub to prevent the flow separation tendencies of 

the vane row. This observation will be further explained with detailed measurements and flow 

visualization pictures later. 

Also in Figure 3.22, the tip region at Rotor 1 Exit portrays a repositioning of the minimum 

pressure from 88% span to 80% span as the tip leakage vortex core extends radially downward 

into the flow path with increased tip clearance. This leaning-over of the radial profile in the tip 

region for TC2 and TC3 is indicative of a close proximity to the stall point which has not yet 

been reached for TC1 at the same corrected mass flow rate (see Figure 3.1). However, the same 

distinguishing characteristic is not present at Rotor 2 Exit or Rotor 3 Exit, suggesting a 

difference in overall performance for the embedded stages, which ingest the increased blockage 

from the upstream stage(s), compared to the first stage which is always met with the same clean 

inlet flow from the IGV. 

Figure 3.22: Radial total pressure profiles at HL. 

As suggested earlier, the HL condition points compared in Figure 3.22 are located at 

proximities to the stall point which vary with tip clearance. Thus, an additional loading condition 

is considered which helps to overcome this inherent difference. The near-stall (NS) points 

denoted by the five-pointed stars in Figure 3.1 are representative of approximately 5% stall 

margin, as defined by Equation (3.2). By matching stall margin, the data comparison between the 

three tip clearance configurations should be more appropriate. 

0.99 1 1.01

12

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88

NRPo,3 [−]

Rotor 1 Exit

Annulus Height [%]

0.99 1 1.01

12

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70

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88

NRPo,4 [−]

Stator 1 Exit

0.99 1 1.01

12

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88

NRPo,5 [−]

Rotor 2 Exit

0.99 1 1.01

12

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50

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88

NRPo,6 [−]

Stator 2 Exit

TC1

TC2

TC3

0.99 1 1.01

12

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80

88

NRPo,7 [−]

Rotor 3 Exit

0.99 1 1.01

12

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NRPo,8 [−]

Stator 3 Exit

NASA/CR—2015-218868 62
This comparison, shown in Figure 3.23, yields some interesting results: the Rotor 1 Exit 

profile shape for TC1 has leaned over in the same manner as TC2 and TC3; also, the normalized 

radial pressure profile shapes for the three tip clearance configurations appear to coalesce at 

Rotor 2 Exit and Rotor 3 Exit. These data at Rotor 2 Exit and Rotor 3 exit show a very slight 

difference for the TC3 results, but this is explained by the fact that the data for TC3 actually 

represent a stall margin closer to 4% instead of the desired 5% which was achieved for TC1 and 

TC2. Therefore, the TC3 data are expected to achieve slightly less pressure near the tip and 

slightly higher pressure near the hub, as shown in Figure 3.23. Despite this apparent insensitivity 

of profile shape with respect to clearance size at Rotor 2 Exit and Rotor 3 Exit in Figure 3.23, 

there is a more noticeable difference at Rotor 1 Exit, and the discrepancy between TC1 and TC3 

at Stator 2 Exit is even more significant than in Figure 3.22. 

Figure 3.23: Radial total pressure profiles at NS. 

3.3.2. Total Temperature 

In the same fashion as the radial total pressure profiles, the normalized total temperature 

(NRT3) represents the circumferential average at each radial position, normalized by the area 

average (radial and circumferential) at the same axial position: 

NRT3,\ =

3,\,~

3,\,AA

. (3.10)

Figure 3.24 presents this normalized radial temperature distribution for the low loading (LL) 

condition on the 100% corrected speedline. At this operating point, the similarity of the pressure 

rise for the three tip clearance configurations leads to a largely insignificant difference between 

the three cases, with only minor differences downstream of each rotor row. For reference, the 

uncertainty of measurements presented in these figures is smaller than the symbol size in the 

figure. 

0.99 1 1.01

12

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70

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88

NRPo,3 [−]

Rotor 1 Exit

Annulus Height [%]

0.99 1 1.01

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88

NRPo,4 [−]

Stator 1 Exit

0.99 1 1.01

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88

NRPo,5 [−]

Rotor 2 Exit

0.99 1 1.01

12

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88

NRPo,6 [−]

Stator 2 Exit

TC1

TC2

TC3

0.99 1 1.01

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88

NRPo,7 [−]

Rotor 3 Exit

0.99 1 1.01

12

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88

NRPo,8 [−]

Stator 3 Exit

NASA/CR—2015-218868 63
Figure 3.24: Radial total temperature profiles at LL. 

As the loading of the compressor is increased, more significant differences appear in these 

normalized radial temperature profiles. Of note, the NL and PE operating points (Figure 3.25 and 

Figure 3.26, respectively) begin to show slight differences of the temperature distribution in the 

tip region at Rotor 1 Exit, denoting the difference of work distribution as the tip leakage flow 

affects the rotor performance. Also at these loading conditions, the Stator 1 Exit total 

temperature distributions are largely unaffected by the tip clearance changes, but Stator 2 Exit 

and Stator 3 Exit begin to show slight shape changes in these radial distributions. 

Figure 3.25: Radial total temperature profiles at NL. 

0.995 1 1.0051.01

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NRT

o,3

 [−]

Rotor 1 Exit

Annulus Height [%]

0.995 1 1.0051.01

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NRT

o,4

 [−]

Stator 1 Exit

0.995 1 1.0051.01

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NRT

o,5

 [−]

Rotor 2 Exit

0.995 1 1.0051.01

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NRT

o,6

 [−]

Stator 2 Exit

TC1

TC2

TC3

0.995 1 1.0051.01

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NRT

o,7

 [−]

Rotor 3 Exit

0.995 1 1.0051.01

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NRT

o,8

 [−]

Stator 3 Exit

0.995 1 1.0051.01

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NRT

o,3

 [−]

Rotor 1 Exit

Annulus Height [%]

0.995 1 1.0051.01

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NRT

o,4

 [−]

Stator 1 Exit

0.995 1 1.0051.01

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NRT

o,5

 [−]

Rotor 2 Exit

0.995 1 1.0051.01

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NRT

o,6

 [−]

Stator 2 Exit

TC1

TC2

TC3

0.995 1 1.0051.01

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88

NRT

o,7

 [−]

Rotor 3 Exit

0.995 1 1.0051.01

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NRT

o,8

 [−]

Stator 3 Exit

NASA/CR—2015-218868 64
Figure 3.26: Radial total temperature profiles at PE. 

As the compressor is further throttled toward stall, the HL (Figure 3.27) and NS (Figure 

3.28) normalized radial total temperature distributions show the differences at the rotor exit 

planes identified for the other loading conditions persist. The Stator 1 Exit profiles remain 

relatively unaffected by tip clearance changes across the entirety of the 100% corrected speedline. 

The Stator 2 Exit profiles are largely unchanged from the NL condition to the NS condition, with 

only slight differences at the LL condition. However, the Stator 3 Exit profiles highlight a slight 

difference in work distribution for the TC1 configurations compared to the larger tip clearances 

(TC2 and TC3) at the near stall (NS) operating condition. In particular, the TC1 configuration 

shows more work distributed in the hub region of Stator 3, compared to TC2 and TC3 which are 

nearly identical for this normalized total temperature distribution shape. 

0.995 1 1.0051.01

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NRT

o,3

 [−]

Rotor 1 Exit

Annulus Height [%]

0.995 1 1.0051.01

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NRT

o,4

 [−]

Stator 1 Exit

0.995 1 1.0051.01

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NRT

o,5

 [−]

Rotor 2 Exit

0.995 1 1.0051.01

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NRT

o,6

 [−]

Stator 2 Exit

TC1

TC2

TC3

0.995 1 1.0051.01

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NRT

o,7

 [−]

Rotor 3 Exit

0.995 1 1.0051.01

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NRT

o,8

 [−]

Stator 3 Exit

NASA/CR—2015-218868 65
Figure 3.27: Radial total temperature profiles at HL. 

Figure 3.28: Radial total temperature profiles at NS. 

3.4. Flow Visualization 

In addition to the quantitative measurement techniques presented so far, a flow 

visualization method was introduced to investigate the surface flow topologies on the stator 

vanes and evaluate regions of flow separation. These flow visualization tests were conducted at 

two loading conditions, NL and HL, for all three tip clearances, and an additional datasets were 

also collected at the peak efficiency and near stall operating points for the baseline tip clearance 

0.995 1 1.0051.01

12

20

35

50

70

80

88

NRT

o,3

 [−]

Rotor 1 Exit

Annulus Height [%]

0.995 1 1.0051.01

12

20

35

50

70

80

88

NRT

o,4

 [−]

Stator 1 Exit

0.995 1 1.0051.01

12

20

35

50

70

80

88

NRT

o,5

 [−]

Rotor 2 Exit

0.995 1 1.0051.01

12

20

35

50

70

80

88

NRT

o,6

 [−]

Stator 2 Exit

TC1

TC2

TC3

0.995 1 1.0051.01

12

20

35

50

70

80

88

NRT

o,7

 [−]

Rotor 3 Exit

0.995 1 1.0051.01

12

20

35

50

70

80

88

NRT

o,8

 [−]

Stator 3 Exit

0.995 1 1.0051.01

12

20

35

50

70

80

88

NRT

o,3

 [−]

Rotor 1 Exit

Annulus Height [%]

0.995 1 1.0051.01

12

20

35

50

70

80

88

NRT

o,4

 [−]

Stator 1 Exit

0.995 1 1.0051.01

12

20

35

50

70

80

88

NRT

o,5

 [−]

Rotor 2 Exit

0.995 1 1.0051.01

12

20

35

50

70

80

88

NRT

o,6

 [−]

Stator 2 Exit

TC1

TC2

TC3

0.995 1 1.0051.01

12

20

35

50

70

80

88

NRT

o,7

 [−]

Rotor 3 Exit

0.995 1 1.0051.01

12

20

35

50

70

80

88

NRT

o,8

 [−]

Stator 3 Exit

NASA/CR—2015-218868 66
(TC1). All nominal loading tests feature orange paint, while blue paint was used for the peak 

efficiency and high loading conditions. A combination of paint colors was used at the near stall 

condition. Further description of the measurement technique can be found in Smith and Key 

(2015). 

Figure 3.29(a) shows an example image of a stator with corner separations at both the hub 

and tip endwall junctions on the suction side of a Stator 1 vane. Flow goes from left to right in 

the figure; the stator hub is at the bottom of the image and the tip is at the top as noted. This vane 

is positioned at the compressor casing splitline, and therefore may be imaged straight-on, 

whereas all later images presented were acquired from the trailing edge looking upstream. The 

corner separations are marked in Figure 3.29(b) and appear as semi-triangular patterns. For this 

Stator 1 vane at the high loading condition, the tip has a larger corner separation than the hub, as 

a result of many design factors including but not limited to flow angles, diffusion factor, and 

Rotor 1 performance. The streaklines in the separated regions reveal the recirculation patterns in 

the flow, as sketched in Figure 3.29(c). 

Several critical points are identifiable in the surface flow topology. At least four saddle 

points (S) and four nodes (N) are present, which follows the rule given by Flegg (1974) that the 

number of saddle points and nodes must be equal for a vane row with no tip gap. It should be 

noted that there are regions that may have indistinguishable critical points due to resolution of 

paint or flow pattern complexities, like the region where the tip and hub meet and near the 

endwalls. When comparing many Stator 1 vane images from the same dataset, there is another 

possible node location near the saddle point at the tip. This type of detailed analysis of the 

surface flow topology is useful to comparisons with computational models, particularly with 

respect to the acceptability of turbulence models. 

Figure 3.29: Basic surface flow topology for Stator 1 at the high loading condition (a) image, 

(b) corner separations, and (c) drawn flow topology with saddle points (S) and nodes (N) labeled.

NASA/CR—2015-218868 67
Figure 3.30 provides an overview of the suction side surface flow visualization images on 

Stator 1, Stator 2, and Stator 3 (left to right) at the four loading conditions (increasing loading 

top to bottom) for TC1. The image of Stator 1 at peak efficiency loading is labeled, indicating 

the vane hub, tip, leading edge (LE) and trailing edge (TE). This orientation remains the same for 

all following images. On the suction surface of each stator, the regions with paint surface flow 

patterns are regions with recirculating flow and separations. The corner separations increase in 

size further downstream through the compressor, where Stator 3 has larger separations than 

Stator 1. The flow conditions in the rear stages contain increased turbulence and more unsteady 

disturbances due to wakes from upstream blade rows resulting in earlier vane boundary layer 

separation. Furthermore, as loading increases, the vanes experience a stronger adverse pressure 

gradient making the boundary layer more susceptible to separations, and thus, the corner 

separations grow in size. 

At the nominal loading (NL) condition, Figure 3.30(a-c), all three stators have larger 

recirculation regions at the tip compared to the respective hub regions. The amount of separated 

area increases with each downstream stator, though Stator 1 and Stator 2 are quite similar. These 

trends fall within expectations of this relatively low loss operating condition, but are also specific 

to the design of the compressor. When loading is increased from nominal loading to the peak 

efficiency loading condition, this trend changes and each stator’s separation pattern is unique. 

Changes in the radial distribution of the flow with increased loading are specific to the 

compressor design, and are governed by various aspects of stage matching (Cumpsty 2004). The 

flow patterns on Stator 1 are similar at nominal and peak efficiency conditions, whereas the 

Stator 2 and Stator 3 hub corner separations grow more and both have boundary layer separation 

along the entire span. 

At peak efficiency, the Stator 2 tip still has a larger separated region than the hub, while the 

Stator 3 hub and tip separations are nearly equivalent. For this PE case, shown in Figure 3.30(df), Stator 1 experiences similar inlet conditions as in Figure 3.30(a) at NL, but the effects of 

increased loading, stronger wakes and adverse pressure gradient are more substantial to the inlet 

conditions for Stator 2 and Stator 3. All three vanes have large increases in corner separation size 

at the hub and tip between the peak efficiency point and the high loading condition (Figure 

3.30(g-i)), which is consistent with the increased losses expected at high loading conditions. By 

the point at which the high loading operating condition is reached, Stator 1 shows full spanwise 

boundary layer separation, and Stator 2 and Stator 3 have separations extending much further 

upstream on the vane surface. The final set of images, Figure 3.30(j-l), are at the near stall 

loading condition. The flow visualization at NS shows that all three stators have significant 

suction side boundary layer separation. The tip region (approximately 75% span and above) of 

Stator 1 is completely separated, beginning from the leading edge. This is difficult to see in 

Figure 3.30(j), but can be identified more easily in Figure 3.31, looking downstream at the 

leading edge. 

NASA/CR—2015-218868 68
Figure 3.30: Surface flow visualization of corner separations for Stator 1 (left), Stator 2 (middle), 

and Stator 3 (right) at four loading conditions: NL (a-c), PE (d-f), HL (g-i), and NS (j-l) for TC1.

NASA/CR—2015-218868 69
Figure 3.31: Stator 1 at NS conditions viewed from the leading edge looking downstream. 

The surface flow patterns in Figure 3.30 have several identifiable nodal points and saddle 

points in the surface topology of all three stators. Specifically, the spiral node in the tip 

recirculation region at the trailing edge of Stator 3 is visible in all three loading conditions. For 

the nominal loading and peak efficiency conditions, it is similar in size and is located at 60% 

span, but at the high loading condition, the spiral node grows in size and moves toward the hub 

to approximately midspan. 

Increased rotor tip clearance causes larger tip leakage flows, thereby resulting in additional 

blockage near the tip endwall at the inlet of each stator. This blockage redistributes the flow 

causing more fluid to travel through the lower half of the vane span. This locally increases the 

mass flow rate and unloads the hub, resulting in reduced corner separations at the hub. 

Meanwhile the large blockage at the tip due to the increased tip leakage flow disturbance 

weakens the tip and increases the amount of separation on the downstream vane. These trends 

can be observed on all three stators and are presented for NL and HL, at the three rotor tip 

clearances TC1, TC2, and TC3. 

Figure 3.32 shows surface flow patterns at NL for all three stators and increasing rotor tip 

clearance, with TC1 at the top of the figure and TC3 at the bottom. Even at this relatively low 

loss loading condition, the increase in size of the rotor tip leakage flow disturbance is significant 

enough to shift the flow in the downstream stator from the tip to the hub. All three stators 

experience an increase in loss at the tip as rotor tip clearance increases, resulting in an apparent 

increase of tip corner separations. The increased size of the tip corner separation is particularly 

apparent between the TC1 and TC2 cases, for which the tip corner separation begins further 

upstream along the vane suction surface and extends deeper along the spanwise direction. The 

hub corner separations of Stator 1 and Stator 2 reduce in size from TC1 to TC2, but their size 

does not appear to reduce further when the clearance is increased to TC3. In fact, the Stator 3 

hub corner separation remains similar in size for all three rotor tip clearances. Based on these 

observations, the increase in rotor tip clearance has a larger effect from TC1 to TC2 at NL. The 

growth in blockage from the tip leakage flow disturbance from TC2 to TC3 does not 

significantly change the radial distribution of the flow. 

NASA/CR—2015-218868 70
Figure 3.32: Flow visualization of vane corner separations at NL for TC1 (top), 

TC2 (middle), and TC3 (bottom). 

