DESIGN, MANUFACTURING AND TESTING OF A WOUND

COMPOSITE AXIAL COMPRESSOR WITH INTEGRATED SHROUD

By

Thomas J Qualman II

A DISSERTATION

Submitted to

Michigan State University

in partial fulfillment of the requirements

for the degree of

Mechanical Engineering—Doctor of Philosophy

2018

ABSTRACT

DESIGN, MANUFACTURING AND TESTING OF A WOUND

COMPOSITE AXIAL COMPRESSOR WITH INTEGRATED SHROUD

By

Thomas J Qualman II

There are a vast number of applications in which composite axial turbomachines could

prove beneficial, however, current manufacturing methods can be time or cost prohibitive.

Over the past several years, a novel manufacturing process aimed at reducing costs and

improving structural properties has been under development at Michigan State University.

This prior work established the basic manufacturing concepts and demonstrated the feasi-

bility of utilizing the novel wound composite impellers for a number of applications in the

renewable energy sector. The features of scalability and modularity put forth by this pre-

vious work have been promising. The resulting geometries that could be produced were

limited to variations of the unconventional “star” pattern design. These designs generate

approximately a solid body (i.e. forced vortex) swirl distribution.

This work reflects a substantial improvement of the manufacturing process for wound

composite impellers, while also allowing the manufacturing of any modern 3D shaped im-

pellers with preferred swirl distributions. This was achieved through sophistication of the

mandrel (mold) design and fiber layup process. By dry winding the fiber and then utiliz-

ing a vacuum assisted resin transfer molding method for the infusion process, high quality

impellers were obtained. Both front and rear stage impellers of a multistage axial com-

pressor were manufactured thereby demonstrating the flexibility of the new manufacturing

process. Additionally, shaft driven and integrated drive motor configurations were designed

and demonstrated. Overlaying a laser scanned model onto the CAD model of an impeller

was used to visually inspect how well the resulting blade shapes matched the original de-

sign and the preliminary data indicated good agreement. Measurements of basic geometric

dimensions like shroud diameter, tip radius, hub radius at leading/trailing edges and axial

chord length were within 20 % agreement with their design values with most below 10 % and

several within 1 % agreement.

To my parents, for all the sacrifices you’ve made to provide me with all of the support

and resources needed to make this possible.

iv

ACKNOWLEDGMENTS

I would like to thank my advisor, Dr. Norbert M¨uller, for initially hiring me as an under-

graduate student and seeing enough potential in me to not only encourage me to attend

graduate school but to offer me an opportunity to do this under his guidance. Without this,

I definitely would not be where I am today so to him I offer my sincerest gratitude. I would

also like to thank my other committee members Dr. Abraham Engeda, Dr. Neil Wright and

Dr. Wei Liao for their assistance and guidance throughout this process.

Other people that were critical to the success of this work were Younis Najim, Zack Hoyle,

Blake Gower, Andy Vander Klok, Johannes Pohl and Michael Jaehnig. Thank you for all

of your help and camaraderie over the years. Working with you all in the Turbomachinery

Laboratory has really been a joy and has helped to make my experience at MSU a great one.

Lastly, I would like to thank my parents. Without them none of this would have been

possible. Whether emotional, financial or whatever I needed you would do what you could to

help me with whatever problem I may have been facing, you’ve truly given me unconditional

love and support and for that I am eternally grateful. I also need to thank my girlfriend,

Andrea, for her patients in all of the years I’ve had to spend so much time focused on my

education. I know it hasn’t always been easy and I thank you for being there to love and

support me throughout this journey.

v

TABLE OF CONTENTS

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

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

ix

xi

KEY TO SYMBOLS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . xviii

KEY TO ABBREVIATIONS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . xxv

Chapter 1

Introduction and Objectives . . . . . . . . . . . . . . . . . . . . .
1.1 Dissertation Overview . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
1.2 Background . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
1.2.1 Axial Turbomachinery . . . . . . . . . . . . . . . . . . . . . . . . . .
1.2.2 Composites
. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
1.3 Previous Work on Wound Composite Rotors . . . . . . . . . . . . . . . . . .
1.3.1 Water as a Refrigerant Application . . . . . . . . . . . . . . . . . . .
1.3.2 Geothermal NCG Removal . . . . . . . . . . . . . . . . . . . . . . . .
1.3.3 Marine/Hydro Turbine Application . . . . . . . . . . . . . . . . . . .
1.3.4 Other Previous Work . . . . . . . . . . . . . . . . . . . . . . . . . . .
1.4 Research Objectives . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

Chapter 2

2.2.1 Experimental Testing Facilities
2.2.2

Scope of Work . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
2.1
Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
2.2 Design Framework . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
. . . . . . . . . . . . . . . . . . . . .
Integrated Permanent Magnet Motor . . . . . . . . . . . . . . . . . .
2.3 Scope of Work . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
Impeller Design . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
2.3.1
2.3.2
Impeller Manufacturing
. . . . . . . . . . . . . . . . . . . . . . . . .
2.3.3 Experimental Testing . . . . . . . . . . . . . . . . . . . . . . . . . . .
Spin Pit Testing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
2.3.4

Chapter 3 Compressor Design . . . . . . . . . . . . . . . . . . . . . . . . . . .
3.1 Axial Compressor Design . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
3.1.1
Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
3.1.2 Axial Compressor Basics . . . . . . . . . . . . . . . . . . . . . . . . .
3.1.3 Axial Compressor Design Methodology . . . . . . . . . . . . . . . . .
3.2 Rotor Design . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
3.2.1 Design Criteria . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
3.2.2 Results . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

vi

1
1
1
1
4
7
7
15
25
31
32

35
35
36
36
37
38
38
40
40
41

42
42
42
43
55
63
63
65

4.1

Chapter 4 Numerical Investigation . . . . . . . . . . . . . . . . . . . . . . . .
Introduction to CFD . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
4.1.1 Governing Equations . . . . . . . . . . . . . . . . . . . . . . . . . . .
4.1.2 Reynolds-Averaged Navier-Stokes Equations . . . . . . . . . . . . . .
4.1.3 Turbulence Modeling . . . . . . . . . . . . . . . . . . . . . . . . . . .
4.2 Compressor Modeling . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
4.3 Meshing of Fluid Domain . . . . . . . . . . . . . . . . . . . . . . . . . . . .
4.4 CFX Solver
. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
4.5 Simulation Setup . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
4.6 Convergence Results
. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
4.7 Results . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

68
68
68
70
73
76
77
80
83
84
87

Chapter 5 Composite Manufacturing . . . . . . . . . . . . . . . . . . . . . . .
5.1
Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
5.2 Mandrel Modeling
. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
5.3 Mandrel Manufacturing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
3D Printed . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

96
96
97
98
5.3.1
98
5.3.2 Cast Wax . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 103
5.4 Fiber Winding Process . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 105
5.5 Resin Infusion/Curing Process . . . . . . . . . . . . . . . . . . . . . . . . . . 112
5.6
. . . . . . . . . . . . . . . . . . . . . . . . . . . 114
5.7 Results . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 119
Shaft Based Designs
. . . . . . . . . . . . . . . . . . . . . . . . . . . 119
Integrated Motor Based Designs . . . . . . . . . . . . . . . . . . . . . 128

Impeller Extraction Process

5.7.1
5.7.2

Integrated Motor - IM-v1.0
Integrated Motor - IM-v2.0

Chapter 6 Experimental Testing . . . . . . . . . . . . . . . . . . . . . . . . . . 135
6.1
Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 135
6.2 Electric Motor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 135
. . . . . . . . . . . . . . . . . . . . . . . 135
6.2.1
6.2.2
. . . . . . . . . . . . . . . . . . . . . . . 140
6.2.3 Variable Frequency Drive . . . . . . . . . . . . . . . . . . . . . . . . . 145
6.3 Assembly Balancing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 146
6.4 Testing Facility . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 146
6.5 Compressor Integration . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 149
6.6 Testing Results . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 150
IM-v1.0 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 150
IM-v2.0 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 152
6.6.2.1 Aerodynamic Results . . . . . . . . . . . . . . . . . . . . . . 152
6.6.2.2 Vibration Analysis . . . . . . . . . . . . . . . . . . . . . . . 153

6.6.1
6.6.2

Chapter 7

Spin Pit Testing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 158
Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 158
7.1
7.2 Spin Pit . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 158
7.3 Hub Adapter Redesign . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 160
Structural Analysis . . . . . . . . . . . . . . . . . . . . . . . . . . . . 162

7.3.1

vii

Introduction to FEA . . . . . . . . . . . . . . . . . . . . . . 162
. . . . . . . . . . . . . . . . . . . . . . . . . . . . . 169
7.4 Testing and Results . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 172

7.3.1.1
7.3.1.2 Results

Chapter 8 Conclusion and Future Work . . . . . . . . . . . . . . . . . . . . . 175
8.1 Conclusions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 175
8.2 Contributions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 176
8.3 Future Work . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 177

APPENDICES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 180
APPENDIX A Dynamic Balancing . . . . . . . . . . . . . . . . . . . . . . . . . . 181
APPENDIX B Composite Characterization . . . . . . . . . . . . . . . . . . . . . 185

BIBLIOGRAPHY . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 190

viii

LIST OF TABLES

Table 1.1: Axial compressor classifications . . . . . . . . . . . . . . . . . . . . . . . .

Table 1.2: Initial Winding Schemes . . . . . . . . . . . . . . . . . . . . . . . . . . . .

3

9

Table 1.3: Initial case study results . . . . . . . . . . . . . . . . . . . . . . . . . . . .

12

Table 1.4: Effect of shroud on structural strength of impeller

. . . . . . . . . . . . .

14

Table 1.5: NCG application design specifications . . . . . . . . . . . . . . . . . . . .

24

Table 1.6: Preliminary NCG application design result

. . . . . . . . . . . . . . . . .

25

Table 1.7: Small scale counter-rotating test loop results

. . . . . . . . . . . . . . . .

32

Table 3.1: Properties of 10 in schedule 40 PVC pipe

. . . . . . . . . . . . . . . . . .

63

Table 3.2: Geometric assumptions/specifications

. . . . . . . . . . . . . . . . . . . .

64

Table 3.3: Thermodynamic assumptions and specifications . . . . . . . . . . . . . . .

64

Table 3.4: Optimization Constrains Summary . . . . . . . . . . . . . . . . . . . . . .

64

Table 3.5: Geometric results summary . . . . . . . . . . . . . . . . . . . . . . . . . .

66

Table 3.6: Performance results - original . . . . . . . . . . . . . . . . . . . . . . . . .

66

Table 3.7: Performance results - adjusted . . . . . . . . . . . . . . . . . . . . . . . .

66

Table 4.1: Performance results - original compared to CFD . . . . . . . . . . . . . .

88

Table 4.2: Performance results - adjusted compared to CFD . . . . . . . . . . . . . .

88

Table 4.3: Comparison of analytical and CFD blade and flow angles

. . . . . . . . .

93

Table 5.1: Carbon fiber properties . . . . . . . . . . . . . . . . . . . . . . . . . . . . 106

Table 5.2: Geometry comparison . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 134

Table 6.1: Vacuum pump specifications

. . . . . . . . . . . . . . . . . . . . . . . . . 147

Table 6.2: Available sensors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 148

ix

Table 6.3: ISO 10816-3 vibration severity zones at approximately 15,700 rpm . . . . . 153

Table 7.1: 17–4PH H900 material properties . . . . . . . . . . . . . . . . . . . . . . . 162

Table 7.2: Grade 5 titanium alloy material properties . . . . . . . . . . . . . . . . . . 162

x

LIST OF FIGURES

Figure 1.1: Domains of turbomachinery . . . . . . . . . . . . . . . . . . . . . . . . .

Figure 1.2: Example stress vs strain curves for FRP and its constituents . . . . . . .

Figure 1.3: Typical setup for VARTM infusion process . . . . . . . . . . . . . . . . .

Figure 1.4:

Ideal vapor compression cycle for R718 and R134a

. . . . . . . . . . . .

2

6

7

8

Figure 1.5: First novel impeller winding . . . . . . . . . . . . . . . . . . . . . . . . .

10

Figure 1.6: Small scale 8B in cylindrical mandrel . . . . . . . . . . . . . . . . . . . .

10

Figure 1.7: 8B pattern rotors with various hub/tip ratios

. . . . . . . . . . . . . . .

11

Figure 1.8: 8C pattern rotors with various hub/tip ratios

. . . . . . . . . . . . . . .

11

Figure 1.9: Modified 8B rotor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

12

Figure 1.10: Maximum nodal displacements

. . . . . . . . . . . . . . . . . . . . . . .

13

Figure 1.11: Impellers with varying blade twist angles . . . . . . . . . . . . . . . . . .

14

Figure 1.12: Effect of blade twist angle . . . . . . . . . . . . . . . . . . . . . . . . . .

14

Figure 1.13: Basic geothermal power plant setup . . . . . . . . . . . . . . . . . . . . .

16

Figure 1.14: Plant net power output as a function of NCG fraction comparing com-
. . . . . . . . . . . . . . .

pressor, hybrid and steam jet ejector systems

20

Figure 1.15: Modular rotor counter-rotor pair

. . . . . . . . . . . . . . . . . . . . . .

22

Figure 1.16: Modular multistage compressor assembly . . . . . . . . . . . . . . . . . .

22

Figure 1.17: A prototype wound impeller with integrated magnets in shroud . . . . .

23

Figure 1.18: Cost breakdown for production of a tidal turbine blade . . . . . . . . . .

27

Figure 1.19: Marine life collision probability contours . . . . . . . . . . . . . . . . . .

28

Figure 1.20: Normalized extracted power variation with rotating speed for different
blade angles . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

29

xi

Figure 1.21: Wound composite turbine being tested in a tow tank . . . . . . . . . . .

30

Figure 1.22: Prototype small scale counter-rotating test loop . . . . . . . . . . . . . .

31

Figure 1.23: Small scale vacuum chamber for spin testing low mass composite impellers 32

Figure 2.1: Basic test loop schematic with secondary devices (vacuum pump and water
evaporation vessel) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

36

Figure 3.1: Analytical design constraint map . . . . . . . . . . . . . . . . . . . . . .

43

Figure 3.2: Control volume for a generalized turbomachine
Figure 3.3: h − s diagram for compression process

. . . . . . . . . . . . . .

46

. . . . . . . . . . . . . . . . . . .

48

Figure 3.4:

Isentropic vs polytropic compression process . . . . . . . . . . . . . . . .

48

Figure 3.5: Velocity Triangles . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

49

Figure 3.6:

Incidence and deviation definition with respect to the circumferential di-
rection for a rotating reference frame . . . . . . . . . . . . . . . . . . . .

52

Figure 3.7: Schematic of meridional view without optional OGV . . . . . . . . . . .

56

Figure 3.8:

IGV circular arc blade definition . . . . . . . . . . . . . . . . . . . . . .

58

Figure 3.9: R-CR circular arc blade definition . . . . . . . . . . . . . . . . . . . . . .

59

Figure 3.10: Resulting rotor design . . . . . . . . . . . . . . . . . . . . . . . . . . . .

65

Figure 4.1: Simulation domain . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

77

Figure 4.2: Turbulent boundary layer flow regimes . . . . . . . . . . . . . . . . . . .

78

Figure 4.3: Governing equation RMS residuals convergence on the mesh with the
highest density . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

85

Figure 4.4: Efficiency convergence on the mesh with the highest density . . . . . . .

85

Figure 4.5: Comparison of coarsest and finest meshes used in grid convergence study

86

Figure 4.6: Mesh independence study results

. . . . . . . . . . . . . . . . . . . . . .

86

Figure 4.7: CFD compressor mapping results - Πt-t and ηpoly,t-t vs ˙mcorr

. . . . . .

87

xii

Figure 4.8: Blade loading at midspan . . . . . . . . . . . . . . . . . . . . . . . . . .

89

Figure 4.9: Leading edge velocity distribution with small separation bubble . . . . .

90

Figure 4.10: Blade-to-blade velocity distributions

. . . . . . . . . . . . . . . . . . . .

91

Figure 4.11: Relative Mach number stream surface contour LE . . . . . . . . . . . . .

91

Figure 4.12: Relative Mach number stream surface contour TE . . . . . . . . . . . . .

92

Figure 4.13: Analytical vs CFD comparison of spanwise distribution of absolute and
relative flow angles including average incidence - LE . . . . . . . . . . . .

Figure 4.14: Analytical vs CFD comparison of spanwise distribution of absolute and
relative flow angles including average deviation - TE . . . . . . . . . . .

Figure 4.15: Circumferentially area averaged Cm contour in the meridional plane . . .

Figure 4.16: Circumferentially area averaged static pressure contour in the meridional
plane . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

92

93

94

94

Figure 5.1: Mandrel modeling progression . . . . . . . . . . . . . . . . . . . . . . . .

97

Figure 5.2: Printed mandrel with broken “finger” . . . . . . . . . . . . . . . . . . . .

99

Figure 5.3: Example of region of low print quality . . . . . . . . . . . . . . . . . . .

99

Figure 5.4: Printed infill at less than 100 % . . . . . . . . . . . . . . . . . . . . . . . 100

Figure 5.5: Polyjet mandrel . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 101

Figure 5.6: Soluble core dissolvable mandrel . . . . . . . . . . . . . . . . . . . . . . . 102

Figure 5.7: Mold of fluid path for wax casting . . . . . . . . . . . . . . . . . . . . . . 103

Figure 5.8: Wax mandrel aluminum base

. . . . . . . . . . . . . . . . . . . . . . . . 104

Figure 5.9: Wax mandrel hub region . . . . . . . . . . . . . . . . . . . . . . . . . . . 105

Figure 5.10: 1st cast wax mandrel . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 105

Figure 5.11: Basic winding steps . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 107

Figure 5.12: Hub winding step . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 108

xiii

Figure 5.13: Side view of winding in progress . . . . . . . . . . . . . . . . . . . . . . . 108

Figure 5.14: Mandrel with motor housing . . . . . . . . . . . . . . . . . . . . . . . . . 109

Figure 5.15: Pre-winding of hub region . . . . . . . . . . . . . . . . . . . . . . . . . . 110

Figure 5.16: Integrated motor first 6 winding steps of loop 1 . . . . . . . . . . . . . . 110

Figure 5.17: Integrated motor first 6 winding steps of loop 2 . . . . . . . . . . . . . . 111

Figure 5.18: Hub winding region of soluble core . . . . . . . . . . . . . . . . . . . . . 112

Figure 5.19: Axial winding progress . . . . . . . . . . . . . . . . . . . . . . . . . . . . 112

Figure 5.20: Flow paths filled with caulk . . . . . . . . . . . . . . . . . . . . . . . . . 114

Figure 5.21: Infusing cylinder

. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 114

Figure 5.22: Vacuum chamber used in infusion process

. . . . . . . . . . . . . . . . . 114

Figure 5.23: Post infusion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 115

Figure 5.24: Extraction process of Shaft-v1.2 . . . . . . . . . . . . . . . . . . . . . . . 117

Figure 5.25: Extracting IM-v2.1 - trailing edge partially finished . . . . . . . . . . . . 117