The high loading (HL) operating condition is presented in Figure 3.33 using the same 

format for three tip clearance configurations. This HL operating condition is further from the 

design point of the compressor and shows more loss compared to the nominal loading (NL) 

condition. As a result, the rotor tip clearance effects on stator loss are more dramatic. The higher 

pressure ratios increases the leakage flow through the clearance gap and results in larger stator 

separation regions. The amount of boundary layer separation on Stator 1 is nearly unchanged 

between TC1 and TC2, but there is evidence of a stronger rotor tip leakage flow disturbance in 

the streamlines near the tip. Of note, TC2 has significant radial shift of streamlines at the top 14% 

span of the separated region. When the rotor tip clearance is increased to TC3, the effects of 

NASA/CR—2015-218868 71
increased flow blockage at the tip are far more significant. The stator tip corner separation 

originates from a position near the vane leading edge, and the radial streamlines extend down 

through the outer 22% span. In contrast, the hub region has been strengthened by the 

redistributed flow, resulting in a smaller hub corner separation. 

The surface topology of Stator 2 changes the most of the three stator vane for the high 

loading (HL) condition in Figure 3.33. At the baseline tip clearance (TC1), there is a large spiral 

node in the hub corner separation, the tip region has a smaller node, and the two separations 

intersect at about 42% span. As the rotor tip clearance is increased to TC2, the fluid is 

redistributed through the Stator 2 passage due to the increased blockage from the Rotor 2 tip 

leakage flow disturbance, as evidenced by the shifts of the two nodes. Also for this TC2 tip 

clearance configuration, the spiral node at the hub is reduced in size and originates further 

downstream in comparison to TC1. The small node at the tip is less defined and is surrounded by 

more radial streamlines. The point where the two corner separations merge has moved toward 

the hub to about 37% span. Finally, at the largest tip clearance (TC3), the spiral node at the hub 

of Stator 2 is smaller and only extends to about 23% span. For this tip clearance condition, the 

Stator 2 surface flow near the tip region shows strong radial streamlines which extend down to 

about 72% span. 

Similar to both Stator 1 and Stator 2, Stator 3 experiences a shift in radial location of losses 

with increased tip clearance. The hub corner separation decreases in size, indicating the hub is 

strengthened as the tip clearance increases. Furthermore, the tip region contains large amounts of 

blockage with significant radial flow apparent in the surface streaklines. 

NASA/CR—2015-218868 72
Figure 3.33: Flow visualization of vane corner separations at HL for TC1 (top), 

TC2 (middle), and TC3 (bottom). 

NASA/CR—2015-218868 73
3.5. Stator Wake Measurements 

3.5.1. Steady Pressure and Temperature Rake Data 

In addition to the circumferentially-averaged data presented in Section 3.3, the discrete 

measurements from the total pressure and total temperature rakes have also been considered in 

the circumferential direction across one vane pitch at the exit of each stator vane row. Beginning 

with data collected at Stator 1 exit, Figure 3.34 highlights measurements collected at three of the 

radial measurement positions: 50%, 70%, and 88% annulus height. These data have been 

normalized by the circumferential average of the data collected at 50% annulus height for each 

of the tip clearances. 

At each of the three loading conditions in Figure 3.34, the Stator 1 total pressure wake 

shapes at 50% and 70% annulus height are relatively unchanged with tip clearance height. 

However, there are slight variations in the measurements collected at the 88% span position, 

where the flow is most directly affected by the tip leakage flow disturbance. At the low loading 

(LL) condition, the suction side of the wake increases in width at 70% and 88% annulus height. 

The Stator 1 Exit wake shapes at nominal loading (NL) appear unaffected by rotor tip 

clearance height changes at 50% annulus height. The wake depths and widths at 70% are 

unchanged, but there is a slight decrease of pressure rise off the pressure surface edge of the 

vane – an effect which is greatest between the TC2 and TC3 tip clearance configurations. At this 

loading condition, however, the greatest effect is at 88% annulus height, where the change of 

rotor tip clearance from TC1 to TC3 shows a decrease of the normalized wake shape off the 

suction surface (near 40% vane passage) on the order of one percent. 

At the high loading (HL) condition, the Stator 1 data in Figure 3.34 also show the thickness 

and minimum normalized pressure of the wake regions are unaffected by the change of rotor tip 

clearance. However, the pressure in the primary throughflow at 88% annulus height shows a 

constant offset in the section off the suction surface of the vane. The one percent change 

observed in the small circumferential region from 30 to 50% vane passage at NL extends across 

the rest of the vane passage at the HL condition. 

The normalized total temperature behind Stator 1 in Figure 3.34 shows the accumulation of 

high total temperature fluid off the pressure surface of the vanes. This region represents the fluid 

from the rotor wake which collects on the stator pressure surface, as described by Kerrebrock 

and Mikolajczak (1970). This same effect has been previously documented at the baseline tip 

clearance, TC1, by Key (2014). At the LL and NL conditions, Figure 3.34 shows no discernable 

change of the normalized circumferential total temperature distribution. However, the HL 

condition shows an increase of the peak normalized total temperature in the region off the vane 

pressure surface (associated with rotor wake fluid) with increased rotor tip clearance – the most 

significant of which is observed at 70% annulus height. 

NASA/CR—2015-218868 74
Figure 3.34: Stator 1 total pressure wakes and pitchwise total temperature distribution. 

0 20 40 60 80 100

0.950 0.970 0.990 1.010 

Normalized Po,4 / Po,AIPSS PS

LL

0 20 40 60 80 100

0.996 1.000 1.004 1.008 1.012 

S1 Vane Passage [%]

Normalized T

o,4

 / T

o,AIP

0 20 40 60 80 100

0.950 0.970 0.990 1.010 

Normalized Po,4 / Po,AIP

NL

SS PS

50% Span70% Span88% Span

0 20 40 60 80 100

0.996 1.000 1.004 1.008 1.012 

S1 Vane Passage [%]

Normalized T

o,4

 / T

o,AIP

TC1

TC2TC3

0 20 4

0.950 0.970 0.990 1.010 

Normalized Po,4 / Po,AIPS

0 20 4

0.996 1.000 1.004 1.008 1.012 

S1 Van

Normalized T

o,4

 / T

o,AIP

NASA/CR—2015-218868 75
Measurements comparing total pressure and total temperature rake data are also included at 

Stator 2 exit for three rotor tip clearance configurations and three loading conditions, Figure 3.35. 

Whereas the Stator 1 exit results showed minimal effect at the LL operating condition, Stator 2 

displays a more significant impact at the 88% measurement location: a nearly constant pressure 

across the bulk throughflow at 88% for TC1 develops into a pressure trough extending 

approximately 60% vane passage at a maximum depth of approximately two percent with respect 

to the bulk throughflow. The total temperature variations observed at HL for Stator 1 exit data 

are not present for the Stator 2 exit data. 

The circumferential variations of Stator 3 exit pressures and temperature are shown in 

Figure 3.36. As with Stator 2, the most distinct effect of the pressure wake shapes with rotor tip 

clearance changes is observed at the low loading (LL) condition. At this condition, the increase 

from TC1 to TC2 creates a depression on the order of one percent with respect to the bulk 

throughflow, in a circumferential region extending across the majority of the throughflow area. 

In contrast, the increase from TC2 to TC3 shows the same depression depth, but the location of 

minimum pressure has moved approximately 20% vane passage toward the vane pressure surface. 

Similar to the other stators, the Stator 3 results in Figure 3.36 show circumferential total 

temperature distributions which are largely unaffected by increased rotor tip clearance at LL and 

NL. At the high loading (HL) condition, there is an increase of normalized total temperature 

across the entire circumference at 88% span; this observation is in contrast to the slight increase 

of the peak rotor fluid accumulation which was observed only near the pressure surface of Stator 

1 (and most significantly at 70% span), and the Stator 2 measurements which were largely 

unaffected by rotor tip clearance. 

NASA/CR—2015-218868 76
Figure 3.35: Stator 2 total pressure wakes and pitchwise total temperature distribution. 

0 20 40 60 80 100

0.950 0.970 0.990 1.010 

Normalized Po,6 / Po,AIPSS PS

LL

0 20 40 60 80 100

0.996 1.000 1.004 1.008 1.012 

S2 Vane Passage [%]

Normalized T

o,6

 / T

o,AIP

0 20 40 60 80 100

0.950 0.970 0.990 1.010 

Normalized Po,6 / Po,AIP

NL

SS PS

50% Span70% Span88% Span

0 20 40 60 80 100

0.996 1.000 1.004 1.008 1.012 

S2 Vane Passage [%]

Normalized T

o,6

 / T

o,AIP

TC1

TC2TC3

0 20 4

0.950 0.970 0.990 1.010 

Normalized Po,6 / Po,AIPS

0 20 4

0.996 1.000 1.004 1.008 1.012 

S2 Van

Normalized T

o,6

 / T

o,AIP

NASA/CR—2015-218868 77
Figure 3.36: Stator 3 total pressure wakes and pitchwise total temperature distribution. 

0 20 40 60 80 100

0.950 0.970 0.990 1.010 

Normalized Po,8 / Po,AIPSS PS

LL

0 20 40 60 80 100

0.996 1.000 1.004 1.008 1.012 

S3 Vane Passage [%]

Normalized T

o,8

 / T

o,AIP

0 20 40 60 80 100

0.950 0.970 0.990 1.010 

Normalized Po,8 / Po,AIP

NL

SS PS

50% Span70% Span88% Span

0 20 40 60 80 100

0.996 1.000 1.004 1.008 1.012 

S3 Vane Passage [%]

Normalized T

o,8

 / T

o,AIP

TC1

TC2TC3

0 20 4

0.950 0.970 0.990 1.010 

Normalized Po,8 / Po,AIPSS 

0 20 4

0.996 1.000 1.004 1.008 1.012 

S3 Van

Normalized T

o,8

 / T

o,AIP

NASA/CR—2015-218868 78
3.5.2. Detailed Steady Total Pressure Traverses 

To supplement the measurements collected downstream of the stator vanes using the total 

pressure rakes, a series of detailed traverses were conducted with a miniature Kiel head total 

pressure probe. These measurements consisted of up to 30 radial plunge positions (at least four 

times the resolutions from the rake), with an emphasis on capturing the flow features in the 

endwall regions, as well as a higher resolution of 25 pitch-wise positions with respect to the 

stationary vanes. Due to the long-duration runs required for this measurement technique, these 

detailed measurement campaigns were conducted at the exit of each of the three stator vane rows, 

at two loading conditions (NL and HL), for the smallest and largest tip clearance configurations 

(TC1 and TC3). 

As alluded to by the radial profiles in Section 3.3.1, an increase rotor tip clearance height 

has the effect of producing increased blockage due to the leakage flow disturbance in the outer 

regions of the annulus. This corresponding redistribution of mass flow from the tip region toward 

the hub region unloads the hub region of the downstream stator to reduce the wake thickness in 

the lower region of the annulus and reduce the corner separation regions in the hub corner, as 

shown in Section 3.4. However, the reduced flow in the tip region leads to increased flow 

separation at the outer diameter of the stator vane. For measurements collected at Stator 1 exit 

for a nominal loading (NL) condition, Figure 3.37 shows this slight decrease of wake thickness 

at approximately 30% annulus height as the rotor tip clearance is increased from TC1 to TC3. In 

this same figure, the reduction of flow (and corresponding reduced total pressure) in the tip 

region is also apparent due the presence of the leakage flow from Rotor 1. 

Figure 3.37: Detailed total pressure traverses at Stator 1 exit, NL. 

Vane Passage [%]

TC1

Annulus Height [%]

0 20 40 60 80 100

0

10

20

30

40

50

60

70

80

90

100

Vane Passage [%]

TC3

0 20 40 60 80 100

Po,4 / Po,AIP,RA

1.040 

1.049 

1.057 

1.066 

1.074 

1.083 

1.091 

1.100 

NASA/CR—2015-218868 79
Figure 3.38: Detailed total pressure traverses at Stator 1 exit, HL. 

For the same Stator 1 exit measurement plane, but at a high loading (HL) operating 

condition, Figure 3.38 portrays this same, but more noticeable effect. At this position closer to 

the stall point, the lower 30% of the annulus shows a benefit from the increased tip clearance, but 

the outer 70% annulus height presents a significant increase of stator wake width and depth. In 

this figure, the bulk flow outside the wake also shows a significant decrease of total pressure rise 

extending down to approximately 70% annulus height. 

Similar comparisons can also be made between two tip clearance configurations at the exit 

of Stator 2. At the NL operating condition, Figure 3.39 shows a significant decrease of total 

pressure in the tip region as the increased tip leakage disturbance from Rotor 2 is ingested by the 

stator vane. The compounding effect of the increased tip clearance height is displayed by this 

embedded stage, making it more difficult to directly compare contour levels between the two tip 

clearance configurations. However, the effect of the increased blockage appears to affect the 

flow most significantly in the outer 20% of the annulus height. 

At the high loading (HL) condition at Stator 2 exit, Figure 3.40, the most noticeable effect 

of the increased tip clearance height is observed. Referring back to the radial total pressure 

profiles at this same operating condition, Figure 3.22, a distinct shape difference is observed for 

the change from TC1 to TC3. The data in Figure 3.40 show that TC1 exhibits a significant hub 

corner separation at this high loading condition which extends up to approximately 30% annulus 

height; however, the wake thickness remains relatively constant across the remainder of the 

annulus. As the rotor tip clearance height is increased, the redistribution of mass flow from the 

tip toward the hub region has the expected result of energizing the hub region and decreasing the 

likelihood of flow separation, thereby reducing the hub corner separation zone. However, this 

benefit comes at the expense of a drastic increase of wake thickness, as observed for the upper 50% 

annulus height. 

Vane Passage [%]

TC1

Annulus Height [%]

0 20 40 60 80 100

0

10

20

30

40

50

60

70

80

90

100

Vane Passage [%]

TC3

0 20 40 60 80 100

Po,4 / Po,AIP,RA

1.070 

1.077 

1.084 

1.091 

1.099 

1.106 

1.113 

1.120 

NASA/CR—2015-218868 80
Figure 3.39: Detailed total pressure traverses at Stator 2 exit, NL. 

Figure 3.40: Detailed total pressure traverses at Stator 2 exit, HL. 

The detailed total pressure measurements collected at Stator 3 exit are also shown for the 

same NL and HL operating conditions on the 100% corrected speedline in Figure 3.41 and 

Figure 3.42, respectively. At the high loading condition, Figure 3.42, the TC1 rotor tip clearance 

configuration exhibits the thickest wake of all three stators (relatively), but with the shallowest 

Vane Passage [%]

TC1

Annulus Height [%]

0 20 40 60 80 100

0

10

20

30

40

50

60

70

80

90

100

Vane Passage [%]

TC3

0 20 40 60 80 100

Po,6 / Po,AIP,RA

1.150 

1.157 

1.164 

1.171 

1.179 

1.186 

1.193 

1.200 

Vane Passage [%]

TC1

Annulus Height [%]

0 20 40 60 80 100

0

10

20

30

40

50

60

70

80

90

100

Vane Passage [%]

TC3

0 20 40 60 80 100

Po,6 / Po,AIP,RA

1.180 

1.187 

1.194 

1.201 

1.209 

1.216 

1.223 

1.230 

NASA/CR—2015-218868 81
depth (see also Figure 3.36). As the tip clearance is increased to TC3, the observed effect is the 

same for Stator 3 as it was for the other two vane rows: the hub region (up to approximately 30% 

annulus height) shows a decreased wake thickness, but the outer 70% annulus height is 

negatively affected by the decreased flow in the tip region that has been redistributed toward the 

hub. 

In addition to the high resolution of measurements, the increased quantity of data points 

from these detailed radial traverses provide the ability to compare with the calculated total 

pressure ratio determined from the seven-element total pressure rakes. On average, the total 

pressure ratios calculated as an area average of the detailed radial traverse data agree with the 

values calculated from the pressure rakes within 0.04%, less than one-fourth of the uncertainty 

for the total pressure ratio. 

Figure 3.41: Detailed total pressure traverses at Stator 3 exit, NL. 

Vane Passage [%]

TC1

Annulus Height [%]

0 20 40 60 80 100

0

10

20

30

40

50

60

70

80

90

100

Vane Passage [%]

TC3

0 20 40 60 80 100

Po,8 / Po,AIP,RA

1.270 

1.276 

1.281 

1.287 

1.293 

1.299 

1.304 

1.310 

NASA/CR—2015-218868 82
Figure 3.42: Detailed total pressure traverses at Stator 3 exit, HL. 