Figure 5.26: Close up of TE region where winding became difficult due to wax piece

voids . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 118

Figure 5.27: Removing excess resin at OD . . . . . . . . . . . . . . . . . . . . . . . . 118

Figure 5.28: Shaft-v1.0 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 119

Figure 5.29: Shaft-v1.0-Large (10 in OD) . . . . . . . . . . . . . . . . . . . . . . . . . 120

Figure 5.30: Shaft-v1.1 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 121

Figure 5.31: Shaft-v1.2 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 122

Figure 5.32: Shaft-v1.3 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 123

Figure 5.33: Shaft-v1.4 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 124

Figure 5.34: Shaft-v1.3 v1.4 comparison . . . . . . . . . . . . . . . . . . . . . . . . . 125

xiv

Figure 5.35: Shaft-v1.5 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 125

Figure 5.36: Shaft-v1.6 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 126

Figure 5.37: 3D scan - Shaft-v1.5 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 127

Figure 5.38: Impeller scan overlaid with CAD model for visual geometric comparison

128

Figure 5.39: IM-v1.0 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 129

Figure 5.40: IM-v2.0 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 130

Figure 5.41: Voids in hub flow surface (LE) - IM-v2.0 . . . . . . . . . . . . . . . . . . 131

Figure 5.42: Hub flow surface with no visible voids (TE) - IM-v2.0 . . . . . . . . . . . 131

Figure 5.43: IM-v2.1 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 132

Figure 5.44: A small void in hub flow surface (LE) - IM-v2.1 . . . . . . . . . . . . . . 132

Figure 5.45: Hub flow surface with no visible voids (TE) - IM-v2.1 . . . . . . . . . . . 133

Figure 6.1: TE view of integrated motor CAD design with magnet slots and stator . 136

Figure 6.2: Sectioned view of integrated motor CAD design with motor cap . . . . . 137

Figure 6.3: Kapton tape application to stator laminations . . . . . . . . . . . . . . . 137

Figure 6.4: Stator winding process for 4 pole, 3 phase wye configuration PMSM . . . 138

Figure 6.5:

Integrated motor mock up in 3D printed rotor . . . . . . . . . . . . . . . 139

Figure 6.6:

Installing integrated motor into IM-v1.0 . . . . . . . . . . . . . . . . . . 140

Figure 6.7:

IM-v1.0 assembly with integrated motor installed . . . . . . . . . . . . . 140

Figure 6.8: CAD model of PMSM for integration with IM-v2.0 . . . . . . . . . . . . 141

Figure 6.9: CAD model of IM-v2.0 with integrated motor . . . . . . . . . . . . . . . 142

Figure 6.10: IM-v2.0 mounted to the hub adapter - no motor . . . . . . . . . . . . . . 143

Figure 6.11: IM-v2.0 mounted to the hub adapter integrated motor assembly . . . . . 144

Figure 6.12: Aerodynamic testing facility . . . . . . . . . . . . . . . . . . . . . . . . . 147

xv

Figure 6.13: Possible compressor section probe arrangement

. . . . . . . . . . . . . . 148

Figure 6.14: Compressor assembly with hub annuli and mounting stanchions . . . . . 149

Figure 6.15: Sectioned view - CAD model of compressor section of test loop . . . . . 150

Figure 6.16: FFT of 0.5 s of accelerometer data during steady state operation at 119 Hz

excitation frequency . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 151

Figure 6.17: Total-to-total pressure difference measured across rotor with averaged (or-

ange) data overlaid onto original (blue) data . . . . . . . . . . . . . . . . 152

Figure 6.18: Inlet-accelerometer data during 0.5 s period of operation at design speed

154

Figure 6.19: Outlet-accelerometer data during 0.5 s period of operation at design speed 155

Figure 6.20: Inlet-FFT of accelerometer signal with first 6 higher harmonics labeled . 155

Figure 6.21: Inlet-FFT of accelerometer signal with NX/2 harmonics labeled . . . . . 156

Figure 6.22: Outlet-FFT of accelerometer signal with first 6 higher harmonics labeled 156

Figure 6.23: Outlet-FFT of accelerometer signal with NX/2 harmonics labeled . . . . 157

Figure 7.1: Spin testing facility . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 160

Figure 7.2: CAD model of modified hub adapter and shaft for spin testing . . . . . . 161

Figure 7.3: Modified hub adapter - reduced geometric domain for FEA . . . . . . . . 163

Figure 7.4: Grid convergence - Maximum shear stress, maximum principle stress and

von Mises stress vs element count . . . . . . . . . . . . . . . . . . . . . . 168

Figure 7.5: FEA of hub adapter assembly at 40,000 rpm . . . . . . . . . . . . . . . . 169

Figure 7.6: FEA of hub adapter assembly at 35,000 rpm . . . . . . . . . . . . . . . . 170

Figure 7.7: FEA of hub adapter assembly at 30,000 rpm . . . . . . . . . . . . . . . . 170

Figure 7.8:

IM-v2.1 assembly for spin testing . . . . . . . . . . . . . . . . . . . . . . 171

Figure 7.9: Mounted assembly hanging under bulkhead . . . . . . . . . . . . . . . . 172

Figure 7.10: SpinIV software and SA4 live feed . . . . . . . . . . . . . . . . . . . . . . 173

xvi

Figure 7.11: SpinIV monitor of air turbine speed vs time as the valve opening was

adjusted . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 173

Figure A.1: EasyBalance 2.2 Rotor Setup Screen . . . . . . . . . . . . . . . . . . . . 183

Figure A.2: EasyBalance 2.2 polar diagram display of rotor balanced to within G6.3

tolerance . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 183

Figure B.1: Setup for determining density . . . . . . . . . . . . . . . . . . . . . . . . 187

Figure B.2: Proposed infusion chamber design . . . . . . . . . . . . . . . . . . . . . . 189

xvii

KEY TO SYMBOLS

Speed of sound/sample apparent mass

Axial chord length

SST model coefficient

Control volume coefficient

Cross-sectional area

Coefficient matrix

Aspect ratio

Sample+sinker+wire apparent mass

Right hand side vector

Blade chord length

Constant pressure specific heat

Constant volume specific heat

Velocity in stationary frame/log-layer constant

Pressure coefficient

Turbulent eddy viscosity constant

Diffusion factor

de Haller number

Young’s modulus

Specific work

Permissible residual specific unbalance

Frequency

Applied force

Global applied load vector

Applied load vector

xviii

General

a

ac

a1

ap

A

[A]

AR

b
{b}

c

cp

cv

C

Cp

Cµ

DF

dH

E

˜e

eper

f

F
{F}
{F a}

i }
{F nr
F1/F2

G

G

h

h/t

H

i

k

[K]

[KT
i ]

L

m

M

˙m

mf

Restoring load vector

Blending functions

G number

Shear modulus

Enthalpy per unit mass

Hub-tip ratio

Blade height

Incidence angle

Turbulent kinetic energy

Global stiffness matrix

Jacobian tangent matrix

Characteristic length scale/undeformed length

Mass

Mach number

Mass flow rate

Mass fraction

MW

Molecular weight

n

N

Nb

p

pb

pc

pv

P

Pk

Number of elements/unit normal

Harmonic multiple number

Number of blades

Pressure

Burst pressure

Collapse pressure

Flow energy per unit mass

Power

Production of turbulent kinetic energy

Psep

Separator pressure

xix

P rt

q

˙Q

r
{r}

rb
{˜rϕ}
R
{R}

Re

s

s

S

SG

SE

SM

t

tb

T

u
{u}
{ui}
uτ

u+

U

U

Uper

Turbulent Prandtl number

Heat transfer per unit mass

Volume flow rate

Radius

Raw residual vector

Radius of blade curvature

Normalized residual vector

Specific gas constant

Out of balance vector

Reynolds number

Entropy per unit mass

Pitch

Invariant measure of the strain rate

Specific gravity

Energy source vector

Momentum source vector

Time

Blade thickness

Temperature

Internal energy per unit mass

Global solution vector

Displacement vector

Friction velocity

Dimensionless near wall velocity

Blade speed

Velocity vector

Permissible residual unbalance

xx

Ut

V
∀

w

W

x

X

y

y+

Z

Zb
∇·

O(cid:0)dx2(cid:1)

⊗

Greek

α

β

γ

Γ

Γef f

δ

∆




Velocity tangent to wall at distance ∆y

Velocity

Volume

Work done per unit mass/geometric parameter/sinker+wire ap-

parent mass

Velocity in relative frame

Geometric parameter

Harmonic multiplier

Geometric parameter

Dimensionless near wall distance

Axial position dimension

Number of blades

Divergence operator

Higher order of smallness terms

Tensor product operator

Mean component

Absolute velocity flow angle/SST correlation coefficient

Relative velocity flow angle/SST correlation coefficient

Heat capacity ratio/shear strain

Temperature ratio

Effective diffusion coefficient

Deviation angle/Kronecker delta function/boundary layer thick-

ness/change in length

Difference operator

Normal strain

xxi

ε

ζ

η

θ

κ

λ

µ

µt

ν

π

Π

ρ

σ

τ

τA

τω

φ

ϕ
{ϕ}

Φ

ψ

ω

Ω

Turbulence dissipation rate

Stagger angle

Efficiency

Camber angle

von K´arm´an constant

Thermal conductivity/eigenvalue

Dynamic viscosity

Turbulent eddy viscosity

Poisson’s ratio

Pressure ratio

Pressure ratio

Density

Blade solidity/SST correlation coefficient/normal stress

Torque/stress tensor/shear stress

Torque on axis A

Wall shear stress

Flow coefficient/angle of attack/turbulence model coefficient

Solution variable

Solution vector

Dependent variable vector

Stage loading coefficient

Angular speed(rad/s)/turbulent frequency

Rotational speed(rpm)

Subscripts

0

1

Total or stagnation properties

Inlet/1st principle axis/thermodynamic process point

xxii

2

3

4

ave

ax

c

corr

e

f

h

i

in

ij

ip

isen

j

k

LE

m

out

p

poly

ref

s

s-s

t

TE

Outlet/2nd principle axis/thermodynamic process point

3rd principle axis/thermodynamic process point

Thermodynamic process point

Average

Axial

Composite

Corrected property

equivalent stress

Fiber

Hub

Index/time step/equilibrium iteration number

Inlet property

Index notation

Integration point

Isentropic

Index

Index

Leading edge

Mean (radius)/meridional (velocity)/composite matrix

Outlet property

Polytropic

Polytropic

Reference property

Isentropic

Static-to-static

Tip

Trailing edge

xxiii

t-t

up

x

y

z

θ

Total-to-total

Upwind integration point

x-direction

y-direction

z-direction

Circumferential direction

Superscripts
∗
(cid:48)

(cid:48)

a

nr

o

T

Critical property

Blade angle

Fluctuating component

Applied force

Newton-Raphson restoring force

Previous time level

Transpose

xxiv

KEY TO ABBREVIATIONS

Acronyms

0D/1D/2D/3D

Zero-/one-/two-/three-dimensional

ABS

ACA

ASTM

ATM

CAD

CFD

CFRP

CR

DAQ

DIC

DNS

EGS

EIA

FDM

FEA

FFF

FFT

FoS

FRP

GE

GPP

HATT

HIPS

Acrylonitrile butadiene styrene

Advisory Committee for Aeronautics

American Society for Testing and Materials

Automatic Topology and Meshing

Computer aided design

Computational fluid dynamics

Carbon fiber reinforced polymer

Counter rotor

Data acquisition

Digital image correlation

Direct numerical simulation

Enhanced geothermal system

Energy information administration

Fused deposition modeling

Finite element analysis

Fused filament fabrication

Fast Fourier transform

Factor of safety

Fiber reinforced polymer

General Electric

Geothermal power plant

Horizontal axis tidal turbine

High impact polystyrene

xxv

HVAC

Heating, ventilation, and air conditioning

ID

IGV

ILU

ISO

LE

MCT

MSU

NCG

NDT

NI

NREL

OD

Inner diameter

Inlet guide vane

Incomplete lower upper factorization

International organization for standardization

Leading edge

Marine current turbine

Michigan State University

Noncondensable gases

Non-destructive testing

National Instruments

National Renewable Energy Laboratory

Outer diameter

OFGEM

Office of Gas and Electricity Markets

OGV

PLA

PMSM

R

R718

R134a

RANS

RMS

RTM

Outlet guide vane

Polyactic acid

Permanent magnet synchronous motor

Rotor

Water vapor

1,1,1,2-Tetrafluoroethane

Reynolds-averaged Navier-Stokes

Root mean square

Resin transfer molding

SIMPLE

Semi-implicit method for pressure-linked equations

SST

TE

Shear stress transport

Trailing edge

VARTM

Vacuum assisted resin transfer molding

xxvi

VATT

VFD

Vertical axis tidal turbine

Variable frequency drive

xxvii

Chapter 1

Introduction and Objectives

1.1 Dissertation Overview

The research presented in this dissertation is focused on furthering the development of the

continuous wound fiber manufacturing process for the production of composite axial tur-

bomachine rotors that is patented by Michigan State University (MSU) [1]. This chapter

serves to first provide a brief introduction to axial turbomachinery and the current state of

composite manufacturing of turbomachines in order to illustrate the original motivation for

this technology. The novel manufacturing process is presented in this context. A summary

of the prior work pertaining to the advancement of this technology is then provided. Lastly,

building off of this previous work, the research goals of this dissertation are introduced.

1.2 Background

1.2.1 Axial Turbomachinery

“A turbomachine is a device that exchanges energy with a fluid using a continuously flowing

fluid and rotating blades” [2]. Though it can be difficult to pinpoint when exactly the concept

of a turbomachine first came into being, it is thought to have originated thousands of years

ago to harness the power of moving water to improve irrigation systems and to grind grains.

“It was used by the Greeks to turn water wheels for grinding wheat into flour more than 2000

years ago” [3]. Another important development from this time period was the principle of

reaction. This was demonstrated by Heron of Alexandria through his Aeolipile device. “The

device can be described as a reaction turbine, since it makes use of the reaction force that

appears due to the momentum change in the jet of steam which is applied to the bent pipe”

1

[4]. Over time, this concept grew to become a large field with a wide range of applications.

There are two main branches of turbomachinery that emerged, “those that absorb power to

increase the fluid pressure or head (ducted fans, compressors and pumps); secondly, those

that produce power by expanding fluid to a lower pressure or head (hydraulic, steam and

gas turbines)” [5]. In addition to being classified as power producing or absorbing, there are

many types of geometric designs that can be categorized according to the fluid flow path.

This can be axial, radial or something in between. Typically, the specific application dictates

what type of turbomachine should be used. These domains can be seen in Figure 1.1 which

serves to illustrate how these different applications and device styles fit together.

The two main types of compressors and turbines that have become prevalent due to their

wide range of application are the axial and the centrifugal compressor and radial turbine.

Centrifugal compressors and radial turbines are typically utilized in situations where larger

pressure ratios and smaller flow rates are required. Axial compressors and turbines, on the

other hand, are utilized when dealing with smaller pressure ratios and larger flow rates.

Three main classifications of axial compressors that emerged based on application and level

(a) Physical Domain

(b) Fluid-Aero-Thermo Domain

Figure 1.1: Domains of turbomachinery

of performance are given in Table 1.1.

2

Table 1.1: Axial compressor classifications [6]

Application Flow

Inlet Relative Pressure Ratio Efficiency
Mach Number per Stage
per Stage

Industrial
Aerospace
Research

Subsonic
Transonic
Supersonic

0.40–0.80
0.70–1.10
1.05–2.50

1.05–1.20
1.15–1.60
1.80–2.20

88–92 %
80–85 %
75–85 %

Despite Heron demonstrating the concept of reaction thousands of years ago, most tur-

bomachines used were impulse turbines that harnessed flowing water to extract useful work.

The first attempt at building an axial-flow reaction turbine was performed by Osborne

Reynolds in 1875 but stopped when he concluded the tip leakages losses would prevent it

from being competitive with other technology of the time. Sir Charles Parsons indepen-

dently worked to develop his own multistage axial flow reaction turbine (initially for power

generation and ship propulsion), which soon became the standard for power plants. Parsons

also had patented an axial-flow compressor design as early as 1884. His original designs

were based on an axial turbine design operating in reverse which proved to perform poorly.

“The fundamental problem was the lack of aerodynamic knowledge in that turbine blading

was being used for axial compressor design. The challenge of diffusing flow and stability

raised its ugly head here and would remain a problem for axial flow compressors well into

the 1950s” [7]. It would require years of research into these flow phenomena to establish the

fundamental understanding required to overcome these design challenges.

It was A. A. Griffith in his 1926 paper “An Aerodynamic Theory of Turbine Design”

who first proposed that the blades needed special consideration in order to prevent them

from operating with stall (adverse pressure gradients leading to decelerated and stagnated

flow which largely contribute to the inefficient performance). He used the thin-airfoil theory

proposed by Ludwig Prandtl and his student, Max Munk, to design an airfoil shape that led

to a significant increase in performance. Griffith’s theoretical breakthrough convinced the

Advisory Committee for Aeronautics (ACA) to support a small-scale experiment on a single

3

stage axial compressor and single stage axial turbine. The results of this experiment sparked

a renewed interest in developing the potential of axial-flow compressors over the next several

decades. This led to massive gains in performance which pushed the structural integrity of

the materials being used to their limits. This has caused the need for development of new

materials and manufacturing processes capable of withstanding these extreme operating

conditions.

1.2.2 Composites

“The whole is greater than the sum of its parts”, a quote often attributed to Aristotle,

perfectly describes the underlying concept of composites. “A composite is when two or more

different materials are combined together to create a superior and unique material. The first

uses of composites date back to the 1500s B.C. when early Egyptians and Mesopotamian

settlers used a mixture of mud and straw to create strong and durable buildings” [8]. It was

not until the 1940s that the field of composites really grew into prominence. There were

several forces that caused this [9]:

“1. Military vehicles, such as airplanes, helicopters, and rockets, placed a premium on

high-strength, light-weight materials. While the metallic components that had been

used up to that point certainly did the job in terms of mechanical properties, the

heavy weight of such components was prohibitive. The higher the weight of the plane

or helicopter itself, the less cargo its engines could carry.

2. Polymer industries were quickly growing and tried to expand the market of plastics to a

variety of applications. The emergence of new, light-weight polymers from development

laboratories offered a possible solution for a variety of uses, provided something could

be done to increase the mechanical properties of plastics.

3. The extremely high theoretical strength of certain materials, such as glass fibers, was

being discovered. The question was how to use these potentially high-strength mate-

4

rials to solve the problems posed by the military’s demands.”

Fiber Reinforced Polymer (FRP) composites are typically composed of two main compo-

nents, the matrix and the reinforcement. The reinforcement material is usually very strong

and rigid but can be brittle. The matrix material is usually weaker but more flexible. By

combining these two components it is possible to enhance the strength and elasticity of the

matrix to achieve a product that is superior to the individual components as can be seen in

Figure 1.2. There are a wide range of choices for both the matrix or fiber components as

well as manufacturing methods that allow for the properties of FRPs to be customizable to

the specific application. The choice of fiber is typically made between glass fibers, aramid

fibers (most notably Kevlar) and carbon fibers. Fiber orientation within the matrix is one

of the most important aspects for determining how strong the resulting composite will be.