Vane Passage [%]

TC1

Annulus Height [%]

0 20 40 60 80 100

0

10

20

30

40

50

60

70

80

90

100

Vane Passage [%]

TC3

0 20 40 60 80 100

Po,8 / Po,AIP,RA

1.300 

1.309 

1.317 

1.326 

1.334 

1.343 

1.351 

1.360 

NASA/CR—2015-218868 83
CHAPTER 4: CHARACTERIAZATION OF COMPRESSOR STALL 

INCEPTION 

Reid A. Berdanier, Natalie R. Smith, Anna M. Young, and Nicole L. Key 

4.1. Measurement Technique 

For this study, the stall inception mechanism for the compressor was systematically 

assessed for each of the three tip clearance configurations. In addition to the four corrected 

speedlines shown in Figure 3.1, the compressor stall behavior was also investigated for the 60% 

and 52% corrected speedlines. In each case, the stall point was determined by closing the throttle 

in incremental steps to slowly increase the loading of the compressor. At the onset of stall, a 

human-interface control mechanism was initiated to open the throttle and allow recovery to a 

stable operating condition. This process was repeated several times to determine a representative 

average of corrected mass flow rate and total pressure ratio for each test case. The previous 

statements regarding uncertainty of mass flow rate and overall total pressure ratio apply for these 

measurements as well. For a given operating speed and tip clearance configuration, the 

compressor was also stalled in up to five different vane positions with respect to the stationary 

instrumentation to develop a representative average for the overall map. During the stall 

inception tests, the full set of inter-stage measurement rakes was removed and only the inlet and 

exit conditions of the compressor (axial measurement planes 0, 1, and 9 in Figure 2.6) were 

monitored for overall pressure rise information. 

To detect the formation of stall events in the compressor, a series of high-frequency 

response Kulite XTL-140 pressure transducers were installed throughout the compressor. These 

piezoresistive pressure transducers each have an outer diameter of 0.101 in. with a threaded 

mounting system to accommodate insertion and removal at several unique positions. As with the 

other fast-response pressure transducers used in this study, these XTL-140 sensors also featured 

a standard protective B-screen, reducing the natural frequency of the sensor to approximately 20 

kHz. For these stall tests, six circumferentially-distributed sensors were positioned 

approximately 15% axial chord upstream of each rotor blade row, following the design outlined 

by Houghton and Day (2010), for 18 total sensors. The measurements from these sensors were 

collected through the same data acquisition chain described in Section 2.5.1. 

4.2. Stall Inception Analysis 

Over the years, researchers (McDougall et al., 1990; Day, 1993) have identified two 

different stall inception mechanisms in axial compressors. The first type, modal oscillations, or 

“modes,” represents long length-scale disturbances which affect the entire compressor. These 

modal oscillations are often observed for some time prior to the definitive stall point for the 

compressor. The second type, “spikes,” appears as a short length-scale disturbance which 

develops as a result of a localized stalling, typically in one blade row. These spike signals may 

appear in high frequency response data collected from hot-wire velocity measurements or 

dynamic pressure transducers. 

Camp and Day (1998) continued the discussion of these stall inception types and developed 

a model for predicting the stall inception mechanism of a given compressor, based on the shape 

NASA/CR—2015-218868 84
of the total-to-static pressure rise characteristic for the machine. These authors concluded modal 

stall inception characteristics are expected to appear if the compressor reaches a peak total-tostatic pressure rise prior to entering stall (i.e., a zero slope condition). On the other hand, if the 

compressor stalls prior to reaching a peak total-to-static pressure rise value, the corresponding 

stall inception mechanism will likely be a spike-type stall event. 

The total-to-static pressure rise characteristics are shown for the Purdue three-stage 

compressor, with each of three tip clearance configurations at four operating speeds, in Figure 

3.10. The characteristics in this figure show that there are three lines which potentially show a 

peak value has been reached: the 100% corrected speedlines for TC1 and TC2, and the 90% 

corrected speedline for TC1. Based on the model presented by Camp and Day (1998), it is 

expected that these three speedlines are the most likely candidates to portray modal stall 

inception traits. The other nine speedlines all appear to be continuing to rise at the stall point (the 

lowest flow rate point on each line) – typical signs of spike-type stall inception. 

Beyond the predictive capabilities of the characteristics in Figure 3.10, dynamic static 

pressure traces at several of these operating conditions are presented. First, measurements 

collected with the baseline tip clearance (TC1) are shown for the 100% corrected speedline, 

Figure 4.1. This figure presents the offset dynamic pressure traces collected at each of the 

eighteen locations around the compressor (six circumferentially-distributed sensors at each rotor). 

As expected from the total-to-static characteristic, this speed and tip clearance configuration 

exhibits modal stall inception traits. Of note, the long length-scale disturbances are observed for 

several rotor revolutions prior to the onset of stall (approximately identified by the zero location 

on the abscissa). Tracking these “modal” oscillations through the compressor shows the mode 

speed is approximately one-fourth of the rotor rotational speed. At the onset of stall, the stall cell 

rotates at approximately one-half the rotor rotational speed. 

Figure 4.1: Dynamic pressure traces at the inception of stall for TC1 at 100% .. 

Day (1993) showed that a change of rotor tip clearance may affect the stall inception 

mechanism of the machine. As previously discussed, the 100% corrected speed for the 

intermediate clearance height (TC2) also suggests the existence of modal stall. The dynamic 

static pressure traces for this speed and tip clearance configuration are presented in Figure 4.2. 

This figure may show weak long length-scale modal propagations (also at a rotational rate on the 

−15 −10 −5 0 5 10

 R1 

 R2 

 R3 

Time / Rotor Revolution

Nominal Offset Pressure

NASA/CR—2015-218868 85
order of one-fourth the rotor speed) in the measurements upstream of Rotor 1, as suggested from 

the total-to-static pressure rise characteristic. However, the distinct modal traits from Figure 4.1 

are certainly less apparent (and nearly absent in the Rotor 2 or Rotor 3 data). Further, some 

spikes can also be identified in the Rotor 1 pressure traces. Of note, one spike develops in the 

first sensor upstream of Rotor 1 at eight revolutions prior to the onset of stall, and it rotates 

approximately at the rotor rotational speed. The traces in Figure 4.2 confirm Camp and Day’s 

(1998) observations that these two stall inception mechanisms can occur simultaneously in the 

same machine. In this case, however, Figure 4.2 appears to be dominated more by spikes than the 

long length-scale modal fluctuations. 

Figure 4.2: Dynamic pressure traces at the inception of stall for TC2 at 100% .. 

Hoying et al. (1999) has previously identified that the trajectory of the tip leakage flow 

approaches a tangential angle as the blade loading increases, influencing the potential for an 

unstable condition in which the leakage flow may jump upstream of the adjacent rotor blade and 

cause spike-type stall behavior. This diagnosis from Hoying et al. would predict that the TC3 tip 

clearance configuration (and possibly also TC2) may display a near-tangential leakage flow 

trajectory for the near stall operating conditions on the 100% corrected speedline (based on the 

spikes observed emanating from the Rotor 1 dynamic pressure traces in Figure 4.2, e.g.), 

whereas the TC1 trajectory would be less inclined due to its modal stall tendencies (Figure 4.1). 

To investigate this further, dynamic static pressures were measured over Rotor 1 using the 

removal instrumentation block described in Section 2.5.1. These data were collected at the near 

stall (NS) operating condition, defined by a position of 5% stall margin using Equation (3.2), as 

explained in Section 3.1.1. The results are analyzed in terms of root-mean-square (RMS) 

unsteadiness with respect to the ensemble average: 

RMS>\

A = Ñ

1



ÖÜ>\

A − á>\Aàâ

;

ä

ã

äå=

,

(4.1)

−15 −10 −5 0 5 10

 R1 

 R2 

 R3 

Time / Rotor Revolution

Nominal Offset Pressure

NASA/CR—2015-218868 86
for which the phase-locked ensemble average (EA) is defined by: 

á>\Aà =

1



ÖÜ>\

Aâä

ã

äå=

. (4.2)

These NS results, Figure 4.3, show this predicted result by considering the flow field 

unsteadiness in terms of normalized RMS pressures. 

More recent studies continuing the evaluation of spike-type stall inception suggested by 

Hoying et al. (1999) have suggested the development of spike-type stall may not be strictly 

dependent on the leakage flow, but rather a radial vortex which develops as a result of leading 

edge separation (Weichert and Day, 2014; Pullan et al., 2015). It is this radial vortex structure 

which propagates in front of the adjacent blade and leads to the spike. However, the increased 

incidence at the tip due to the leakage flow, especially with larger tip clearance heights, as shown 

in Figure 4.3, is expected to increase the likelihood of such leading edge separation. 

Figure 4.3: Static pressure field unsteadiness over Rotor 1, presented as an RMS with respect to 

the EA, for each tip clearance configuration at NS. 

Considering again the total-to-static pressure rise characteristics for the baseline tip 

clearance configuration (TC1), Figure 3.10, the 90% corrected speedline also appears to reach a 

maximum and begin to decrease – a sign of potential modal stall inception traits. The dynamic 

pressure traces for this case, Figure 4.4, reveal results which are more similar to TC1 at 100% 

corrected speed than TC2. In particular, Figure 4.4 presents long length-scale modal tendencies 

which persist through all three rotor rows. However, the spike-type signals are still present in the 

Rotor 1 pressure traces. Of note, there are spikes propagating from the first sensor of Rotor 1 at 

13, 11, and five revolutions prior to the onset of stall, all of which fall in the troughs of the modal 

waves. Relative to Figure 4.2, these results for TC1 at 90% corrected sped represent more modal 

stall inception traits than spikes; however, both mechanisms are certainly present for this 

condition as well. 

Axial Chord [%]

t / BPP [−]

TC1

−20 0 20 40 60 80 100 120

0

1

2

PRMS/Pamb

0.002 

0.004 

0.006 

0.008 

0.010 

0.012 

0.014 

0.016 

0.018 

0.020 

0.022 

Axial Chord [%]

t / BPP [−]

TC2

−20 0 20 40 60 80 100 120

0

1

2

PRMS/Pamb

0.002 

0.004 

0.006 

0.008 

0.010 

0.012 

0.014 

0.016 

0.018 

0.020 

0.022 

Axial Chord [%]

t / BPP [−]

TC3

−20 0 20 40 60 80 100 120

0

1

2

PRMS/Pamb

0.002 

0.004 

0.006 

0.008 

0.010 

0.012 

0.014 

0.016 

0.018 

0.020 

0.022 

NASA/CR—2015-218868 87
Figure 4.4: Dynamic pressure traces at the inception of stall for TC1 at 90% .

. 

In contrast with the results presented thus far, the part-speed results for TC1 (68% 

corrected speed) are shown in Figure 4.5. These dynamic pressure traces are dominated by spiketype stall inception traits propagating from Rotor 1. These spikes travel circumferentially around 

Rotor 1 at a rate approximately equal to the rotor rotational speed, and they are likely the result 

of localized separation regions related to high incidence angles for Rotor 1. However, it is 

interesting to note in this case that it appears to be Rotor 2 which first trips the compressor into 

stall, suggesting that the rear stages of the compressor could be attempting to “assist” Rotor 1 

until they are no longer capable of doing so. The data in Figure 4.5 are shown on the same scale 

as the previous 100% and 90% speedline traces for consistency. Although it may not be clear 

from the scale, if the results in Figure 4.5 were shown on a larger scale, modal stall 

characteristics would still not be present. 

Figure 4.5: Dynamic pressure traces at the inception of stall for TC1 at 68% .. 

−15 −10 −5 0 5 10

 R1 

 R2 

 R3 

Time / Rotor Revolution

Nominal Offset Pressure

−15 −10 −5 0 5 10

 R1 

 R2 

 R3 

Time / Rotor Revolution

Nominal Offset Pressure

NASA/CR—2015-218868 88
Although not all of the test cases have been presented here, these four selected results 

represent the observed trends in the compressor at different speeds and with different tip 

clearance configurations. A summary of the test cases, Figure 4.6, highlights several regions of a 

tip clearance vs. speed matrix; the four cases presented thus far are noted by black dots in Figure 

4.6. As suggested by Figure 3.10 and shown in Figure 4.1, the 100% corrected speedline for TC1 

portrays distinctly modal stall inception traits. This test case is representative of the original 

design conditions for the compressor, suggesting the compressor is likely well-matched, and it 

stalls via long length-scale perturbations. The 90% corrected speedline for TC1 and the 100% 

corrected speedline for TC2 exemplify a transitional range away from the distinct modes to also 

incorporate spike-type stall inception traits. Ultimately, though, the compressor exhibits clear 

spike-type stall inception mechanisms (i.e., Figure 4.5) for the conditions which depart most 

significantly from the original design intent of the machine. 

Figure 4.6: Summary of stall inception trends for all tip clearance configurations at all 

investigated operating speeds. 

NASA/CR—2015-218868 89
CHAPTER 5: TIME-RESOLVED FLOW FIELD MEASUREMENTS 

Reid A. Berdanier and Nicole L. Key 

5.1. Over-Rotor Static Pressures 

The tip leakage flow trajectory can be tracked using the dynamic static pressure 

measurement system described in Section 2.5.1. These data have been collected at a sampling 

frequency of 100 kHz with a low-pass filter cut-off frequency of 40 kHz for a total of 500 

revolutions, as defined by the 1/rev signal from the laser tachometer. The large number of 

revolutions makes it possible to represent an average flow field using the phase-locked ensemble 

averaging (EA) technique: 

á>\Aà =

1



ÖÜ>\

Aâä

ã

äå=

, (5.1)

defined for each revolution ç and each time é. Using Equation (5.1), the ensemble average static 

pressure field over Rotor 1 at NL is shown for each of the three tip clearance configurations in 

Figure 5.1. The data presented in Figure 5.1 represent a mean rotor tip flow, calculated by 

dividing the ensemble-averaged signal into the 36 separate rotor blade passages, and averaging 

across the 36 passages. This average result is shown twice, assuming periodicity, to more easily 

discern the applicable flow features. Also in Figure 5.1, the axial position of the utilized sensors 

is shown at the bottom of each figure as a series of black dots. For these initial comparisons, the 

data were collected at one position with respect to the upstream and downstream vane rows. 

However, these static pressure data were all collected at the same position with respect to the 

stator vanes and the measurement location was selected to ensure that the sensors were not 

positioned in the wake from the upstream stator vane. Further consideration is given to the 

variability of the tip leakage flow due to interaction with the stationary vane rows in the next 

section. 

Figure 5.1: Ensemble-average static pressure field over Rotor 1 for each tip clearance 

configuration at NL. 

Axial Chord [%]

t / BPP [−]

TC1

−20 0 20 40 60 80 100 120

0

1

2

<P>/Pamb

0.881 

0.893 

0.905 

0.918 

0.930 

0.942 

0.954 

0.966 

0.979 

0.991 

1.003 

Axial Chord [%]

t / BPP [−]

TC2

−20 0 20 40 60 80 100 120

0

1

2

<P>/Pamb

0.870 

0.883 

0.897 

0.910 

0.924 

0.938 

0.951 

0.964 

0.978 

0.991 

1.005 

Axial Chord [%]

t / BPP [−]

TC3

−20 0 20 40 60 80 100 120

0

1

2

<P>/Pamb

0.871 

0.884 

0.898 

0.911 

0.924 

0.938 

0.951 

0.964 

0.977 

0.991 

1.004 

NASA/CR—2015-218868 90
The phase-locked ensemble average data in Figure 5.1 identify a low pressure region, 

representing the tip leakage flow. However, the leakage flow trajectory becomes less distinct 

downstream of approximately 60% axial chord. As an alternative method for identifying the 

leakage flow path, previous authors have also used the root-mean square (RMS) unsteadiness 

with respect to the ensemble average: 

RMS>\

A = Ñ

1



ÖÜ>\

A − á>\Aàâ

;

ä

ã

äå=

. (5.2)

This definition of RMS unsteadiness provides the ability to more definitively identify regions of 

pressure fluctuation. For example, the same data from Figure 5.1 (previously presented as an 

ensemble average) are recast in terms of the RMS with respect to the ensemble average in Figure 

5.2. This figure shows that the RMS definition provides the capability to more easily distinguish 

the tip leakage flow trajectory up to nearly 80% axial chord. 

Figure 5.2: Static pressure field unsteadiness over Rotor 1, presented as an RMS with respect to 

the EA, for each tip clearance configuration at NL. 

Similar to the data presented at NL, an ensemble-average of the static pressure field over 

Rotor 1 at a high loading (HL) operating condition is shown in Figure 5.3 for each tip clearance. 

Compared to the NL results, these HL data show that the tip leakage flow region has increased 

its trajectory angle across the rotor passage, as expected from the increased incidence angle on 

the rotor and the increased pressure difference across the blade tip. Initially, there is no 

discernable difference in the trajectory angle between these tip clearance configurations in Figure 

5.3, although the RMS with respect to the ensemble average can again be used to more easily 

track the leakage flow trajectory across the rotor passage. 

These RMS data, Figure 5.4, highlight flow patterns at the high loading operating condition 

which are not clearly observable in the ensemble average results of Figure 5.3. Of note, the 

trajectory angle of the tip leakage flow increases as the tip clearance height increases from TC1 

to TC3. This observed trend can be attributed to the fact that the HL conditions for the three tip 

clearances represent positions which are at varying positions with respect to the stall point, as 

Axial Chord [%]

t / BPP [−]

TC1

−20 0 20 40 60 80 100 120

0

1

2

PRMS/Pamb

0.002 

0.004 

0.006 

0.008 

0.010 

0.012 

0.014 

0.016 

0.018 

0.020 

0.022 

Axial Chord [%]

t / BPP [−]

TC2

−20 0 20 40 60 80 100 120

0

1

2

PRMS/Pamb

0.002 

0.004 

0.006 

0.008 

0.010 

0.012 

0.014 

0.016 

0.018 

0.020 

0.022 

Axial Chord [%]

t / BPP [−]

TC3

−20 0 20 40 60 80 100 120

0

1

2

PRMS/Pamb

0.002 

0.004 

0.006 

0.008 

0.010 

0.012 

0.014 

0.016 

0.018 

0.020 

0.022 

NASA/CR—2015-218868 91
discussed in Section 3.3.1. In particular for the TC3 tip clearance configuration, Figure 5.4 

shows the leakage flow is approaching a tangential angle. 