The most common choices of matrix are thermoset resins (those which require curing) such

as: epoxies, polyurethanes, phenolic and amino resins, bismaleimides and polyamides [10].

There are currently a number of different composite manufacturing processes that are

employed depending on the specific design or application of the composite part to be manu-

factured. FRPs are typically divided into two categories based on how the fiber is utilized.

There are short fiber reinforced composites or continuous fiber reinforced composites. The

use of short fibers is typically the most cost effective both in terms of fiber costs as well as

manufacturing costs but the strength of the resulting composite materials is low compared

to those made with continuous fiber. The continuous fiber reinforced composites, while more

costly, provide the highest strength to weight ratios.

The short fiber reinforced composites are usually manufactured by mixing the short fibers

into the uncured matrix and then this mixture is injected into a mold where it undergoes

a curing process similar to injection molding. The properties of the short fiber reinforced

composites can be tailored by adjusting the length of the short fibers as well as the ratio of

fiber to matrix.

Most other methods of composite manufacturing typically entail hand layup of woven

5

Figure 1.2: Example stress vs strain curves for FRP and its constituents [11]

fiber mats, either dry or prepreg, into a mold. For the dry fiber case this is followed by

a resin infusion process which is typically uses a low pressure vacuum to draw the resin

into the mold, called Vacuum Assisted Resin Transfer Molding (VARTM), as can be seen in

Figure 1.3. Another variation of this is to use positive pressure to force the resin into the

mold which is referred to as just Resin Transfer Molding (RTM). Dry fiber can be easier

to work with but needs special care to ensure that the resin is able to infuse properly. In

the prepreg case they are usually cured under low pressure and elevated temperature (such

as in an autoclave). One of the downsides to prepreg materials is because the resin is in B

stage (partial cure), they can need to be stored in a freezer to prevent the resin from curing

further [12].

With respect to turbomachinery, there are many reasons why being manufactured out of

composite materials could be beneficial. For example, in the aerospace industry, the thrust

6

Figure 1.3: Typical setup for VARTM infusion process [11]

to weight ratio of gas turbines is of paramount importance. Therefore, using composite

compressor impellers can provide the strength necessary while reducing the weight compared

to traditional metal parts. General Electric (GE) has been using composites for the front

stage fan of several jet engines such as the GE90 and GE9X. The improvements to the

thrust to weight ratio allowed for the GE90 to earn the distinction of the most powerful jet

engine while the GE9X is the largest jet engine in the world. This does not come without

cost though as each blade is independently manufactured by hand layup of approximately

1700 pieces of carbon fiber laminates requiring over 340 h per blade [13]. This is a common

problem faced by the composite turbomachine industry and therefore finding ways to reduce

the time and cost in the manufacturing process while maintaining or improving quality is a

major focus of the industry.

1.3 Previous Work on Wound Composite Rotors

1.3.1 Water as a Refrigerant Application

The original motivation for this technology arose while investigating the application of water

as a refrigerant. Typically, conventional refrigerants are harmful to the environment and have

contributed significantly to ozone depletion while the use of water vapor (R718) can avoid

these negative effects. However, in order for it to be comparable to classical refrigerants (like

7

R134a), R718 requires close to 200 times the volume flow and two times the pressure ratio

[14]. The concept of using water vapor (R718) as a refrigerant is not new, but in the past

has proven challenging to implement due to the location on the vapor dome the refrigeration

cycle must take place in comparison to other refrigerants as shown in Figure 1.4. Due to the

thermodynamic properties of water vapor, the speed of sound is also approximately twice as

large as that of R134a at the same temperature. These requirements mean that the allowable

Figure 1.4: Pressure-specific volume diagram of ideal vapor compression cycles for R718 and
R134a [15]

tip speeds can push traditional manufacturing methods to their limits and beyond.

In an attempt to help realize the potential of R718 as a refrigerant, M¨uller proposed a

novel manufacturing concept to utilize commercially available filament winding machines to

manufacture the rotor out of a continuous strand of fiber [16][17]. The goal being to greatly

reduce the weight of the rotor while simultaneously increasing the strength and scalability.

This would also significantly lower the cost of manufacturing compared to current methods

which require hand layup of woven fabric and resin transfer molding by allowing the process

to be automated. Through incorporation of integrated motors, counter-rotating multistage

8

compressors could more easily be created because this avoids the mechanical complications of

counter-rotating drive shafts and gearing while providing even more flexibility in operation

as each rotor’s speed can be controlled independently. This extra control may help with

off-design and startup operation and can inspire further research into this topic. This idea

of independently controlled stages is not limited to use with composite turbomachinery and

theoretically could be applied to any multistage turbomachine application.

Table 1.2: Initial Winding Schemes [16]

Some of the proposed continuous fiber patterns can be seen in Table 1.2. In the initial

attempts at manufacturing a continuous fiber wound rotor, a simple cylinder with slots

machined into it was used as the winding mandrel (see Figure 1.6) and the winding was

performed by hand. The use of this mandrel type allowed for a variety of unique impeller

designs to be investigated as shown below. Both wet winding and prepreg methods were

attempted by Eyler [18]. The wet winding process is shown in Figure 1.5a while a finished

wheel wound from prepreg fibers can be seen in Figure 1.5b.

While these first attempts at manufacturing helped to illustrate the continuous fiber

9

(a) Wet winding (8B Pattern)

(b) Prepreg winding (8C Pattern)

Figure 1.5: First novel impeller winding [18]

Figure 1.6: Small scale 8B in cylindrical mandrel

10

layup concept, the aerodynamics of these new “star” pattern designs had yet to be investi-

gated. The blades follow a forced vortex distribution where CU /r = constant which means

that there is more turning at the tip than at the hub. Li worked to analyze the aerody-

namic potential (for compressing water vapor as a refrigerant) of a number of these 8B and

8C designs with various hub/tip ratios h/t through Computational Fluid Dynamics (CFD)

analysis and found them to be somewhat limited in their performance potential as can be

seen in Table 1.3.

(a) 0.43

(b) 0.54

(c) 0.75

Figure 1.7: 8B pattern rotors with various hub/tip ratios [19]

(a) 0.43

(b) 0.54

(c) 0.75

Figure 1.8: 8C pattern rotors with various hub/tip ratios [19]

Li also performed a preliminary structural Finite Element Analysis (FEA) into the static,

dynamic and failure behavior of the 8B and 8C patterns using classical lamination theory

for unidirectional composites to estimate the material properties for Kevlar-49 and Carbon

Fiber Reinforced Polymer (CFRP). Considering only the centrifugal forces, he found that a

Kevlar impeller similar to that in Figure 1.5a could be able to withstand tip speeds of up to

11

Table 1.3: Initial case study results [19]

Case No. Rotor h/t Ratio Peak Π Peak η

1
2
3
4
5
6

8B
8B
8B
8C
8C
8C

0.43
0.54
0.75
0.43
0.54
0.75

1.107
1.121
1.162
1.115
1.124
1.163

64 %
65 %
68 %
66 %
65 %
73 %

˙m(cid:0)kg/s(cid:1)

0.95
0.75
0.45
0.80
0.70
0.30

830 m/s before reaching the tensile stress limits [20]. The results for the 8B and 8C impellers

with h/t ratio of 0.43 (as seen in Figures 1.7a and 1.8a) made of CFRP indicated that each

design could sufficiently handle the centrifugal forces during start-up, steady state, and shut

down periods of operation at up to tip speeds of approximately 225 m/s with a maximum
shroud deflection of approximately 2 mm (at 50 ◦C) [19].

Patil furthered this investigation by performing FEA on slightly modified version of the
8B impeller (with h/t ratio of 0.54 at 90 ◦C) for Kevlar-49, S-glass and carbon fiber. The

impeller model is shown in Figure 1.9. A comparison of the effect of the material on the

Figure 1.9: Modified 8B rotor [21]

12

radial deflection at tip speeds of approximately 250 m/s indicates that Kevlar and carbon

fiber significantly outperform S-glass as can be seen in Table 1.10. The Kevlar and S-glass

impellers also were predicted to fail near 300 m/s while the CFRP impeller was predicted

to initiate failure near 325 m/s [21]. While these predicted failure tip speeds are low, the

failure locations were around the flow hub area, and therefore it is predicted that the tip

speed limits could be much higher through improvement/strengthening of this region.

Figure 1.10: Maximum nodal displacements [21]

Patil also expanded upon the FEA analysis of the 8B design with no hub considerations

in order to investigate the effects of shroud thickness and blade twist angle on the strength of

the CFRP impeller. The three cases analyzed were the standard shroud thickness (2.8 mm),

a double-thickness shroud (5.6 mm) and the third case doubled the shroud thickness through

application of a woven ply rather than straight fibers. The meaning of the blade twist angle

is illustrated in Figure 1.11 while the results can be seen in Figure 1.12. It can be seen that

the maximum total deflection decreases with increasing blade twist angle while the maximum

out of plane deflection reaches an approximately constant value at angles of 20° and above.

The maximum radial deflection remains relatively constant (all values within approximately
±15% of the mean maximum radial deflection) with values ranging from 0.426–0.567 mm

13

Table 1.4: Effect of shroud on structural strength of impeller [21]

Case
Max radial deflection{mm}
Max out-of-plane deflection{mm}
Max total blade deflection{mm}
Failure Speed{rpm}

Base Double Woven Ply

0.56
0.60
1.87
12,000

0.38
0.54
1.71
13,000

0.47
0.54
1.69
13,000

and with the peak value occurring at 12° [21].

(a) 4°

(b) 16°

(c) 28°

Figure 1.11: Impellers with varying blade twist angles [21]

Figure 1.12: Effect of blade twist angle

14

1.3.2 Geothermal NCG Removal

Another potential industry that could benefit through application of this new technology

would be Geothermal Power Plants (GPP). The U.S. Energy Information Administration

has projected that the worldwide energy demand will increase by 48 % from 2012–2040 [22].

Meeting these demands will require utilization of many different energy sources. As climate

change becomes more severe, green and renewable energy sources are quickly becoming a

major focus in the energy sector. One issue presented by many sources of renewable energy

resources such as wind, solar and tidal is that they are limited by their cyclic nature. This

can be problematic for meeting the continuous base load requirements as well as spikes in

demand during peak usage periods. Therefore, having renewable energy sources that can

be continuously harnessed are crucial to meeting these needs. GPPs aim to extract power

from the near endless and constant supply of heat being generated in the core of the Earth.

According to the U.S. Energy Information Administration (EIA) renewable energy sources

accounted for roughly 13 % while geothermal alone only accounted for about 0.4 % of the

total U.S. energy production in 2015 [23]. It is projected that by 2050 geothermal could

provide approximately 3 % of global electricity demands and 5 % of the global demand for

heating and cooling [24].

Geothermal systems work by harnessing the thermal energy that is stored within the

Earth’s interior layers. An example layout of a GPP can be seen in Figure 1.13. Due to nat-

ural variations in the composition of the Earth’s crust and mantle, however, the temperature

as a function of depth can also vary significantly between different locations. This means

that some locations will be better suited for a geothermal power plant than others. There are

also several different types of GPPs that are utilized depending on these conditions. These

are given as:

1. dry steam plants

2. flash steam plants

15

3. binary cycle plants

4. direct use/geothermal heat pumps

Figure 1.13: Basic geothermal power plant setup [25]

Dry steam and flash steam systems are utilized at locations where the geothermal fluid
is at intermediate to high temperature, which is defined as ≥150 ◦C [24]. The dry steam

system is used when the geothermal fluid is already hot enough to consist predominately

of dry steam that can be directly used to drive the turbine and generator. When the

geothermal fluid is not hot enough to exist as dry steam but rather as a saturated mixture,

then a flash steam system is utilized to first flash evaporate the saturated mixture before

using the resulting steam to drive the turbine/generator system. Binary cycle systems are
used when the geothermal fluid is at low to intermediate temperatures of 70–170 ◦C [24]. Due

to the lower temperature of the geothermal fluid, the binary cycle systems make use of a

heat exchanger to extract energy from the geothermal fluid in order to vaporize a secondary

fluid with a significantly lower boiling point. This auxiliary vapor thermodynamic cycle

then drives the turbine and generator. Direct use and geothermal heat pump systems are

16

typically used in areas with low temperatures or in combination with other systems. They

can be used to provide building heating/cooling or hot water. These are larger scale versions

of geothermal systems someone would install at their home.

Locations where dry steam systems can be utilized are relatively uncommon and most

have already been developed, so there is minimal potential for expanding this type of GPP

usage. Because they can operate with lower geothermal fluid temperatures, the number of

viable locations for flash steam and binary GPPs to be installed is significantly higher. This is

especially true as the development of Enhanced Geothermal System (EGS) technology, where

geothermal reservoirs are engineered by creating the permeable pathways required within the

subsurface, has gained traction. According to a study done by Bertani, of these various plant

topologies, flash steam systems compose approximately 62 % of installed capacity and 63 %

of geothermal energy produced, yet only account for roughly 39 % of the total number of

installed units [26].

Since the geothermal fluid originates from deep within the Earth, it typically contains

a wide range of contaminants that include solids as well as gases. A majority of these

contaminants are separated and removed during the flash vaporization process. There are

certain Noncondensable Gases (NCGs) that cannot be easily converted to liquid phase for

removal with the flashing process. The noncondensable gases that are most commonly found

in geothermal resources include carbon dioxide (CO2), hydrogen sulfide (H2S), methane

(CH4), ammonia (NH3), hydrogen (H2), and nitrogen (N2). Of these, carbon dioxide usually

constitutes the majority of the NCG concentration with trace amounts of the remaining gases,

and together they are typically in concentrations of less than 5 % of the geothermal fluid

by weight [27]. Despite these low overall concentrations, the presence of these NCGs cause

GPP inefficiencies, increase operating costs and can contribute to green house gas emissions

if not handled properly. A National Renewable Energy Laboratory (NREL) report analyzing

alternative means of removing NCGs provides details of the practical problems associated

with elevated levels of noncondensable gases in geothermal power systems and is summarized

17

here [28]:

ˆ The NCGs reduce the heat transfer efficiency of the power plant condensers. This

causes an increase in condenser operating pressure, reducing the turbine output power.

To overcome this, larger condensers with greater heat transfer area and thus higher

costs would be needed.

ˆ The NCGs contribute a partial pressure that adds to the turbine back pressure, also

reducing power output.

ˆ Under-performing gas removal systems have the effect of an under-designed condenser,

increasing the turbine back pressure.

ˆ The NCGs contain lower recoverable specific energy than does steam. The gases dilute

the geothermal steam and reduce gross turbine output.

ˆ Most geothermal steam sources contain higher concentrations of NCGs (often by orders

of magnitude) than those seen in conventional fossil-fueled power plants. This causes

proportionally higher capital cost and operating components for gas removal to the

cost of electricity from geothermal plants.

ˆ Acidic gases like carbon dioxide and hydrogen sulfide are highly water-soluble and

contribute to corrosion problems in piping and equipment that contact steam and

condensate.

ˆ Conversely, as acidic gases escape from flashing geothermal brine, the pH of the brine

increases. This raises the risk of scale formation in brine piping and equipment, creating

a potentially expensive maintenance problem in the process systems that handle both

the steam and spent brine, including brine re-injection wells. Geothermal steam also

entrains brine mist that causes the buildup of scale in power turbines and in flow

systems.

18

In order to overcome the negative effects of the NCGs, they must be evacuated from

the condenser. The NCGs, along with a small amount of residual steam, is extracted and

compressed from vacuum to approximately ambient pressures. This has traditionally been

accomplished using a number of different vacuum systems. Due to the large size of these sys-

tems as well as their power consumption (whether parasitic steam or electricity usage), they

require significant design consideration when planning a new geothermal plant or retrofitting

an existing plant. The traditional NCG removal systems are:

1. Steam jet ejectors

2. Liquid ring vacuum pumps

3. Radial blowers/centrifugal compressors (turbocompressors)

4. Hybrid systems

Steam jet ejectors are the most commonly used method for NCG removal currently

because of their relatively low capital costs, and with no moving parts, the maintenance

costs are also lower. They utilize a parasitic stream of steam and capitalize on the Venturi

effect to create suction and “eject” the NCGs from the condenser. However, there are several

drawbacks to steam jet ejectors. The steam required for them to operate is steam that cannot

drive the turbine and thus reduces the power output. Steam jet ejectors are only about 15 %

efficient and are better suited to handle lower NCG concentrations (<3 %) [29]. The capacity

of a steam jet ejector is directly linked to its physical size, meaning there are practical limits

on how much compression and throughput a unit can provide which could lead to multiple

units being required.

Liquid ring vacuum pumps are typically used in conjunction with a first stage steam jet

ejector in hybrid system. They are a type of positive displacement pump that are better

suited for low flow applications which do not require a very large compression ratio. Due

to their more complicated design they have higher capital and maintenance costs but they

operate at efficiencies of around 50 %.

19

When there are concentrations of NCGs >3 % the parasitic steam requirements for a

steam jet ejector become uneconomical so radial blowers or centrifugal compressors are typ-

ically employed. One of the main benefits of a centrifugal compressor is that they are able

to operate at much higher efficiencies (approximately 70–80 %). Studies done by ¨Ozcan,

compared the performance of a two-stage steam jet ejector system, a hybrid system (one

steam jet ejector paired with one liquid ring vacuum pump) and a two-stage centrifugal

compressor system with respect to a single flash GPP application. The results indicate that

the NCG fraction is one of the most important factors influencing GPP performance and

that the compressor system “is the most efficient and robust system where the influence of

the NCG fraction is limited” [30]. This relationship of net plant power output vs. NCG con-

centration is illustrated in Figure 1.14 for a given flash separator pressure. The high capitol

Figure 1.14: Plant net power output as a function of NCG fraction comparing compressor,
hybrid and steam jet ejector systems [29]

cost and large size, paired with concerns regarding maintenance and reliability have been the

deterring factors, historically, for choosing this type of system. The traditionally metallic

compressor rotors are also at risk for corrosion damage. If these issues can be overcome, a

compressor system becomes the logical choice for a majority of NCG removal scenarios.

20

M¨uller and Patil proposed a modular multistage counter-rotating axial compressor with

integrated composite wound impellers as a new approach to alleviate some of these issues. A

summary of the proposed benefits of this new NCG removal system over traditional systems

are given [21]:

ˆ High volume flows

ˆ High efficiency

ˆ No parasitic steam use

ˆ Relatively low capital cost

ˆ Reduced cooling water needs

ˆ Corrosion resistant impellers

ˆ Least influenced by NCG fraction

ˆ Modularity allows for wide range of pressures

ˆ Integrated design and fewer parts lead to reduced maintenance

ˆ Robust system can be installed in-line with minimal footprint

ˆ Ideal for retrofitting existing plant or establishing new plants

The use of counter-rotation provides an opportunity to increase the power density and

thus reduce the overall size through elimination of the stationary guide vanes found between

each stage of unidirectional multistage compressor. An example of this type of modular

setup can be seen in Figures 1.15 and 1.16 where each rotor can be independently driven by

its own Variable Frequency Drive (VFD). Through integration of the magnets into the outer

shroud (as shown in Figure 1.17) and a clever approach to bearings and sealing it could be

possible to remove the need for a central shaft all together. This could allow for many of

the mechanical parts to be essentially removed from the corrosive fluid stream.