At the HL operating condition shown in Figure 5.4, the larger tip clearance configurations 

(TC2 and TC3) show regions of large unsteadiness (RMS) where the tip leakage flow 

disturbance impinges on the adjacent blade. These data do not provide a definitive explanation 

for whether or not this machine displays the “double-leakage” discussed by Khalsa (1996). 

However, there is reason for speculation, particularly related to the region of high flow 

unsteadiness on the suction side of the rotor blade in the region of 30 to 60% axial chord for TC2 

and 20 to 50% axial chord for TC3. An advanced data collection campaign implementing nonintrusive measurements techniques (such as the PIV methodologies introduced in this project) 

inside the blade passage could provide more insight into this phenomenon. 

Figure 5.3: Ensemble-average static pressure field over Rotor 1 for each tip clearance 

configuration at HL. 

Figure 5.4: Static pressure field unsteadiness over Rotor 1, presented as an RMS with respect to 

the EA, for each tip clearance configuration at HL. 

Axial Chord [%]

t / BPP [−]

TC1

−20 0 20 40 60 80 100 120

0

1

2

<P>/Pamb

0.875 

0.890 

0.906 

0.921 

0.937 

0.952 

0.967 

0.983 

0.998 

1.014 

1.029 

Axial Chord [%]

t / BPP [−]

TC2

−20 0 20 40 60 80 100 120

0

1

2

<P>/Pamb

0.881 

0.896 

0.910 

0.925 

0.939 

0.954 

0.969 

0.983 

0.998 

1.012 

1.027 

Axial Chord [%]

t / BPP [−]

TC3

−20 0 20 40 60 80 100 120

0

1

2

<P>/Pamb

0.897 

0.910 

0.922 

0.935 

0.947 

0.960 

0.973 

0.985 

0.998 

1.010 

1.023 

Axial Chord [%]

t / BPP [−]

TC1

−20 0 20 40 60 80 100 120

0

1

2

PRMS/Pamb

0.002 

0.004 

0.006 

0.008 

0.010 

0.012 

0.014 

0.016 

0.018 

0.020 

0.022 

Axial Chord [%]

t / BPP [−]

TC2

−20 0 20 40 60 80 100 120

0

1

2

PRMS/Pamb

0.002 

0.004 

0.006 

0.008 

0.010 

0.012 

0.014 

0.016 

0.018 

0.020 

0.022 

Axial Chord [%]

t / BPP [−]

TC3

−20 0 20 40 60 80 100 120

0

1

2

PRMS/Pamb

0.002 

0.004 

0.006 

0.008 

0.010 

0.012 

0.014 

0.016 

0.018 

0.020 

0.022 

NASA/CR—2015-218868 92
clearance height can be determined. These trajectories, shown in Figure 5.5, have been identified 

as the locus of the peak RMS values. The comparison for these trajectories has been specifically 

selected for the Rotor 1 data because they represent comparisons of tip leakage flow trajectories 

which are less affected by upstream flow conditions. In all cases, Rotor 1 experiences the same 

“clean” inlet flow exiting the IGV, whereas the rotors in the aft stages of the compressor 

experience more variability due to ingestion of the upstream flow disturbances. 

For the two high flow rate conditions, NL and LL, Figure 5.5 shows an increase of tip 

clearance height has the effect of moving the leakage flow trajectory closer to the blade suction 

surface. This figure also more clearly reveals a non-linear (“kinked”) leakage flow trajectory, 

which is especially apparent for the TC3 tip clearance configuration at the high flow rate 

conditions, LL and NL. This observation was also noted by Yoon et al. (2006) near mid-chord, 

but the results shown here represent a less drastic turn of the leakage flow than was observed by 

those authors. Chen (1991) theoretically predicts this non-linear trajectory, but not until the end 

of the passage, where the image vortices required to satisfy kinematic constraints change with 

the absence of the blade as an effective wall. 

In addition to these observed trends at high flow rate operating conditions, Figure 5.5 also 

depicts a noticeable change of the leakage flow trajectory concavity at the high loading condition 

as the tip clearance height increases. The discussion of Figure 5.4 noted the change of the flow 

angle between the three tip clearance configurations, but the locus of peak RMS values shows 

the two larger tip clearance heights turn more noticeably toward the adjacent blade. It is likely 

that this observed change of flow trajectory path is also due to the relative proximity of the HL 

points to the stall condition with different tip clearance heights. 

Figure 5.5: Rotor 1 tip leakage trajectories for all three tip clearance configurations at three 

loading conditions on the 100% corrected speedline. 

The over-rotor static pressure fields presented thus far have compared the results for one 

rotor with three tip clearance heights. In contrast, Figure 5.6 shows the RMS unsteadiness of the 

static pressure field over each of the three rotor rows at the NL condition for TC2. As with the 

Rotor 1 results shown above, these figures represent a mean flow field, for which the data have 

NASA/CR—2015-218868 93

Using the RMS pressures shown in Figure 5.2, Figure 5.4, and a similar figure for the low 

loading condition, the effect of the tip leakage flow trajectory on loading condition and tip 
been separated into the number of blades in the row (36, 33, or 30 for Rotor 1, Rotor 2, and 

Rotor 3, respectively) and subsequently averaged. These averaged data are then repeated 

(assuming periodicity) to more clearly distinguish the flow features in the blade passages. 

In Figure 5.6, a clear distinction can be made between the data for Rotor 1 and the data for 

the Rotor 2 or Rotor 3. As shown above in Figure 5.5, the trajectory of the Rotor 1 tip leakage 

flow is non-linear, but this is less present in Rotor 2 or Rotor 3. Furthermore, the Rotor 2 and 

Rotor 3 results portray a wider region of high flow unsteadiness (RMS), indicating a different 

shape of the leakage flow at these positions. 

In the same manner, Figure 5.7 shows the RMS unsteadiness over each rotor for TC2 at HL. 

As with the nominal loading condition, a clear distinction can be drawn between the Rotor 1 data 

and the results for Rotor 2 or Rotor 3. Although the angle of the leakage flow trajectory is similar 

for the three rotors, a wider region of high flow unsteadiness is present for the Rotor 2 and Rotor 

3 results, representing a larger leakage flow disturbance for these rotors. 

The measurements in the fixed frame of reference for the data in Figure 5.6 and Figure 5.7 

are nearly identical in position with respect to the upstream stator vane row and the associated 

low-pressure wake region. Thus, the results in these figures further emphasize the differences 

measured over Rotor 1, which always ingests the same clean inlet flow from the IGV, compared 

to the latter Rotor rows, which are subject to flow disturbance ingestion from the upstream rotor 

rows and, in particular, the affected flow in the outer region near the rotor tip. 

Figure 5.6: Static pressure field unsteadiness, presented as an RMS, over each rotor 

for TC2 at NL. 

Axial Chord [%]

t / BPP [−]

Rotor 1

−20 0 20 40 60 80 100 120

0

1

2

PRMS/Pamb

0.002 

0.004 

0.006 

0.008 

0.010 

0.012 

0.014 

0.016 

0.018 

0.020 

0.022 

Axial Chord [%]

t / BPP [−]

Rotor 2

−20 0 20 40 60 80 100 120

0

1

2

PRMS/Pamb

0.002 

0.004 

0.006 

0.008 

0.010 

0.012 

0.014 

0.016 

0.018 

0.020 

0.022 

Axial Chord [%]

t / BPP [−]

Rotor 3

−20 0 20 40 60 80 100 120

0

1

2

PRMS/Pamb

0.002 

0.004 

0.006 

0.008 

0.010 

0.012 

0.014 

0.016 

0.018 

0.020 

0.022 

NASA/CR—2015-218868 94
Figure 5.7: Static pressure field unsteadiness, presented as an RMS, over each rotor 

for TC2 at HL. 

5.2. Influence of Stator Wakes on Leakage Flow 

The wealth of data presented so far for this project have suggested the loading condition of 

the compressor can affect the overall influence of the tip leakage flow. In particular, the overrotor static pressure field time series in Section 5.1 above showed the trajectory of the leakage 

flow through the rotor passage changes significantly as the compressor loading is increased 

toward stall. With this understanding, the rotor tip leakage flow can also be noticeably affected 

by position with respect to the upstream stator vane row. 

In the wake of a stator vane (i.e., Stator 1), the absolute velocity deficit creates a 

corresponding increase of incidence into the downstream rotor row (i.e., Rotor 2). The local 

increase of incidence angle loads up the front of the rotor, affecting the strength and trajectory of 

the tip leakage flow. As a result, the location of the probe in the stationary frame of reference 

will dictate whether or not the measured flow parameter (pressure, velocity, etc.) will be affected 

by the rotor passing through the upstream vane wake. 

This phenomenon is illustrated in Figure 5.8 with respect to the rotor passing through the 

upstream stator wake. In Figure 5.8(a), the rotor leading edge (Rotor 2, in this case) is positioned 

between stator wakes, and a corresponding velocity triangle is shown. At this location, the 

leakage flow propagates through the rotor passage and is measured by a series of probe locations 

at the rotor exit plane. In Figure 5.8(b), the rotor leading edge is located in the wake of the 

upstream stator, and the associated velocity triangle appears as a deficit of velocity in the 

absolute frame of reference. In this case, as the rotor loading increases, the leakage flow region 

increases in size, and the trajectory changes as its inception point moves upstream toward the 

rotor leading edge. 

Axial Chord [%]

t / BPP [−]

Rotor 1

−20 0 20 40 60 80 100 120

0

1

2

PRMS/Pamb

0.002 

0.004 

0.006 

0.008 

0.010 

0.012 

0.014 

0.016 

0.018 

0.020 

0.022 

Axial Chord [%]

t / BPP [−]

Rotor 2

−20 0 20 40 60 80 100 120

0

1

2

PRMS/Pamb

0.002 

0.004 

0.006 

0.008 

0.010 

0.012 

0.014 

0.016 

0.018 

0.020 

0.022 

Axial Chord [%]

t / BPP [−]

Rotor 3

−20 0 20 40 60 80 100 120

0

1

2

PRMS/Pamb

0.002 

0.004 

0.006 

0.008 

0.010 

0.012 

0.014 

0.016 

0.018 

0.020 

0.022 

NASA/CR—2015-218868 95
Figure 5.8: Schematic of the interaction between rotor tip leakage flow and upstream 

stator wake. (a) R2 passing in the middle of the S1 passage, (b) R2 passing 

through the S1 wake. 

The modulation of the tip leakage flow due to an interaction with the upstream stator wake 

is exemplified qualitatively using flow visualization results on the compressor casing endwall, 

Figure 5.9. In this figure, the casing endwall has been photographed over each of the three rotor 

rows following flow visualization tests at each of three loading conditions for TC1. In this figure, 

the approximate axial location of leakage flow inception can be identified by the path of absent 

paint (near mid-chord for nominal loading, and moving forward toward the blade leading edge as 

the loading increases). Also in the photographs of Figure 5.9, the stator wakes can be identified 

by similar regions of absent paint – the wakes for Stator 1 and Stator 2 are most clearly identified 

at the high loading and near stall operating conditions. 

As suggested by the cartoon schematic in Figure 5.8, when the rotor passes through the 

upstream stator wake (in the absolute reference frame), the velocity deficit causes the position of 

leakage flow inception to move axially upstream. The strength of this modulation is different for 

Rotor 1 than for Rotor 2 or Rotor 3 due to the difference of the IGV wake compared to wake of 

Stator 1 or Stator 2. Furthermore, the modulation pattern on the casing endwall changes as the 

loading condition changes. The fundamentals of this rotor-stator interaction have been 

documented by few authors in the past (Mailach et al., 2008; Krug et al., 2015), but these 

photographs motivate a more in-depth analysis. The following measurements expand upon the 

findings of a recent study at part-speed operating conditions in the Purdue three-stage axial 

compressor (Smith et al., 2015b). 

NASA/CR—2015-218868 96
Figure 5.9: Flow visualization on the casing endwall highlighting a modulation of the tip 

leakage flow for TC1. 

NASA/CR—2015-218868 97
5.2.1. Influence of Stator Wakes on Leakage Flow Trajectory 

As described in Section 5.1, the over-rotor static pressure field measurements presented 

above were all acquired at one fixed position with respect to the upstream and downstream vane 

rows. However, the interaction of the rotor tip leakage flow and the upstream stator wakes 

suggested by previous authors, and alluded to in the photograph of Figure 5.9, motivates 

additional analysis. Thus, over-rotor static pressure measurements were repeated with the 

baseline tip clearance, TC1, at several loading conditions for each rotor. 

The cartoon schematic in Figure 5.10 outlines the measurement process conducted using 

the over-rotor static pressure measurement array. The sensors in this measurement system are in 

fixed positions in the compressor casing, but the stator vanes can be moved (either together or 

independently). Thus, Figure 5.10 shows representative measurements that may be measured at 

two vane positions with respect to the fixed measurement locations: (a) the sensors are located 

between the upstream stator wakes and measure the freestream flow, or (b) the sensors are 

approximately in the upstream stator wake. In each case, the time-resolved measurements are 

phase-locked with the rotor rotation, so the shaded measurement region identifies one blade 

passage of time-resolved static pressure data for each of the two vane positions. The cartoons in 

Figure 5.10 show measurements over Rotor 2 for which Stator 1 and Stator 2 are moved together, 

but similar measurements were collected over Rotor 1 and Rotor 3, and some data were also 

collected over Rotor 2 when Stator 1 and Stator 3 were moved independently from one another. 

Figure 5.10: Schematic of leakage flow measurements over the rotor for two vane wake-leakage 

flow interaction positions, (a) and (b). A boxed region identifies a representative snapshot of the 

leakage flow identified by the static pressure measurements. 

The first of these measurements, Figure 5.11, highlights the Rotor 1 leakage flow variations 

when moving all of the stator vanes simultaneously to 25 unique equally-spaced pitchwise 

positions with respect to the fixed dynamic pressure sensor positions for the nominal loading 

(NL) condition. Figure 5.11 shows the unsteadiness represented by the RMS with respect to the 

ensemble average. For each representation of the pressure measurements, the diagrams may be 

followed in a clockwise direction, as identified by the directional arrow. 

To more clearly separate the modulation of the leakage flow from the 25 pitchwise 

positions shown in Figure 5.11, two of the pitchwise vane positions have been selected for 

NASA/CR—2015-218868 98
comparison. These positions, identified in Figure 5.11 as vpA and vpB, are approximately 50% 

vane passage out of phase from one another. 

The RMS unsteadiness results for these two discrete vane positions are enlarged and 

compared in Figure 5.12(a-b). For these data, the leakage flow trajectories have been identified 

by connecting the locus of peak unsteadiness positions. The comparison of trajectories, Figure 

5.12(c), shows a small but identifiable difference between the trajectories for the two vane 

positions, vpA and vpB, on the order of 1.5 degrees. This difference may seem small, but it is 

significant, especially compared to the effects observed for different tip clearances and loading 

conditions. Specifically, this modulation of the leakage flow trajectory identified in Figure 5.12(c) 

for the stator-rotor interaction is 50% more than the leakage flow trajectory change for the tip 

clearance increase from TC1 to TC2 identified in Figure 5.5. 

This quantification shows that the motion of the rotor into and out of the upstream stator 

wake can change the leakage flow trajectory more than a doubling of the tip clearance height 

from 1.5% to 3.0% annulus height. To adequately study the tip leakage flow, data acquisition at 

one pitchwise position with respect to the stationary vane rows is insufficient. This pitchwise 

modulation must be carefully considered, especially when attempting to compare experimental 

results with CFD solutions. 

NASA/CR—2015-218868 99
Figure 5.11: Static pressure field over Rotor 1, averaged for all 36 blades, in terms of RMS static pressure, at ea

across one vane passage (clockwise) for TC1 at NL. 

 (a) (b) (

Figure 5.12: Static pressure unsteadiness (RMS) over Rotor 1 for vpA (a) and vpB (b) from Figure 5.11, and 

Axial Chord [%]

t / BPP [−]

vpA

−20 0 20 40 60 80 100 120

012

Axial Chord [%]

t / BPP [−]

vpB

−20 0 20 40 60 80 100 120

012

NASA/CR—2015-218868 100
The results from Figure 5.11 and Figure 5.12 identified a modulation of the leakage flow 

trajectory that agrees with the flow visualization photographs. However, the IGV wake is 

expected to have a smaller effect on the leakage flow for Rotor 1 than the Stator 1 or Stator 2 

wakes for Rotor 2 and Rotor 3, respectively. Furthermore, it is important here to more clearly 

distinguish the effects of the upstream vane row from those of the downstream vane row to 

ensure that the downstream potential field does not significantly impact the leakage flow 

development. Thus, the static pressure field over Rotor 2 was chosen to investigate both of these 

points. 

To address these questions, two separate vane movement processes were introduced – 

effectively clocking the position of the IGV and Stator 1 with respect to Stator 2 and Stator 3. 

(Recall that IGV, Stator 1, and Stator 2 all have the same vane count of 44, which allows vane 

clocking studies in this facility.) Specifically, the first method moved IGV and Stator 1 

simultaneously to the same 25 pitchwise positions as for Figure 5.11, while Stator 2 and Stator 3 

were maintained in fixed locations. In the second method, Stator 2 and Stator 3 were moved 

simultaneously to the 25 pitchwise positions while the IGV and Stator 1 were held fixed. 

The calculated RMS unsteadiness pressure fields from this two-movement process are 

presented in Figure 5.13 at NL. This figure shows results for the first movement (Move IGV & 

S1, Fix S2 & S3) for the 25 pitchwise positions. As before, two of these positions are identified 

as vpA and vpB, and the results for those two vane positions are shown in Figure 5.14(a-b). 

Comparing the RMS unsteadiness from these two pitchwise vane positions, there is a region of 

high unsteadiness appearing for positions greater than 70% axial chord for vpB which does not 

appear for vpA. Referring to Figure 5.13, this region of high unsteadiness appears and disappears 

while moving clockwise around the 25 pitchwise positions. In addition, the trajectory is also 

compared for these two identified vane positions, in Figure 5.14(c). For this first vane movement 

process, the trajectory of the leakage flow modulates by approximately 1.5 degrees between vpA 

and vpB. In this case, the relative trajectories of vpA and vpB are opposite from the results 

shown in Figure 5.12(c). The positions of the IGV and Stator 1 with respect to the measurement 

sensors are nearly identical, but it is expected that the IGV wake will affect Rotor 1 differently 

than the effect of Stator 1 on Rotor 2 due the nature of the IGV as an accelerating vane row. This 

first vane movement process has isolated the upstream vanes from the downstream vanes for 

Rotor 2 to emphasize the effect of the upstream vane row on the leakage flow development. 

A similar analysis of RMS unsteadiness is also presented for the second vane movement 

process (Move S2 & S3, Fix IGV & S1), Figure 5.15. For this case, the 25 positions in Figure 

5.15 may identify a slight change of the pressure field unsteadiness near the trailing edge of the 

rotor blades where the downstream vane row may affect the flow, but there is no appreciable 

change of the leakage flow pattern or trajectory throughout the 25 positions. This becomes more 

apparent by selecting the same two relative positions, identified as vpA and vpB for comparison 

in Figure 5.16(a-b). The RMS unsteadiness for these two positions is nearly identical, and the 

trajectories traced in Figure 5.16(c) reinforce this invariability. Ultimately, these results in Figure 

5.13 through Figure 5.16 highlight the effect that the upstream vane wakes may impose on the tip 

leakage flow (as suggested in the schematic of Figure 5.8), and they separate observations from 

the potential field associated with the downstream vane row. 

NASA/CR—2015-218868 101
Figure 5.13: Static pressure field over Rotor 2, averaged for all 33 blades, in terms of RMS static pressure, at ea

across one vane passage (clockwise) for TC1 at NL. Only IGV and S1 are moved, S2 and S3 are 

 (a) (b) (

Figure 5.14: Static pressure unsteadiness (RMS) over Rotor 2 for vpA (a) and vpB (b) from Figure 5.13, and 

Axial Chord [%]

t / BPP [−]

vpA

−20 0 20 40 60 80 100 120

012

Axial Chord [%]

t / BPP [−]

vpB

−20 0 20 40 60 80 100 120

012

NASA/CR—2015-218868 102
Figure 5.15: Static pressure field over Rotor 2, averaged for all 33 blades, in terms of RMS static pressure, at ea

across one vane passage (clockwise) for TC1 at NL. Only S2 and S3 are moved, IGV and S1 are 

 (a) (b) (

Figure 5.16: Static pressure unsteadiness (RMS) over Rotor 2 for vpA (a) and vpB (b) from Figure 5.15, and 

Axial Chord [%]

t / BPP [−]

vpA

−20 0 20 40 60 80 100 120

012

Axial Chord [%]

t / BPP [−]

vpB

−20 0 20 40 60 80 100 120

012

NASA/CR—2015-218868 103
To this point, the Rotor 1 leakage flow trajectory modulation was identified at the nominal 

loading condition. However, the IGV is expected to affect the Rotor 1 flow field differently than 

the downstream stators will affect the rotors of the latter stages. As a result, the Rotor 2 pressure 

field was analyzed using two separate vane movement processes to investigate the embedded 

stage and separate the effects of the upstream vane wakes from the downstream potential field. 

At this point, further analysis is warranted to evaluate the effect of changing loading condition, 

and Rotor 3 was selected as the vehicle for this analysis. 

Figure 5.17 identifies a modulation of the leakage flow trajectory for Rotor 3 using the 

simultaneous movement of all vane rows described previously for Figure 5.11. In Figure 5.17, 

the Rotor 3 unsteadiness data at the nominal loading (NL) condition show the modulation of the 

leakage flow for the 25 vane positions. Considering the pitchwise position identified as vpA, 

highlighted in Figure 5.18(a), the locations of high RMS are identified to approximately midchord, and then attenuate slightly prior to reappearing as a high-intensity region at the 

intersection with the trailing edge of the adjacent blade. In contrast, for vpB (Figure 5.18(b)), the 

high RMS regions begin to attenuate further forward in the passage (near 60% axial chord), but 

the high-intensity region at the trailing edge also moved forward in the passage to approximately 

80% axial chord. For Rotor 3 at this NL condition, the approximately trajectories identified in 

Figure 5.18(c) vary less than the results shown previously for Rotor 1 or Rotor 2 at the same 

loading condition. 

At the high loading (HL) operating condition for Rotor 3 in Figure 5.19, a similar trend can 

also be identified. The comparison of vpA and vpB at this HL condition in Figure 5.20(a-b) 

shows a more dramatic difference than for the NL condition. At HL, Figure 5.20(a) shows the 

high unsteadiness region of the leakage flow attenuates at approximately 40% axial chord before 

reappearing as a region of high unsteadiness at its intersection with the adjacent blade, near 65% 

axial chord. In contrast, Figure 5.20(b) shows for vpB that the same attenuation occurs near 40% 

axial chord, but the “island” of high unsteadiness can then be identified near 55% axial chord, 

but not in contact with the adjacent blade. However, Figure 5.20(b) also shows a region of high 

unsteadiness in contact with the adjacent blade centered near 85% axial chord which does not 

align with the trajectory path connecting the locus of high unsteadiness for the rest of the passage. 

The leakage flows in Figure 5.20(a-b) are clearly different, and their approximate trajectories are 

traced in Figure 5.20(c). 

NASA/CR—2015-218868 104
Figure 5.17: Static pressure field over Rotor 3, averaged for all 30 blades, in terms of RMS static pressure, at ea

across one vane passage (clockwise) for TC1 at NL using two separate vane movement techniq

 (a) (b) (

Figure 5.18: Static pressure unsteadiness over Rotor 3 for vpA (a) and vpB (b) from Figure 5.17, and the traject

Axial Chord [%]

t / BPP [−]

vpA

−20 0 20 40 60 80 100 120

012

Axial Chord [%]

t / BPP [−]

vpB

−20 0 20 40 60 80 100 120

012

NASA/CR—2015-218868 105
Figure 5.19: Static pressure field over Rotor 3, averaged for all 30 blades, in terms of RMS static pressure, at ea

across one vane passage (clockwise) for TC1 at HL. 

 (a) (b) (

Figure 5.20: Static pressure unsteadiness over Rotor 3 for vpA (a) and vpB (b) from Figure 5.19, and the traject

Axial Chord [%]

t / BPP [−]

vpA

−20 0 20 40 60 80 100 120

012

Axial Chord [%]

t / BPP [−]

vpB

−20 0 20 40 60 80 100 120

012

NASA/CR—2015-218868 106
5.2.2. Rotor Exit Total Pressures 

To further capture the effect of the upstream vane wake on the rotor tip leakage flow, the 

time-resolved total pressure measurements downstream of the rotor blade rows have been 

completed for up to 28 radial positions at each of the 26 pitchwise positions (pp) across one vane 

pitch. (Of these 26 positions, 25 are unique positions, since one pair of positions, pp1 and pp26, 

are periodic with respect to the vane pitch.) The cartoon schematic in Figure 5.21 shows the 

relative locations of the probe positions at the rotor exit plane for simultaneous movements of all 

vane rows as the rotor interacts with the upstream vane wake to modulate the rotor tip leakage 

flow. 

Figure 5.21: Schematic of leakage flow measurements at rotor exit planes for two vane 

wake-leakage flow interaction positions, (a) and (b). 

The time-resolved total pressure measurements are presented here as an unsteadiness using 

the RMS with respect to the ensemble average, as defined in Equation (5.2). This presentation of 

results accommodates the identification of recirculating flow regions expected in the tip leakage 

flow disturbance, as well as the rotor wake. These RMS measurements are presented for the 

Rotor 1 exit plane at the NL and HL operating conditions in Figure 5.22 and Figure 5.23, 

respectively. 

In these figures, the RMS with respect to the ensemble average (500 revolutions, for these 

data) has been calculated at each circumferential vane position, and then averaged across all 

rotor blades (36 blades for Rotor 1) to identify an averaged rotor passage. The IGV wake is very 

thin and has a minimal effect on the downstream rotor row. As a result, Figure 5.22 and Figure 

5.23 show a weak modulation of the tip leakage flow disturbance across one vane pitch 

(following the figures in a clockwise direction, denoted by the titles pp1-pp26). However, an 

increased unsteadiness at the center of the leakage flow disturbance region is identifiable, 

particularly for the TC3 tip clearance configuration measurements at the NL operating condition 

in Figure 5.22. In addition to the pitchwise modulation of the flowfield exemplified by the 26 

circumferential measurement positions, a mean contour represents the rotor wake averaged 

across one vane pitch. The PS and SS of the rotor blade are also labeled in this mean contour to 

orient the figure. 

At the nominal loading condition, Figure 5.22, a clear distinction can be made between the 

tip leakage flow disturbances for each of the tip clearance configurations, TC1 and TC3. For the 

NASA/CR—2015-218868 107
smaller tip clearance, the unsteadiness due to the tip leakage flow occupies a region which 

extends down from the outer casing endwall to approximately 80% annulus height, whereas the 

larger tip clearance height shows a leakage flow region which reaches approximately 70% 

annulus height, as well as affecting a noticeably larger area in the pitch-wise direction. 

At the same Rotor 1 exit position for the HL operating condition, Figure 5.23 shows the 

TC1 leakage flow region no longer has a coherent circular shape, but rather fills a corner area 

adjacent to the pressure side of the rotor wake. The TC3 tip clearance configuration, on the other 

hand, maintains a shape which closely resembles the result from the NL condition in Figure 5.22. 

However, its size and relatively intensity of the unsteadiness parameter are both increased at the 

higher loading condition. Also at the HL condition, both tip clearance configurations show a 

noticeable change as the hub corner separation off the suction surface previously observed at NL 

grows into a thick wake which extends above 50% annulus height. 

NASA/CR—2015-218868 108
Figure 5.22: Average Rotor 1 exit flow field, in terms of RMS total pressure, at each circumferential location across one vane passage 

(clockwise) at NL. 

TC1

Annulus Height [%]pp1

0

50

100

pp2 pp3 pp4 pp5 pp6 pp7 pp8 pp9 pp10 pp11 pp12 pp13

pp14 pp15 pp16 pp17 pp18 pp19 pp20 pp21 pp22 pp23 pp24 pp25

Annulus Height [%]pp26

0

50

100

P

o,3,RMS / Po,AIP,RAA

0.035 0.031 0.026 0.022 0.018 0.014 0.009 0.005 

Annulus Height [%]

Mean

PS SS

100806040200

TC3

Annulus Height [%]pp1

0

50

100

pp2 pp3 pp4 pp5 pp6 pp7 pp8 pp9 pp10 pp11 pp12 pp13

pp14 pp15 pp16 pp17 pp18 pp19 pp20 pp21 pp22 pp23 pp24 pp25

Annulus Height [%]pp26

0

50

100

P

o,3,RMS / Po,AIP,RAA

0.035 0.031 0.026 0.022 0.018 0.014 0.009 0.005 

Annulus Height [%]

Mean

PS SS

100806040200

NASA/CR—2015-218868 109
Figure 5.23: Average Rotor 1 exit flow field, in terms of RMS total pressure, at each circumferential location across one vane passage 

(clockwise) at HL. 

TC1

Annulus Height [%]pp1

0

50

100

pp2 pp3 pp4 pp5 pp6 pp7 pp8 pp9 pp10 pp11 pp12 pp13

pp14 pp15 pp16 pp17 pp18 pp19 pp20 pp21 pp22 pp23 pp24 pp25

Annulus Height [%]pp26

0

50

100

P

o,3,RMS / Po,AIP,RAA

0.035 0.031 0.026 0.022 0.018 0.014 0.009 0.005 

Annulus Height [%]

Mean

PS SS

100806040200

TC3

Annulus Height [%]pp1

0

50

100

pp2 pp3 pp4 pp5 pp6 pp7 pp8 pp9 pp10 pp11 pp12 pp13

pp14 pp15 pp16 pp17 pp18 pp19 pp20 pp21 pp22 pp23 pp24 pp25

Annulus Height [%]pp26

0

50

100

P

o,3,RMS / Po,AIP,RAA

0.035 0.031 0.026 0.022 0.018 0.014 0.009 0.005 

Annulus Height [%]

Mean

PS SS

100806040200

NASA/CR—2015-218868 110
Time-resolved total pressure measurements at Rotor 2 exit for the NL and HL operating 

conditions, Figure 5.24 and Figure 5.25, respectively, are also presented as the RMS 

unsteadiness with respect to the ensemble average. At Rotor 2 exit, the leakage flow modulation 

is more clearly distinguished by following the measurements at each position across one vane 

pitch. For TC1 at the nominal loading condition, two adjacent high-intensity regions of flow 

unsteadiness are identifiable in the outer tip region for the measurements labeled pp1–pp7. These 

two regions compare with the results observed in Figure 5.6 (for TC2, although the figure for 

TC1 is similar), which is expected based on the position of the static pressure measurements 

falling in this same range with respect to the stator vanes. The strength of the two highunsteadiness regions is also sufficiently significant to persist into the circumferential mean for 

TC1 in Figure 5.24. The TC3 results in Figure 5.24 position the extremes of the pitchwise 

leakage flow modulation in the same locations as for TC1. Further, the measurements are similar 

to the observations for Rotor 1, such that the radial extent of the leakage flow region for TC3 

extends to approximately 70% annulus height, compared to 80% for TC1. 

At the high loading condition for Rotor 2 presented in Figure 5.25, the unsteadiness of the 

flow identified by the RMS fluctuations shows less modulation than at the nominal loading 

condition (the RMS of the passage represents magnitudes which exist more in the middle of the 

prescribed contour range), especially for TC1. However, two regions of high pressure 

unsteadiness are still faintly present in the mean contour for TC1. Comparing the HL condition 

with the NL results for Rotor 2, the circumferential probe position which yields the most intense 

leakage flow structure is not the same. This result is expected as the width of the upstream stator 

wake is significantly changed between the two conditions. Furthermore, the trajectory of the 

leakage flow is expected to change as the rotor loading is increased. 

NASA/CR—2015-218868 111
Figure 5.24: Average Rotor 2 exit flow field, in terms of RMS total pressure, at each circumferential location across one vane passage 

(clockwise) at NL. 

TC1

Annulus Height [%]pp1

0

50

100

pp2 pp3 pp4 pp5 pp6 pp7 pp8 pp9 pp10 pp11 pp12 pp13

pp14 pp15 pp16 pp17 pp18 pp19 pp20 pp21 pp22 pp23 pp24 pp25

Annulus Height [%]pp26

0

50

100

P

o,5,RMS / Po,AIP,RAA

0.035 0.031 0.026 0.022 0.018 0.014 0.009 0.005 

Annulus Height [%]

Mean

PS SS

100806040200

TC3

Annulus Height [%]pp1

0

50

100

pp2 pp3 pp4 pp5 pp6 pp7 pp8 pp9 pp10 pp11 pp12 pp13

pp14 pp15 pp16 pp17 pp18 pp19 pp20 pp21 pp22 pp23 pp24 pp25

Annulus Height [%]pp26

0

50

100

P

o,5,RMS / Po,AIP,RAA

0.035 0.031 0.026 0.022 0.018 0.014 0.009 0.005 

Annulus Height [%]

Mean

PS SS

100806040200

NASA/CR—2015-218868 112
Figure 5.25: Average Rotor 2 exit flow field, in terms of RMS total pressure, at each circumferential location across one vane passage 

(clockwise) at HL. 