21

Figure 1.15: Modular rotor counter-rotor pair [21]

Figure 1.16: Modular multistage compressor assembly [31]

22

It should be clear now that the performance requirements of an NCG removal system are

highly dependent on the properties of the geothermal resource and specific plant design. In

prior work, this author designed the first three stages (an Inlet Guide Vane (IGV), Rotor

(R) and Counter-Rotor (CR)) of a multistage counter-rotating axial compressor with condi-

tions comparable to those found in flash steam plants located in California. Since the NCG

content is a majority carbon dioxide, for simplicity, the working fluid was approximated as a

mixture of H2O and CO2. Several other thermodynamic and geometric specifications were

used for the design process which can be found in Table 1.5. Other design goals, originating

from manufacturing considerations, were given as:

Figure 1.17: A prototype wound impeller with integrated magnets in shroud [31]

23

Table 1.5: NCG application design specifications [32]

Property

Value Units

Mixture MW 24.50
0.55
mf-H2O
0.45
mf-CO2
1.34
γ
2.00
˙m
3.75
Πt-t
8.27
p0,in
309.82
T0,in
8000.00
Ω
Inlet OD
0.26
0.09
Inlet ID
12
Zb

kg/kmol

−
−
−
kg/s
−
kPa
K
rpm
m
m
−

ˆ Constant tip radius

ˆ Constant axial length (chord not constant in radial direction)

ˆ Constant thickness blades

ˆ Circular arc blades

ˆ Maximize pressure ratio while maintaining efficiency

Through development of a simple analytical design approach (which can be found in [32]

and will be summarized later in Chapter 3) and CFD verification, the following results were

obtained.

24

Table 1.6: Preliminary NCG application design result [32]

Property Value

Rotor C-Rotor Total

˜e
P
τ
Πt-t
Γt-t
ηpoly,t-t
ηisen,t-t
dH
DF
φ
ψ

19.13
38.11
45.49
1.16
1.04
88.92
87.69
0.75
0.37
1.00
0.65

22.38
44.60
53.24
1.18
1.05
88.93
87.23
0.71
0.44
1.02
0.70

41.52
82.71
98.72
1.37
1.10
89.81
75.74
-
-
-
-

Units

kJ/kg
kW
N m
−
−
%
%
−
−
−
−

1.3.3 Marine/Hydro Turbine Application

This technology could also be applicable to the hydropower industry. Historically, hy-

dropower has been harnessed by damming a river, but unfortunately most locations where

this is feasible have already been dammed, so the focus is shifting to in-stream turbines and

the mostly untapped resource of tidal energy [33]. “The gravitational forces of the sun and

the moon create two ‘bulges’ in the earth’s oceans: one closest to the moon, and other on

the opposite side of the globe. These ‘bulges’ result in the two tides (high water to low water

sequence) a day - the dominant tidal pattern in most of the world’s oceans”[34]. The cyclical

nature of the tidal currents makes it a reliable and relatively consistent alternative energy

source compared to wind and solar. While the process of harnessing hydropower is similar

to that of wind energy, the power available is proportional to the product of fluid density

and the cube of velocity. With the density of water being approximately 1000 times larger

than air, this means that “A tidal current turbine rated at 2–3 m/s in seawater can result

in four times as much energy per year/m2 of rotor swept-area as similarly rated power wind

turbine”[35].

25

While tidal energy is an appealing source of renewable power it presents its own set

of challenges. The marine environment is one that is extremely harsh and unforgiving.

The salinity of seawater makes it especially corrosive to any metallic component, meaning

that sealing with a protective coatings are necessary to prevent corrosion from damaging

parts. Unlike in the R718 application where centrifugal forces dominate, the Marine Current

Turbine (MCT) application is dominated by forces in the axial direction, which for traditional

unshrouded designs means very large root bending moments acting on the blades. Due to

the high strength-to-weight ratio and corrosion resistance achievable with composites, they

have become increasingly attractive for MCT applications.

There are a wide variety of approaches that are being investigated for harnessing tidal

energy. The two main approaches are either vertical or horizontal axis turbines. Vertical

Axis Tidal Turbines (VATT) have the benefit of operating independent of flow direction as

the axis of rotation is perpendicular to the flow, but they can require relatively high flow

velocities to generate enough torque to begin rotating. The nature of their design also makes

them more suitable for shallow waters than Horizontal Axis Tidal Turbines (HATT) [36].

Due to the working principle of VATT, they are also more sensitive to cavitation [37] and

can be prone to pulsation [38]. “Its drawbacks in the structural robustness have always been

a burden in competing with the horizontal axis turbine” [39].

There are two main approaches for HATT design, one includes a shrouded tip and the

other is unshrouded. Those without shrouds more closely resemble a modern wind tur-

bine design. An example of this is the SeaGen S which was the first marine renewable

energy project to be accredited by the Office of Gas and Electricity Markets (OFGEM) as

a commercial power station [40]. The traditional composite manufacturing methods used in

constructing these by hand layup of wet plies into a mold can be very labor intensive, espe-

cially as the desired laminate thickness increases (requiring many more layers to achieve as

the thickness of each ply is limited to typically no more than 1 mm [41]). In the unshrouded

designs like the SeaGen S, the blade thickness near the hub needs to be very thick to with-

26

stand the huge bending moment that occurs, meaning labor costs can become a major factor

as seen in Figure 1.18. OpenHydro Ltd. was a company that was working on a shrouded

Figure 1.18: Cost breakdown for production of a tidal turbine blade [41]

turbine design [42]. This design incorporates an integrated permanent magnet generator on

the outer shroud allowing a central shaft to be omitted for an open center. Since torque is

a function of not only the force trying to rotate the object, but also the distance from the

axis of rotation, it means that by having the generator located at the outer radius, the forces

are significantly reduced for the same torque utilized for electricity generation. The turbine

is manufactured using traditional composite manufacturing techniques and has previously

experienced structural issues [42] [43]. Its massive size also significantly drives up the cost

of installation and maintenance.

A study done at the University of Strathclyde titled “Marine life interaction with tidal

turbines” took the bird collision model for wind turbines developed by Tucker and modified

it to predict the probability of marine life striking tidal turbines [44] [45]. In Figure 1.19

from this study it is clear that near the hub region has the highest probability of collisions

occurring, so having an open center can provide a path for marine life to pass through and

help to reduce the negative effects on the marine ecosystem.

The novel wound composite impeller technology serves to address many of the issues cur-

27

Figure 1.19: Marine life collision probability contours [44]

rently faced by the marine current turbine industry. The ability to automate many parts of

the manufacturing process can help to lower costs while the continuous fiber construction can

provide the strength-to-weight ratio desired in this application. The ability to manufacture

designs with open centers is also appealing for reducing the impact on marine life.

Wang performed a preliminary numerical investigations in to the application of this

technology for manufacturing marine current turbines. It was found that for an un-ducted

turbine, the average maximum power coefficient over a range of flow speeds (2.57–6.17 m/s)

was approximately 0.19, which is comparable to the value of 0.2 for the horizontal wind

turbine inspired design of IT Power Ltd.

[46] [38]. CFD analysis showed that by using a

diffuser after the turbine, it is possible to increase the maximum power extracted by up to a

factor of five when compared to a bare impeller [47]. By utilizing both a nozzle and a diffuser,

it was possible to increase the work extraction even further [48]. An initial investigation into

the effects of blade angle on the 8B impeller’s performance for a flow speed of 5 m/s shows

28

that as the blade angle increases, the power extracted also becomes larger, as does the range

of rpm that useful power is able to be extracted as seen in Figure 1.20. However, since

Figure 1.20: Normalized extracted power variation with rotating speed for different blade
angles [49]

for tidal energy it is desirable to harness energy in both tidal flow directions, special care

would need to be exercised with respect to symmetric harvesting of energy and the design

of the ducting if a nozzle/diffuser are to be used so that they can also perform in both flow

directions.

A study done by Belloni on the hydrodynamics of ducted and open-center tidal turbines

numerically investigated the performance of bidirectionally ducted turbines. For purely axial

flow conditions it was determined that although the power delivered per-rotor-area is similar

between the bare turbine, in ducted and open-center (also ducted) cases the rotor area per-

device (and thus power generating area) is reduced for the ducted and open-center turbines.

For non-axial (yawed) inflow, the ducting acts as a flow conditioner allowing the turbine to

experience nearly uniform flow over a wide range of yaw angles which allows the loading

experienced by the rotors to remain relatively axisymmetric (compared to the asymmetric

and cyclical blade loading experienced by the bare turbine). It was also shown that the power

extracted by the bare turbine decreases with increasing yaw angle while the ducted and open

center cases actually showed an increase in power production with increasing yaw angle (with

29

a maximum increase occurring at yaw angles of 20–30°) [50]. These results indicate some of

the potential advantages of utilizing a bidirectional ducted open center turbine, which can

provide more power in a wider range of flow conditions while reducing the structural loading

on the turbine and still provide a path for marine life to pass through.