TC1

Annulus Height [%]pp1

0

50

100

pp2 pp3 pp4 pp5 pp6 pp7 pp8 pp9 pp10 pp11 pp12 pp13

pp14 pp15 pp16 pp17 pp18 pp19 pp20 pp21 pp22 pp23 pp24 pp25

Annulus Height [%]pp26

0

50

100

P

o,5,RMS / Po,AIP,RAA

0.035 0.031 0.026 0.022 0.018 0.014 0.009 0.005 

Annulus Height [%]

Mean

PS SS

100806040200

TC3

Annulus Height [%]pp1

0

50

100

pp2 pp3 pp4 pp5 pp6 pp7 pp8 pp9 pp10 pp11 pp12 pp13

pp14 pp15 pp16 pp17 pp18 pp19 pp20 pp21 pp22 pp23 pp24 pp25

Annulus Height [%]pp26

0

50

100

P

o,5,RMS / Po,AIP,RAA

0.035 0.031 0.026 0.022 0.018 0.014 0.009 0.005 

Annulus Height [%]

Mean

PS SS

100806040200

NASA/CR—2015-218868 113
Similar to the other rotor rows, the unsteadiness of the Rotor 3 exit total pressure field is 

shown in Figure 5.26 and Figure 5.27 for the NL and HL operating conditions, respectively. At 

the NL condition, the results for TC1 identify two adjacent regions of high unsteadiness (in 

agreement with other results in this section of the document) – one in the middle of the passage 

and another adjacent to the blade pressure surface. As with Rotor 2, this observation compares 

with the over-rotor static pressures for TC2 at the NL condition (Figure 5.6). For the TC3 

configuration, the circumferential modulation of the leakage flow in Figure 5.26 is less 

discernable compared to the TC1 measurements. This observation suggests the leakage flow 

region with the large tip clearance height is less sensitive to the influence from the upstream 

stator vane. 

For the high loading condition, Figure 5.27, the TC1 unsteadiness is dominated by the rotor 

wake more than the leakage flow. Comparing the pitchwise modulation results, the location with 

the smallest or weakest leakage flow region corresponds to the most significant rotor wake, as 

the decreased leakage flow blockage provides less energizing benefit to the suction surface wake 

separation tendencies in the lower portion of the annulus. Similarly, the positions with the largest 

leakage flow regions show the least suction surface separation (i.e., the thinner rotor wakes). 

Also at the high loading condition, Figure 5.27 shows the leakage flow region for TC3 fills 

nearly the entire rotor passage, with radial extent down to approximately 70% annulus height. In 

contrast, the region of high unsteadiness identified as the leakage flow for TC1 only affects 

approximately half of the rotor passage with radial extent to just below 80% annulus height but 

with a more significant rotor wake. 

Ultimately, these rotor exit total pressure field measurements provide valuable insight into 

the rotor tip leakage flow disturbance as it enters the downstream vane passage. Furthermore, the 

influence of the upstream stator wake can play an important role in the size, shape, and trajectory 

of the leakage flow through the rotor passage – particularly for Rotor 2 and Rotor 3, whereas 

Rotor 1 is weakly affected by the IGV wake. In combination with the over-rotor static pressure 

field measurements, these rotor exit data can provide bounding planes which could be used to 

influence more focused measurements in the rotor passage using non-intrusive measurement 

techniques, such as the PIV methods introduced in Chapter 6. 

NASA/CR—2015-218868 114
Figure 5.26: Average Rotor 3 exit flow field, in terms of RMS total pressure, at each circumferential location across one vane passage 

(clockwise) at NL. 

TC1

Annulus Height [%]pp1

0

50

100

pp2 pp3 pp4 pp5 pp6 pp7 pp8 pp9 pp10 pp11 pp12 pp13

pp14 pp15 pp16 pp17 pp18 pp19 pp20 pp21 pp22 pp23 pp24 pp25

Annulus Height [%]pp26

0

50

100

P

o,7,RMS / Po,AIP,RAA

0.035 0.031 0.026 0.022 0.018 0.014 0.009 0.005 

Annulus Height [%]

Mean

PS SS

100806040200

TC3

Annulus Height [%]pp1

0

50

100

pp2 pp3 pp4 pp5 pp6 pp7 pp8 pp9 pp10 pp11 pp12 pp13

pp14 pp15 pp16 pp17 pp18 pp19 pp20 pp21 pp22 pp23 pp24 pp25

Annulus Height [%]pp26

0

50

100

P

o,7,RMS / Po,AIP,RAA

0.035 0.031 0.026 0.022 0.018 0.014 0.009 0.005 

Annulus Height [%]

Mean

PS SS

100806040200

NASA/CR—2015-218868 115
Figure 5.27: Average Rotor 3 exit flow field, in terms of RMS total pressure, at each circumferential location across one vane passage 

(clockwise) at HL. 

TC1

Annulus Height [%]pp1 050 100 pp2 pp3 pp4 pp5 pp6 pp7 pp8 pp9 pp10 pp11 pp12 pp13

pp14 pp15 pp16 pp17 pp18 pp19 pp20 pp21 pp22 pp23 pp24 pp25

Annulus Height [%]pp26 050 100

P

o,7,RMS / Po,AIP,RAA

0.035 0.031 0.026 0.022 0.018 0.014 0.009 0.005 

Annulus Height [%]

Mean

PS SS

100806040200

TC3

Annulus Height [%]pp1 050 100 pp2 pp3 pp4 pp5 pp6 pp7 pp8 pp9 pp10 pp11 pp12 pp13

pp14 pp15 pp16 pp17 pp18 pp19 pp20 pp21 pp22 pp23 pp24 pp25

Annulus Height [%]pp26 050 100

P

o,7,RMS / Po,AIP,RAA

0.035 0.031 0.026 0.022 0.018 0.014 0.009 0.005 

Annulus Height [%]

Mean

PS SS

100806040200

NASA/CR—2015-218868 116
5.2.3. Quantifying the Size of the Leakage Flow 

The RMS unsteadiness of these rotor exit total pressure data provides a unique opportunity 

to quantify the size of the leakage flow region at the rotor exit measurement plane. Suder (1998) 

uses the calculation of axial velocity gradients to identify a defect region for quantifying flow 

blockage. A similar method has been implemented with these RMS total pressures to identify a 

defect region. Specifically, the gradient of RMS pressure was calculated in the radial and 

circumferential directions, and a cutoff value was assigned to determine the defect region, as 

exemplified in Figure 5.28. Once the defect region was identified, a human-guided selection tool 

in the plotting GUI manually separated the points from the defect region associated with the 

leakage flow from those associated with the rotor wake or the hub endwall boundary layer. These 

separated leakage flow points are also highlighted for the example in Figure 5.28. These points 

associated with the leakage flow disturbance region were then integrated across the passage to 

determine the percentage of the rotor exit flow passage which is affected by the tip leakage flow. 

Figure 5.28: Identification of defect region associated with leakage flow unsteadiness. 

This process was repeated for each of the 26 pitchwise measurement positions presented in 

Figure 5.22 through Figure 5.27. Because there is a potential error associated with the humanguided identification technique, the identification process was repeated three times for all of the 

data sets. This repetition facilitates the calculation of pertinent statistics related to humanintroduced variability. The average of these leakage flow disturbed areas was then calculated at 

each pitchwise position for each measurement condition. 

A passage-averaged value was calculated for each measurement condition, as shown by the 

filled bar plots in Figure 5.29. As shown in the figures above, the leakage flow region can change 

significantly for the different pitchwise measurement positions. Thus, the minimum affected area 

and maximum affected area are noted in Figure 5.29 by range bars about the passage-averaged 

value. For reference, and in the context of the results shown in Figure 5.29, the humanintroduced variability may be represented by the standard deviation with respect to the sample 

0

10

20

30

40

50

60

70

80

90

100

PS SS

Annulus Height [%]

RMS Total Pressure

0

10

20

30

40

50

60

70

80

90

100

Defect Region

NASA/CR—2015-218868 117
mean. The calculated disturbed areas yield a standard deviation of less than 2% of the passage 

area for all conditions with the TC1 configuration. The TC3 data have a standard deviation of 

less than 2.5% of the passage area for Rotor 1 and Rotor 2 data, but approaching 5% of the 

passage area for the Rotor 3 data. 

The information in Figure 5.29 show the expected result of increased leakage flow 

disturbed area as the rotor tip clearance height is increased. Rotor 2 shows the lowest percentage 

of affected area for both tip clearances and both loading conditions, which is perhaps unexpected 

given the knowledge presented thus far about latter stages ingesting the high-loss leakage flow 

from the upstream rotor(s). Figure 5.29 also shows that Rotor 1 has the least pitchwise variability 

of the leakage flow size, which aligns with the fact that the IGV has a thinner wake than Stator 1 

or Stator 2. This result can also be referred to the flow visualization photograph highlighting the 

leakage flow variability for TC1, Figure 5.9, which shows the significant difference of leakage 

flow variability for Rotor 1 compared to Rotor 2 or Rotor 3. 

Figure 5.29 also shows a similarity between Rotor1 and Rotor 2 (compared to Rotor 3) 

which has not been identified until now. At the NL condition, the increase of leakage flow 

affected area between TC1 and TC3 is an identical 10.5% for both Rotor 1 and Rotor 2, whereas 

Rotor 3 shows a smaller increase of 7.6%. At the HL condition, the same comparison is echoed 

as Rotor 1 and Rotor 2 show an identical increase of 15.3% between Rotor 1 and Rotor 2, 

whereas Rotor 3 shows a smaller increase of 10.4%. Based on these data, Rotor 3 shows a trend 

of 30% less increase of leakage flow area between TC1 and TC3 compared to Rotor 1 and Rotor 

2 for these loading conditions. 

Figure 5.29: Percentage of rotor passage affected by tip leakage flow: 

passage-averaged value and range. 

1 2 3

0

5

10

15

20

25

30

35

40

Rotor

Leakage Flow Disturbed Area [% Passage]

NL

TC1

TC3

1 2 3

0

5

10

15

20

25

30

35

40

Rotor

HL

NASA/CR—2015-218868 118
5.3. Thermal Anemometry 

The steady pressure and temperature measurements presented in Chapter 3, combined with 

the time-resolved pressure measurements above, provide valuable insight into the overall 

performance of the Purdue three-stage compressor, as well as the underlying flow physics 

governing the tip leakage flow disturbance under different rotor tip clearance conditions. 

However, a quantification of the flow angles provides an additional method by which to evaluate 

the overall effect of the rotor tip leakage flows and validate computational tools. The slanted hotwire technique implemented for these measurements provides valuable information about the 

flow features in three-dimensions, as described in Section 2.5.3. At each of the axial 

measurement planes 2 through 8 (see Figure 2.6), hot-wire data were collected at up to 28 radial 

positions across the annulus height and a series of circumferential positions across one vane pitch. 

Data were sampled at a rate of 1 MHz for 200 rotor revolutions. Several methods for evaluating 

the results are presented here which feature results to complement the measurement techniques 

already introduced. 

5.3.1. Flow Angles and Velocity 

To analyze the flow angles output from the slanted hot-wire processing technique, the 

radial profiles at each axial measurement plane are considered for two tip clearance 

configurations (TC1 and TC3). For these results, the passage-averaged flow angles were 

calculated using data collected at several measurement positions with respect to the fixed vanes 

(13 circumferential positions for rotor exit planes and 20 circumferential positions for stator exit 

planes). 

The absolute flow angles for both tip clearance configurations at the nominal loading (NL) 

operating condition are shown in Figure 5.30. As the tip leakage flow passes from the rotor 

pressure surface to the suction surface, its trajectory across the rotor passage is interpreted in the 

flow angles as underturning in the rotor relative frame of reference. However, the corresponding 

decrease of local flow velocity in the endwall region as a result of the flow blockage due to the 

tip leakage flow disturbance is significant enough to cause the absolute flow angles to increase in 

the tip region at the rotor exit (Goto, 1992). This effect is clearly observed at the rotor exit planes 

in Figure 5.30. As the rotor tip clearance height is increased, the increased flow blockage from 

the larger leakage flow region is noted as the absolute flow angle at the rotor exit planes also 

increases. 

In Figure 5.30, the Rotor 1 exit angles are distinctly different than those for the Rotor 2 and 

Rotor 3 exit angles. In fact, a comparison of the Rotor 2 and Rotor 3 exit angles for either tip 

clearance configuration, TC1 or TC3, shows nearly identical radial profile shapes and 

magnitudes. This observation aligns well with the previous notes that Rotor 1 performs 

differently than Rotor 2 and Rotor 3 in the Purdue three-stage compressor. 

At this same loading condition, the stator exit angle profiles are less affected by tip 

clearance height than the rotor exit profiles for these passage-averaged, time-averaged results. 

The Stator 1 exit results present an increase of absolute flow angle in the outer 40% annulus 

height as the rotor tip clearance height is increased, indicating that the decreased flow velocity in 

this region causes the stator loading to increase and the stator is less capable of performing the 

same amount of flow turning when the Rotor 1 tip clearance height is increased. At the Stator 2 

and Stator 3 exit planes, the curvature of the absolute flow angle profile changes as the rotor tip 

clearance height is increased, and both stators (most notably Stator 3) show a slight decrease of 

NASA/CR—2015-218868 119
flow angle in the outer 15% annulus height near the tip region. Referring to the steady detailed 

total pressure contours for Stator 2 and Stator 3 at the NL operating condition, Figure 3.39 and 

Figure 3.41, respectively, this change of flow angle is due to the slight shift of the location of 

maximum wake thickness toward the hub. 

The same passage-averaged, time-averaged radial profiles of absolute yaw angle are also 

considered at the high loading (HL) operating condition, Figure 5.31. Similar to the observations 

at the NL operating condition, these HL data also show a difference between the results at Rotor 

1 exit and those at Rotor 2 or Rotor 3 exit. At all three rotor exit planes for HL, however, the 

shift between the two tip clearance configurations is nearly twice that observed for NL. As the 

leakage flow region overtakes a more significant area at the high loading condition, Figure 5.31 

also shows a decrease of absolute yaw angle at Rotor 1 exit (due to the redistribution of mass 

flow) for the lower 70% annulus height which was less noticeable at the NL operating condition. 

Also at the HL condition, Figure 5.31 shows the reduction of the Stator 2 hub corner separation 

for TC3 compared to TC1 has a profound effect on the radial yaw angle profile in the same 

region occupying the lower 35% annulus height which was also identified in Figure 3.40. 

Figure 5.30: Radial profiles of absolute yaw angle at NL. 

30 45 60 75

0

20

40

60

80

100

α

3

 [deg]

Rotor 1 Exit

Annulus Height [%]

−10 0 10 20

0

20

40

60

80

100

α

4

 [deg]

Stator 1 Exit

30 45 60 75

0

20

40

60

80

100

α

5

 [deg]

Rotor 2 Exit

−10 0 10 20

0

20

40

60

80

100

α

6

 [deg]

Stator 2 Exit

TC1

TC3

30 45 60 75

0

20

40

60

80

100

α

7

 [deg]

Rotor 3 Exit

−10 0 10 20

0

20

40

60

80

100

α

8

 [deg]

Stator 3 Exit

NASA/CR—2015-218868 120
Figure 5.31: Radial profiles of absolute yaw angle at HL. 

The constant annulus height of the Purdue three-stage compressor lends itself to a 

simplified parallel streamline assumption, thereby neglecting the need to consider substantial 

contributions of radial flow (i.e., pitch angles) in the flowfield. However, the centrifugal effect 

imparted to the fluid by the rotors has the effect of pushing fluid outward in the radial direction. 

As a result, the flow may be expected to have a slight positive trend of pitch angles across the 

annulus height. Indeed, the NL results of passage-averaged, time-averaged pitch angle profiles, 

Figure 5.32, show a slight positive profile across the entire annulus for TC1. Throughout this 

report, the explanations of results with increased tip clearance height have been based on the 

observation of redistributed flow from the tip toward the hub. For TC3, Figure 5.32 identifies 

this trend as the pitch angles in the flow regions unaffected by the leakage flow (i.e., less than 70% 

annulus height) decrease on the order of five degrees. This net change of pitch angles between 

TC1 and TC3 confirms the theory of flow distribution from the tip toward the hub. 

At the stator exit planes, the sharp increase of pitch angle near the endwall has been 

documented by previous five-hole probe measurements in the same facility. However, this 

observation near the endwall may be attributable to the influence of the hole through which the 

measurement probe is inserted (i.e., for measurements above 95% annulus height). 

At the HL operating condition, Figure 5.33, similar trends of pitch angle with annulus 

height are observed as the rotor tip clearance height is increased from TC1 to TC3. Especially for 

the Stator 2 exit results, the considerable decrease of hub corner separation with increased tip 

clearance height is also noted in the pitch angles. 

30 45 60 75

0

20

40

60

80

100

α

3

 [deg]

Rotor 1 Exit

Annulus Height [%]

−10 0 10 20

0

20

40

60

80

100

α

4

 [deg]

Stator 1 Exit

30 45 60 75

0

20

40

60

80

100

α

5

 [deg]

Rotor 2 Exit

−10 0 10 20

0

20

40

60

80

100

α

6

 [deg]

Stator 2 Exit

TC1

TC3

30 45 60 75

0

20

40

60

80

100

α

7

 [deg]

Rotor 3 Exit

−10 0 10 20

0

20

40

60

80

100

α

8

 [deg]

Stator 3 Exit

NASA/CR—2015-218868 121
Figure 5.32: Radial profiles of pitch angle at NL. 

Figure 5.33: Radial profiles of pitch angle at HL. 

To this point, allusion has been made to the redistribution of flow from the tip toward the 

hub as the tip clearance height is increased and the corresponding increase of the tip leakage flow 

affects the flow as increased blockage. The measured steady and time-resolved pressures have 

shown the redistribution of normalized pressure rise, and the flow visualization have shown the 

effect that the increased flow near the hub can have on the separated flow regions for the 

stationary vane rows. The absolute flow angles shown here in Figure 5.30 and Figure 5.31 

provide further assurance of this trend. However, the axial velocity profiles can provide a more 

definitive representation of the mass flow rate through the machine to confirm the statements 

about redistributed mass flow suggested throughout this report. 

−5 0 5 10 15

0

20

40

60

80

100

φ

3

 [deg]

Rotor 1 Exit

Annulus Height [%]

−5 0 5 10 15

0

20

40

60

80

100

φ

4

 [deg]

Stator 1 Exit

−5 0 5 10 15

0

20

40

60

80

100

φ

5

 [deg]

Rotor 2 Exit

−5 0 5 10 15

0

20

40

60

80

100

φ

6

 [deg]

Stator 2 Exit

TC1

TC3

−5 0 5 10 15

0

20

40

60

80

100

φ

7

 [deg]

Rotor 3 Exit

−5 0 5 10 15

0

20

40

60

80

100

φ

8

 [deg]

Stator 3 Exit

−5 0 5 10 15

0

20

40

60

80

100

φ

3

 [deg]

Rotor 1 Exit

Annulus Height [%]

−5 0 5 10 15

0

20

40

60

80

100

φ

4

 [deg]

Stator 1 Exit

−5 0 5 10 15

0

20

40

60

80

100

φ

5

 [deg]

Rotor 2 Exit

−5 0 5 10 15

0

20

40

60

80

100

φ

6

 [deg]

Stator 2 Exit

TC1

TC3

−5 0 5 10 15

0

20

40

60

80

100

φ

7

 [deg]

Rotor 3 Exit

−5 0 5 10 15

0

20

40

60

80

100

φ

8

 [deg]

Stator 3 Exit

NASA/CR—2015-218868 122
These axial velocity profiles are shown in Figure 5.34 for the axial measurement planes at 

the exit of Rotor 2 and Rotor 3, which show the specific effect of the tip leakage flow on the 

rotor exit flow fields which will include the important multistage effects. The profiles in Figure 

5.34(a-b) correspond to the nominal loading and high loading conditions, respectively, and the 

data have been normalized in all cases by the corresponding tip speed, . 

The results in Figure 5.34 show several interesting trends at both loading conditions. In 

particular, a comparison of the two rotors at NL or HL shows qualitatively similar profiles for 

each of the two tip clearance configurations. At the NL condition, Figure 5.34(a), there is a 

noticeable decrease of axial velocity in the tip region for both rotors and both tip clearances. 

However, the velocity profile has a profile across the span which resembles a parabolic profile. 

This contrasts with the HL condition, Figure 5.34(c), which shows a noticeable decrease of axial 

velocity near the tip (outer 20-30% span), but a more constant axial velocity profile across the 

lower 70-80% span. A comparison with the time-resolved total pressure measurements in Figure 

5.24 through Figure 5.27 shows that this change of velocity profile is due to the thicker rotor 

wake at the high loading conditions, for which the separation off the suction surface of the rotor 

blade becomes more significant across a more substantial portion of the span. 

(a) NL (b) HL 

Figure 5.34: Radial profiles of normalized axial velocity at Rotor 2 and Rotor 3 exit. 

5.3.2. Blockage 

In addition to the flow angles presented in the previous section, the flow blockage at each 

axial measurement plane has been quantified using the method outlined by Suder (1998). As 

discussed in Section 2.5.3.3, the ambient temperature during hot-wire experiments has 

potentially introduced an additional inadvertent error in the data which prevents direct 

comparison of the velocity magnitudes from one test to the next. However, the blockage 

parameter is an alternative presentation method which relies less on the absolute magnitude of 

0.1 0.3 0.5 0.7

0

20

40

60

80

100

Vx,5 / Ut [−]

Rotor 2 Exit

Annulus Height [%]

0.1 0.3 0.5 0.7

0

20

40

60

80

100

Vx,7 / Ut [−]

Rotor 3 Exit

TC1

TC3

0.1 0.3 0.5 0.7

0

20

40

60

80

100

Vx,5 / Ut [−]

Rotor 2 Exit

Annulus Height [%]

0.1 0.3 0.5 0.7

0

20

40

60

80

100

Vx,7 / Ut [−]

Rotor 3 Exit

TC1

TC3

NASA/CR—2015-218868 123
the measurements for calculation. Thus, the blockage data presented here serve as a more 

meaningful method for comparing the effective flow velocities. 

Suder (1998) defines the blockage quantification method as the effective reduction in flow 

area: 

 = 1 −

effective	area

geometric	area

. (5.3)

Equation (5.3) is further defined by evaluating the effective area as a function of the velocitydensity deficit in the flow region, 

∗

: 

 = 1 −

 − ï 

∗ñ



,

(5.4)

for which  represents the geometric flow area. The velocity-density deficit is analogous to the 

displacement thickness introduced through fundamental boundary layer theory: 



∗ = ó o1 −

"ò

>"òAinviscid

r ñ

;ö/NB

U

, (5.5)

where NB is the number of blades (or vanes) in the row of interest. 

For these data, the blockage can be calculated at the rotor exit planes by analyzing timeseries measurements. The time series is first phase-locked ensemble-averaged, then passageaveraged across one vane pitch, and finally averaged over each of the blades in one row; this 

final product represents a mean rotor wake at the given measurement plane and loading condition. 

Similarly, the same blockage parameter can be calculated in a spatial reference frame (instead of 

a time series) at stator exit planes, for example. At these locations, the hot-wire signal is timeaveraged and then presented as a function of pitch-wise location. 

In either case, the physical method described by Suder (1998) is the same: determine a 

defect region via axial velocity gradients and the selection of an arbitrary cutoff value. For this 

study, the cutoff value was chosen to be 3 s-1 for all conditions. In agreement with Suder’s 

conclusion, the arbitrary selection of the cutoff values has a negligible effect on the result (less 

than 5% change of calculated blockage values). Several cutoff values in the range of 1-5 s-1 were 

considered for these data. Examples of the axial velocity data and the corresponding defect 

region are shown in Figure 5.35 and Figure 5.36 for rotor exit data and stator exit data, 

respectively. 

NASA/CR—2015-218868 124
Figure 5.35: Example of defect region identification for Rotor 2 exit (TC1, NL). 

Figure 5.36: Example of defect region identification for Stator 1 exit (TC1, NL). 

Non−Dimensional Time, t/BPP [−]

Annulus Height [%]

Axial Velocity

0 0.2 0.4 0.6 0.8 1

0

10

20

30

40

50

60

70

80

90

100

 Vx [ft/s]

90 

110 

130 

150 

170 

190 

210 

230 

250 

Defect Region

Non−Dimensional Time, t/BPP [−]

Annulus Height [%]

0 0.2 0.4 0.6 0.8 1

0

10

20

30

40

50

60

70

80

90

100

Stator 1 Vane Passage [%]

Annulus Height [%]

Axial Velocity

0 20 40 60 80 100

0

10

20

30

40

50

60

70

80

90

100

 Vx [ft/s]

90 

110 

130 

150 

170 

190 

210 

230 

250 

Annulus Height [%]

Defect Region

Stator 1 Vane Passage [%]

0 20 40 60 80 100

0

10

20

30

40

50

60

70

80

90

100

NASA/CR—2015-218868 125
The calculated blockage values are presented in Figure 5.37 for the nominal loading (NL) 

condition and Figure 5.38 for the high loading (HL) condition. A summary of these results is also 

listed in Table 5.1 as a representative one-dimensional blockage value for each axial 

measurement position, loading condition, and tip clearance configuration. A preliminary analysis 

of the results in Figure 5.37 show the most significant changes in blockage are observed in the 

tip region at Rotor 1 exit, Stator 2 exit, and Stator 3 exit for the two tip clearance configurations 

investigated here (TC1 and TC3). Considering the total pressure measurements at Rotor 1 exit in 

Figure 5.22, it is not surprising that there is a more significant blockage region due to the 

increased tip leakage flow disturbance. However, it is particularly interesting that the same 

disturbance does not appear as significantly in the calculated blockage for the Rotor 2 and Rotor 

3 exit data. Although the Rotor 2 exit data do show an increase of blockage on the order of 10% 

in the region from 80 to 90% annulus height, the Rotor 1 exit data are nearly 20% different 

between the two tip clearance configurations across the region spanning 80 to 100% annulus 

height. 

The Stator 2 exit total pressure field, Figure 3.39, reveals that the increased blockage in the 

tip region at the same location in Figure 5.37 is due to largely to low pressure regions near the 

wall, not the wake itself. On the other hand, the Stator 3 exit total pressure field (Figure 3.41) 

shows a thicker wake in the outer 50% of the flow annulus between the two tip clearance 

configurations. However, this blockage effect is only identified for the outer 30% of the flow 

annulus. 

As the compressor is throttled toward stall, the similarity between the observed blockage 

trends at each rotor exit plane increases, Figure 5.38. Of the three rotors, Rotor 3 actually shows 

a decrease of blockage in the tip region with increased loading for TC1, a result which is not 

surprising considering the difference of Rotor 3 exit measurements at the same conditions 

presented earlier. In Figure 5.26 and Figure 5.27, the mean wake contours of RMS total pressure 

show the leakage flow region is decreased at high loading, but at the expense of a very thick 

wake at HL which extends across a majority of the passage. Based on the calculated values in 

Table 5.1, the more intense fluctuating wake region extending across the blade span is significant 

enough to offset the blockage created by the leakage flow at NL (and more). Also at the three 

rotor exit planes, the blockage differences between the two tip clearance configurations approach 

30% and the radial extent begins at approximately 70% annulus height for all rotors – a 

similarity which was not observed as clearly for the three rotors at the NL condition. 

For the high loading condition, a comparison of the two tip clearance configurations at 

Stator 2 exit confirms the expectations from the steady total pressure traverses shown previously 

in Figure 3.40 and flow visualization photographs. Although the Stator 2 exit blockage in Figure 

5.38 increases across the outer 60% annulus height (on the order of 10% increased blockage), the 

reduction of the corner separation region off the suction surface of the vane is represented by the 

decrease of blockage in the lower 40% annulus height (on the order of 10% decreased blockage). 

NASA/CR—2015-218868 126
Figure 5.37: Radial profiles of blockage at NL. 

Figure 5.38: Radial profiles of blockage at HL. 

Table 5.1: Summary of one-dimensional blockage parameters, listed in percent flow area. 

TC Config. Loading R1 Exit S1 Exit R2 Exit S2 Exit R3 Exit S3 Exit 

TC1 NL 9.1 6.8 9.3 7.4 8.4 6.2 

HL 10.2 10.0 11.8 12.4 11.8 12.3 

TC3 NL 15.5 9.4 9.0 10.9 7.6 8.5 

HL 16.6 9.4 16.4 15.4 13.3 14.2 

0 20 40 60 80

0

20

40

60

80

100

B3 [%]

Rotor 1 Exit

Annulus Height [%]

0 20 40 60 80

0

20

40

60

80

100

B4 [%]

Stator 1 Exit

0 20 40 60 80

0

20

40

60

80

100

B5 [%]

Rotor 2 Exit

TC1

TC3

0 20 40 60 80

0

20

40

60

80

100

B6 [%]

Stator 2 Exit

0 20 40 60 80

0

20

40

60

80

100

B7 [%]

Rotor 3 Exit

0 20 40 60 80

0

20

40

60

80

100

B8 [%]

Stator 3 Exit

0 20 40 60 80

0

20

40

60

80

100

B3 [%]

Rotor 1 Exit

Annulus Height [%]

0 20 40 60 80

0

20

40

60

80

100

B4 [%]

Stator 1 Exit

0 20 40 60 80

0

20

40

60

80

100

B5 [%]

Rotor 2 Exit

TC1

TC3

0 20 40 60 80

0

20

40

60

80

100

B6 [%]

Stator 2 Exit

0 20 40 60 80

0

20

40

60

80

100

B7 [%]

Rotor 3 Exit

0 20 40 60 80

0

20

40

60

80

100

B8 [%]

Stator 3 Exit

NASA/CR—2015-218868 127
5.3.3. Streamwise Vorticity 

The streamwise vorticity provides an additional method for evaluating the leakage flow 

region once the three-dimensional velocity vectors are known. Because the ensemble-average 

technique has been used to calculate the velocity vectors, the presence of an instantaneous vortex 

cannot be evaluated. However, this technique can still determine the average rotational motion in 

the leakage flow region. 

Inoue et al. (1986) define the streamwise vorticity by calculating the projection of two 

vorticity components on the streamwise flow direction (defined by the relative flow angle, ): 

< = ù

1



û>ü

A

û −

1



û†

û ° cos  − ¢

1



û†

û −

ûB

û £ sin  , (5.6)

where the first bracketed terms represents the axial vorticity, B, and the second bracketed term 

represents the tangential vorticity, ü. Unlike the measurements collected by Inoue et al., the 

data collected in this study were only captured at one axial survey plane downstream of each 

rotor. As a result, the gradient of radial absolute velocity, û†⁄û, cannot be evaluated for these 

measurements. Future studies may build in the ability to evaluate this change in the axial flow 

direction. However, if the contributions from this gradient are assumed to be small with respect 

to the other components, they can be neglected. Under this assumption, the streamwise vorticity 

is simplified: 

< = ù

1



û>ü

A

û −

1



û†

û ° cos  − ¢−

ûB

û £ sin  . (5.7)

The derivatives needed for Equation (5.7) were numerically calculated using a central 

differencing scheme, with exception to the edges of the data region, which were calculated as 

single-sided differences. 

Figure 5.39 shows contours of streamwise vorticity, as calculated from Equation (5.7), for 

two different tip clearance configurations and two loading conditions at the Rotor 1 exit plane. 

The calculated vorticity results in Figure 5.39 are normalized by the angular velocity of the 

rotor, Ω. The leakage flow structures identified in Figure 5.39 are very similar to the regions 

identified by the passage-averaged time-resolved total pressure measurements in Figure 5.22 and 

Figure 5.23. Specifically, the leakage flow for TC1 extends downward to approximately 80% 

span, whereas the leakge flow for TC3 extends to approximately 70% span for NL and 65% span 

for HL. However, Figure 5.39 also shows that the regions of high vorticity typically associated 

with the tip leakage vortex (recall, these data are ensemble-averaged and passage-averaged, so 

they do not represent an instantaneous vortex) does not stay attached to the wall, but moves 

radially downward into the passage. These results suggest that, although the total pressure 

unsteadiness (RMS) at the rotor exit plane is an adequate method for identifying the leakage flow 

region, those results may not sufficiently represent any true vortex structures. 

Further analysis of the contours in Figure 5.39 shows the difference of the wake region for 

TC1 versus TC3. Inoue et al. (1986) also identified these regions of positive and negative 

vorticity associated with the passing rotor wake, an observation which was attributed to the 

radial velocities induced by the centrifugal effects of the passing rotor. However, there is a 

significant reduction of both the positive and negative regions of vorticity in the rotor wake as 

the tip clearance height is increased from TC1 to TC3. This observation can be linked to the 

redistribution of flow velocity toward the hub, which energizes the separation-prone fluid on the 

rotor suction surface and effectively reduces the strength of the wake region (although 

NASA/CR—2015-218868 128
comparisons of the yaw angle wake profiles at several spanwise locations, not shown here, 

suggest the width of the wake may not necessarily become narrower). Indeed, a deeper 

investigation of the contribution from each of the three velocity gradients in Equation (5.7) 

shows that the û†⁄û component of the streamwise vorticity has a peak value which not only 

changes magnitude, but also its circumferential position, leading to the change of the wake 

vorticity. This observed trend is in contrast to the other two components in Equation (5.7) which 

have peak values that are unaffected in their circumferential position and show a change only in 

their magnitude. 

Figure 5.39: Contours of streamwise vorticity at Rotor 1 exit for TC1 and TC3. 

In addition to these results for Rotor 1, the streamwise vorticity is also presented for TC1 at 

the exit of Rotor 2 and Rotor 3 in Figure 5.40. These results confirm many of the previous 

analyses suggesting the Rotor 2 and Rotor 3 may behave similarly to each other, but different 

from Rotor 1. In particular, a qualitative comparison of the streamwise vorticity contours in 

Figure 5.40 shows leakage flow vorticity structures which are similar in size and shape. A 

primary difference here is the results at Rotor 3 exit for both NL and HL extend downward 

PS SS

Annulus Height [%]

TC1, NL

0 0.2 0.4 0.6 0.8 1

0

10

20

30

40

50

60

70

80

90

100

PS SS

TC3, NL

0 0.2 0.4 0.6 0.8 1

0

10

20

30

40

50

60

70

80

90

100

PS SS

t / BPP [−]

Annulus Height [%]

TC1, HL

0 0.2 0.4 0.6 0.8 1

0

10

20

30

40

50

60

70

80

90

100

PS SS

t / BPP [−]

TC3, HL

0 0.2 0.4 0.6 0.8 1

0

10

20

30

40

50

60

70

80

90

100

ωst / Ω [−]

−15

−12

−9

−6

−3

0

3

6

9

12

15

NASA/CR—2015-218868 129
approximately 5% deeper into the passage than the results at Rotor 2 exit. This difference 

compares well with the leakage flow regions identified by high unsteadiness in the time-resolved 

total pressures, Figure 5.24 through Figure 5.27, and similar comparisons can be drawn between 

the rotor wake shapes at the two measurement locations. 

Figure 5.40: Contours of streamwise vorticity at for TC1 at Rotor 2 and Rotor 3 exit. 

To supplement the streamwise vorticity contours shown here, the maximum vorticity 

strength has been quantified for each of the test conditions. The vorticity contour plots (e.g, 

Figure 5.39 and Figure 5.40) were used to locate the region of the blade passage which 

corresponds to the leakage flow. In this region, a search tool identified the largest negative value 

(based on the coordinate system in the above figures). 

This information, presented in Figure 5.41, summarizes the streamwise vorticity results 

from all of the test conditions to one-dimensional values for comparison. These data show that as 

the rotor tip clearance is increased from TC1 to TC3, the maximum vorticity strength in the 

leakage flow region increases for Rotor 2 and Rotor 3, but decreases for Rotor 1. A comparison 

of the Rotor 1 vorticity contours for TC1 in Figure 5.39 shows that the change of the shape of the 

PS SS

Annulus Height [%]

Rotor 2, NL

0 0.2 0.4 0.6 0.8 1

0

10

20

30

40

50

60

70

80

90

100

PS SS

Rotor 3, NL

0 0.2 0.4 0.6 0.8 1

0

10

20

30

40

50

60

70

80

90

100

PS SS

t / BPP [−]

Annulus Height [%]

Rotor 2, HL

0 0.2 0.4 0.6 0.8 1

0

10

20

30

40

50

60

70

80

90

100

PS SS

t / BPP [−]

Rotor 3, HL

0 0.2 0.4 0.6 0.8 1

0

10

20

30

40

50

60

70

80

90

100

ωst / Ω [−]

−15

−12

−9

−6

−3

0

3

6

9

12

15

NASA/CR—2015-218868 130
high negative vorticity region identified as the leakage flow from NL to HL is more dramatic 

than any of the other conditions shown here. This change of the leakage flow shape may 

contribute to the different for Rotor 1 at NL in Figure 5.41. Finally, as the loading condition 

changes from NL to HL, there is a consistent decrease of maximum streamwise vorticity strength. 

This trend shows an inverse proportionality of leakage flow size (Figure 5.29) with maximum 

vorticity strength in the leakage flow region which holds for a change in loading condition, but 

not a change in tip clearance height. The results in Figure 5.41 have been calculated for the 

passage-averaged vorticity, as shown in the contour plots above, but follow-on work will analyze 

the change of maximum streamwise vorticity strength with the stator-rotor interaction. This 

future analysis will allow the identification of a range for the bars in Figure 5.41, similar to the 

ranges defined in Figure 5.29. 

Figure 5.41: Passage-averaged maximum streamwise vorticity strength. 

1 2 3

0

4

8

12

16

20

Rotor

Maximum Streamwise Vorticity Strength, |

ωst

/Ω|max [−]

NL

TC1

TC3

1 2 3

0

4

8

12

16

20

Rotor

HL

NASA/CR—2015-218868 131
CHAPTER 6: PARTICLE IMAGE VELOCIMETRY 

Sayantan Bhattacharya and Pavlos P. Vlachos 

The objective of the PIV aspect of this project was to develop a technique to acquire data in 

the tip flow region of an embedded rotor in a multistage compressor. The unique aspect of this 

work is that the same window used to acquire the images was also used to deliver the laser 

illumination to the flow field, thus eliminating the need for inserting a periscopic probe to the 

flow field. The successful implementation of this technique at two loading conditions (nominal 

and high loading) for the TC3 tip clearance configuration is discussed below. 

The three-component vector fields obtained by generalized stereo reconstruction at each 

circumferential location were analyzed. The recorded stereo image coordinate system was 

reoriented and scaled to express the data in terms of the coordinate system defined by axial chord, 

span, and blade passage. The blade tip velocity was subtracted from the circumferential velocity 

component to change from absolute frame of reference to compressor blade coordinate system. 

The flow direction and the coordinate system are expressed in the following schematic (Figure 

6.1). 

Figure 6.1: Schematic of the PIV measurement plane and corresponding coordinate system. 

6.1. Stereo Results 

The twenty measurement planes containing the three-component planar velocity fields 

were combined to reconstruct the volumetric vector field across the blade passage. The effective 

domain was 70% to 96% span, 15% to 90% axial chord and 100% blade passage. The velocity 

field was smoothed with a Gaussian kernel of 2 standard deviations and a window size of 7x7 

NASA/CR—2015-218868 132
grid points to reduce the noise in the flow field. The volume of data was then sliced in constant 

span-wise locations as shown in the following figures. Although the measurements were across 

one complete blade passage, for more intuitive representation the domain was periodically 

repeated along the blade passage direction. 

The radial velocity contours in Figure 6.2 show a radial inflow and outflow represented by 

the blue and the red regions, respectively. For the nominal loading case, the alternating red and 

blue region close to the suction side of the blade is indicative of the tip leakage flow and this 

structure expands as the flow passes through the blade passage. A radial outflow is also observed 

on the pressure side close to the blade tip and within 20% axial chord, which essentially rolls up 

on the suction side and forms the tip leakage vortex. However, for the high loading case, little 

radial inflow is observed close to the tip of the blade within 20% axial chord and the radial 

outflow on the pressure side of the blade is more dominant throughout the passage flow at the 95% 

spanwise slice. At high loading, the radial inflow is observed at spanwise slices away from the 

tip. 

Figure 6.2: Spanwise volume slices of radial velocity for stereo reconstructed velocity field at 

two loading conditions. 

NASA/CR—2015-218868 133
region with negative radial velocity is present from 95% to 75% span and extends through 

almost 70% of the blade passage. The consecutive red and blue region indicating the tip leakage 

flow is present on the suction side of the blade up to 84.5% span, and after that, the secondary 

flow breaks into small regions of radially inward flow. The positive relative radial velocity, †, 

near the pressure side of the blade within 20% axial chord is very strong at the 95% span location 

and decreases gradually towards the hub. 

Figure 6.3: Contours of radial velocity at different span-wise location for NL. A vector field 

representing the projected velocity in the  −  plane is superimposed over the contours. Flow is 

from right to left. 

NASA/CR—2015-218868 134

The radial velocity at each spanwise location is examined in detail for both loading 

conditions. Figure 6.3 shows the spanwise slices for the nominal loading condition. The blue 
For the high loading condition, a strong radial outflow is dominant throughout the spanwise 

slices on the pressure side of the blade and within first 25% of the axial chord (Figure 6.4). Also, 

the blade passage is dominated by a strong positive radial component (red and yellow contours) 

with a few blue regions in between up to 88% span. From 95% span to 75% span, the secondary 

flow towards the hub increases for the high loading case. 

Figure 6.4: Contours of radial velocity at different span-wise location for HL. A vector field 

representing the projected velocity in the  −  plane is superimposed over the contours. Flow is 

from right to left. 

The vectors in both Figure 6.3 and Figure 6.4 are projections of the three-dimensional 

velocity vector onto the  −  plane. The relative velocity vectors in the blade frame of reference 

reveal the primary flow from right to left following the staggered blade passage. The vectors are 

NASA/CR—2015-218868 135
physically consistent in the middle of the passage, except along the edge of the blades and the 

side of the domain of measurement. The presence of spurious vectors in this region is mainly an 

artifact of the smoothing operation along the blade edges where no vectors were initially present. 

6.2. Tomographic Results 

In addition to the stereo results presented above in Figure 6.2 through Figure 6.4, 

tomographic reconstructions were also created, as described in Section 2.6.2. The tomographic 

PIV vector fields were also interpolated and smoothed using Gaussian smoothing. The three 

dimensional vector fields were obtained for twenty phase locked positions along the 

circumference with a 15% overlap between the volumes. The individual volumes were stacked 

up to represent the flow field within the blade passage, and the spanwise slices in the 

measurement volume are shown in Figure 6.5. 

Figure 6.5: Spanwise volume slices of radial velocity for tomographic reconstructed velocity 

field at two loading conditions. 

The results for both loading conditions show adjacent radial inflow and outflow region on 

the suction side of the blades and this structure extended throughout the blade passage. 

Comparing the results with the stereo reconstructed measurements, a few differences are 

NASA/CR—2015-218868 136
observed. First, the strong radial outflow on the pressure side of the blade within 20% axial 

chord, which was observed in the stereo results, is not present in the tomographic PIV results. 

Second, for the high loading case, although the radial outflow is still dominant in the slices 

towards the tip at 95% span, the radial inflow (blue) region is more prominent for this particular 

case even at 95% span as compared to the stereo results. Despite minor differences, overall the 

tomographic PIV results are similar to the ones obtain with stereo-PIV. However, the quality of 

the tomographic PIV solution greatly depends on the tomographic reconstruction and eliminating 

the “ghost” or false reconstructed particles. Thus, further efforts are necessary to achieve a more 

robust volumetric reconstruction and to minimize the noise in the output vector fields. 

6.3. Summary 

These results show the capability to do three-dimensional volumetric PIV measurements in 

a high-speed multistage compressor without introducing any probe in the measurement volume. 

All the previous PIV measurements to date have introduced a light sheet probe which inevitably 

creates a disturbance in the downstream rotor flow region. Here, the laser sheet was introduced 

through a glass window on the casing and the cameras were focused on a 4mm thick laser sheet 

through the same window. Reflection was overcome by using fluorescent particles and filters in 

front of the lenses. Phase-locked measurements were taken in one complete blade passage. Using 

multiple camera information both stereoscopic and tomographic PIV velocity fields were 

obtained. 

NASA/CR—2015-218868 137
CHAPTER 7: SUMMARY AND CONCLUSIONS 

Reid A. Berdanier and Nicole L. Key 

The effect of rotor tip clearances on axial compressor performance has been a focus of 

research for several decades. In general, studies have found that pressure rise capability, 

efficiency, and operability range all decrease as the rotor tip clearance height is increased. 

However, the future of gas turbine engine engineering is moving toward designs which will 

incorporate smaller blade heights in the rear stages of high pressure compressors. As a result, a 

decrease in blade heights and a corresponding increase in relative rotor tip clearances are 

expected in the rear stages of these next-generation compressors. Therefore, a better 

understanding of the fundamental flow physics and multistage performance effects related to 

large tip clearance heights in axial compressors is a necessity. 

7.1. Overview of Methods and Findings 

This work has investigated the effects of large rotor tip clearances on the performance of a 

three-stage axial compressor at Purdue University. A series of experimental measurement 

techniques has been implemented to evaluate compressor performance and interrogate the tip 

leakage flow for three rotor tip clearances: 1.5%, 3.0%, and 4.0% based on annulus height. 

Benchmark compressor performance measurements showed the pressure rise capability and 

isentropic compressor efficiency through the machine decrease approximately linearly with rotor 

tip clearance height. Measurements with seven-element total pressure and total temperature rakes 

inserted at inter-stage locations in the compressor showed a redistribution of flow from the tip 

toward the hub as the tip clearance height was increased. This effect is due to the increased 

blockage related to the tip leakage flow. 

Stall inception evaluated using piezoresistive pressure transducers distributed 

circumferentially around the compressor, approximately 10% axial chord upstream of the rotor 

leading edges. These measurements showed the compressor stalls as a result of long length-scale 

modal disturbances for the 100% corrected speedline with the 1.5% tip clearance configuration. 

Previous research has suggested modal stall characteristics may be expected when the individual 

stages of a multistage compressor are well-matched. At part-speed operating conditions, or when 

the tip clearance height was increased, however, the compressor was more likely to exhibit 

spike-type stall inception mechanisms emanating from Rotor 1. 

Additional time-resolved measurements of total pressure at the rotor exit planes and hotwire measurements of three-dimensional flow velocities were collected for two tip clearance 

configurations at two loading conditions on the 100% corrected speedline: a nominal loading 

point near the peak efficiency point and a low flow rate high loading point. By casting the rotor 

exit dynamic total pressures in terms of an unsteadiness parameter (the RMS with respect to the 

ensemble average), the radial and circumferential extent of the tip leakage flow disturbance was 

evaluated. These measurements, in addition to flow visualization images, helped to identify a 

modulation of the tip leakage flow, affecting its size and depth of penetration into the main flow, 

as the rotor interacts with the wakes propagating from the upstream stator. Also, these timeresolved velocity data were also used to quantify the blockage at each measurement plane and 

NASA/CR—2015-218868 138
correlate the blockage with the observed tip leakage flow disturbance. Additionally, streamwise 

vorticity calculations were used to characterize the strength of the leakage flow from each rotor. 

Finally, a novel three-dimensional PIV measurement technique has been developed for 

application in multistage compressors. This new technique circumvents the need to introduce 

intrusive probes for light sheet delivery. Instead, the laser sheet was introduced through the same 

optical access window used to acquire the images. Issues with reflections from the incident light 

were overcome through the use of fluorescent seed particles and appropriate optical filters on the 

cameras. Phase-locked measurements were collected using four cameras in stereo configurations 

to reconstruct planar velocity fields, as well as an out-of-plane component. Initial tomographic 

reconstructions were performed, as well. 

The measurements collected for this project represent a unique data set which contributes 

to better understanding the tip leakage flow field and its associated loss mechanisms. These data 

will serve the community as a method for validating computational design tools, especially at 

off-design conditions. Through this process, the results presented herein will aid in the 

development of new blade designs which could be desensitized to rotor tip leakage flows and 

their associated performance decrements. 

7.2. Recommendations for Future Research 

The results presented herein provide a wealth of data which will improve the understanding 

of tip leakage flow effects in axial compressors. However, there is still significant processing that 

can be performed to glean additional information from the collected data. For the time-resolved 

measurements, in particular, the results shown here have only analyzed the time domain, and 

further analysis of the frequency content may unveil more information, particularly relating to 

the multistage effects of the machine and any potential aeromechanical forcing. 

The detailed flow field traverses (steady total pressure, time-resolved total pressure, timeresolved velocities) were conducted only for the smallest and largest tip clearance configurations 

at the nominal loading and high loading operating conditions. Thus, there is a multitude of 

additional data which could be collected at additional loading conditions (i.e., a negative 

incidence, low loading condition) or at the intermediate tip clearance height. The measurements 

collected from a third tip clearance could accommodate a more thorough analysis of trends 

associated with vorticity, leakage flow size, etc. with changing tip clearance. 

The non-intrusive PIV technique implemented here is the first of its kind to introduce the 

laser illumination through the same optical access window as the imaging, thereby avoiding 

intrusive periscope methods. With this proof-of-concept, the opportunity is available to nonintrusively measure the compressor flow field in each of the three rotors for three tip clearance 

heights and at several loading conditions. These non-intrusive measurements help to fill in the 

void left by the other traditional measurement techniques by capturing data in the rotor passage, 

but they also provide the opportunity to validate measurements and gain additional confidence in 

measurements from all techniques. 

Finally, the understanding of the rotor tip leakage flow field and its impact on the 

compressor performance developed through this study has created a unique opportunity to 

investigate leakage flow control techniques (such as casing treatments) and other leakage flow 

desensitizing design methodologies in this three-stage axial compressor. Ultimately, it is these 

technologies that will help to drive the future of robust compressor designs. 

NASA/CR—2015-218868 139
LIST OF PUBLICATIONS RESULTING FROM THIS PROJECT 

[1] Berdanier, R.A., Smith, N.R., Fabian, J.C., and Key, N.L., 2015, “Humidity Effects on 

Experimental Compressor Performance—Corrected Conditions for Real Gases,” Journal of 

Turbomachinery, 137(3), 031011 (10 pages). 

[2] Berdanier, R.A. and Key, N.L., 2015, “Experimental Investigation of Factors Influencing 

Operating Rotor Tip Clearance in Multistage Compressors,” International Journal of 

Rotating Machinery, Article ID 146272 (13 pages). 

[3] Smith, N.R., Berdanier, R.A., Fabian, J.C., and Key, N.L., 2015, “Reconciling Compressor 

Performance Differences for Varying Ambient Inlet Conditions,” PowerEnergy2015-49102. 

To be presented at the ASME Power and Energy Conference, San Diego, CA, June 26-July 

2, 2015. 

[4] Berdanier, R.A. and Key, N.L., “Tip Leakage Flow Effects on Multistage Compressor 

Performance for Small Core Engine Applications,” In review for Journal of 

Turbomachinery. 

[5] Smith, N.R. and Key, N.L., “Flow Visualization for Investigating Stator Losses in a 

Multistage Axial Compressor,” In review for Experiments in Fluids. 

[6] Berdanier, R.A. and Key, N.L., “Data Processing Techniques for Measurements Collected 

from a Single Slanted Hot-Wire,” To be presented at the AIAA Fluid Dynamics 

Conference, Dallas, TX, June 22-26, 2015. 

NASA/CR—2015-218868 140
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