Figure 1.21: Wound composite turbine being tested in a tow tank [21]

A very simple prototype turbine made by the MSU team from Kevlar-49/epoxy was

tested in a tow tank at the Marine Hydrodynamics Lab at the University of Michigan. The

turbine was mounted to a moving carriage which was used to control the relative velocity

between the turbine and the water while simultaneously measuring the torque generated

(through use of a load cell and disk brake attached to the shaft). One of the better tests

produced a power of approximately 2.5 kW at a flow speed of 3.6 m/s with the 1.5 m diameter

turbine wheel. The turbine design was not optimized and the manufacturing process was

in its infancy, which produced a relative low power coefficient indicating low efficiency [21].

Still, this proof-of-concept served to validate that the wound composite impeller for use in

marine current turbine applications could serve as a promising way to help lower the high

costs currently associated with the state of the art marine current turbine technology.

30

1.3.4 Other Previous Work

Patil also conducted experimental testing in a small test loop. The test was performed using

a pair of 8B star pattern impellers. They were placed in a rotor counter-rotor setup where

each stage was driven independently by an integrated motor at the shroud. The test setup

can be seen in Figure 1.22. Testing was performed at ambient conditions and a summary

Figure 1.22: Prototype small scale counter-rotating test loop [21]

of the results is given in Table 1.7. It can be seen that a maximum speed of 2750 rpm was

obtained for both rotor and counter-rotor which for such small impellers was only a tip speed

of approximately 20 m/s. A total-to-total pressure difference of 32.89 Pa was generated across

the impellers. While these numbers are not large, they served to demonstrate the feasibility

of the new impeller technology in conjunction with the integrated shroud motor.

Preliminary spin testing of a 10 in “star” pattern impeller was undertaken in a small scale

vacuum chamber which can be seen in Figure 1.23. The pressure inside was brought down

to approximately 5300 Pa before leakages prevented the pressure from being reduced any

further. The rotor was able to achieve a maximum speed of 14,000 rpm which corresponded

to a tip speed of 190 m/s before failure occurred [51].

31

Table 1.7: Small scale counter-rotating test loop results [21]

Property Value Units

ΩR
ΩCR
∆pt-t
∆ps-s

2750
-2750
32.38
29.89

rpm
rpm
Pa
Pa

Figure 1.23: Small scale vacuum chamber for spin testing low mass composite impellers [51]

1.4 Research Objectives

Given the wide range of potential applications where utilizing a composite wound impeller

could provide significant advantages, it is important to gain a better understanding of the

potential this technology possesses. The purpose of this research is to make significant

refinements to the novel wound composite impeller manufacturing process. The preliminary

research previously done at MSU served to highlight several areas that could be focused on in

32

order to increase the flexibility of design and aerodynamic potential of these new impellers.

They are summarized here:

1. Design

(a) The unconventional nature of the “star” pattern designs makes designing for

things like a particular mass flow rate or power difficult. With no existing body of

literature to reference, the process of meeting these types of design specifications

would be much more of an iterative method.

(b) The unique crisscrossing of the blades can prove challenging to accurately model

for FEA. Without experimental validation, it is difficult to place too much confi-

dence in those predictions.

2. Manufacturing

(a) The current state of the novel manufacturing method does not provide control

over blade thickness and results in very thin structures limiting the aerodynamic

design potential.

(b) It currently also results in low fiber volume fractions (mean experimental value

for Kevlar-49/epoxy was determined to be 0.29 while the theoretical maximum

for unidirectional composites is approximately 0.9 with typical values of around

0.6) which corresponds to lower strength as the fiber reinforcement provides the

majority of the strength.

(c) Use of wet-winding and prepreg methods result in a rough surface finish which

reduces efficiency.

With these preliminary investigations confirming suspicions that these exotic designs are

limited aerodynamically by the forced vortex swirl distribution and mandrel constraints, an

alternative approach was needed to improve the flexibility of design and quality of the next

33

generation of these wound composite impellers. Enabling a more conventional axial com-

pressor design appeared to be the most logical solution to achieving more predictable, higher

performance impellers but this would require further engineering of the novel manufacturing

process.

34

Chapter 2

Scope of Work

2.1 Introduction

While the original motivation for the novel manufacturing process arose from a single appli-

cation (R718 as a refrigerant), it quickly become clear that this could prove beneficial to a

wide range of industries and applications. If the limitations of this manufacturing process

could be shown to minimally affect the potential designs that are possible while ensuring

the structural properties are at least equal to, if not better than, current manufacturing

methods, then the potential for the technology becomes huge. This scalable, modular, and

flexible design and manufacturing method could prove to be revolutionary.

The first step is to provide an alternative approach to the simple cylindrical mandrel

that constrains the geometries to variations of the “star” patterns. Once the manufacturing

process has been adapted to produce more traditionally configured impellers (due to the

sizable research base already existing on this topic, this should greatly assist with achieving

better aerodynamic performance for whatever application is desired) it will need to be refined.

This will require working to identify how modifications to the several contributing sources

(rotor design, mandrel design, fiber layup, infusion and curing process) affect the quality of

the resulting impellers. For instance, how does having a large change in hub diameter from

inlet to outlet affect the winding process? Does this cause undesirable buildup of fiber in

some locations while leaving the composite with low fiber content elsewhere? The purpose

of this work is to adapt the manufacturing process to produce the desired geometries and

then begin to quantify and map the strengths and weaknesses for further improvements that

can be identified from these resulting impellers.

35

2.2 Design Framework

2.2.1 Experimental Testing Facilities

The light-weight, high-strength properties the wound composite impellers can provide led

M¨uller to initially consider its application in a low pressure high volume flow situations

like those found in the Heating, Ventilation and Air Conditioning (HVAC) industry. Other

potential applications where wound composite impellers could be utilized operate at pressures

below ambient in harsh environments include for non-condensable gas removal (necessary in

geothermal power plants) and for mechanical vapor compression (for desalination) just to

name a few. Performing initial testing under vacuum also reduces the power required to

drive the impeller. For these reasons, it was decided that when establishing an aerodynamic

testing facility (in the MSU Turbomachinery Lab) that being able to test under vacuum

conditions was desirable. With the impeller manufacturing and testing development still in

the early phases, flexibility of the testing facility was determined to be an important feature.

The size of the available space and the goal of using commercially available parts whenever

possible served to provide some initial constraints. The basic layout of a closed test loop can

be seen in the schematic below.

Figure 2.1: Basic test loop schematic with secondary devices (vacuum pump and water
evaporation vessel) [52]

36

It was decided that schedule 40 PVC pipe with an inner diameter (ID) of 10 in or 254 mm

would be used. Pohl determined that with a collapse pressure of pc = 2.79 bar that this pipe

is considered safe for use with an internal vacuum under outside atmospheric pressure [52].

The bursting pressure (pb) was found to be 9.65 bar and therefore this pipe should be safe for

a wide range of operating conditions with temperatures near ambient. A frame constructed

from 80/20 parts was to serve as the flexible backbone of the test setup to which the closed

loop was mounted.

2.2.2 Integrated Permanent Magnet Motor

While the idea of incorporating an integrated permanent magnet motor into an axial tur-

bomachine is not new (employed at the shroud by OpenHydro Ltd. [42] and inside the hub

by Tocardo International [53] to name two examples), it can be beneficial for a number of

reasons including:

1. Eliminates need for drive shafts and gear boxes allowing for more modular and compact

designs.

2. In multistage applications, the ability to control each stage independently should allow

more freedom in design optimization while increasing the operating range/improving

startup performance. A preliminary study of counter-rotating axial compressor stage

indicated that increasing the counter rotor speed to 1.5 times that of the rotor results

in shifting the stall point to a lower flow coefficient, widening the operating range [54],

however, this is a relatively unexplored topic that is currently being investigated by

the MSU Turbomachinery Laboratory [55].

3. Locating the integrated motor in the hub region can serve to keep the cost and weight

down while facilitating higher tip speeds, but extra care is needed to ensure the electric

motor components are properly isolated from the flow and are able to be kept from

overheating.

37

4. Locating the integrated motor at the shroud region can be beneficial in some instances

such as:

(a) When the working fluid is extremely corrosive and it is desirable to keep as many

of the electrical and mechanical parts outside of the fluid environment as possible.

(b) When the forces acting on the impeller blades are very large, by moving the force

transmitting location away from the axis of rotation the bending moment at the

blades of the impeller can be reduced (i.e. marine current turbine)

(c) When motor overheating is an issue, having the motor at the outer radius can

enable easier and more flexible cooling options.

2.3 Scope of Work

The work in this thesis can be divided into four main sections. The first (Chapters 3 and 4)

covers the development of an appropriate design scheme to allow for easy generation of axial

compressor impellers suited for manufacturing with the process presented in this research.

CFD analysis is then used to confirm the merits of the aerodynamic design before moving

forward. The second (Chapter 5) details the improvements to the novel manufacturing

process employed to realize the next generation of wound composite impellers. The third

(Chapter 6) covers the experimental testing of the prototype’s aerodynamics. The final

section (Chapter 7) deals with spin pit testing of an impeller.

2.3.1 Impeller Design

The preliminary designs were limited by several design choices aimed at increasing the ease of

manufacturing during these early stages of development while further design sophistication

can occur after the manufacturing process has become more refined. While it is perceived

that airfoil shaped blades could eventually be manufactured via the novel wound composite

technique, it can present its own set of challenges such as fiber bunching in some areas (for

38

example, at the leading and trailing edges) while having lower fiber content in other areas.

Therefore, by specifying that the blades be constant thickness with a circular arc camber

line and constant axial length, the design calculations are made to be simpler. This also

serves to help with beginning to mapping out how the various geometric features can affect

the manufacturing process and resulting impeller quality. The permanent magnet motor

is to be integrated into the hub region (in order to gain experience with this approach as

some experimenting with shroud motors was already undertaken previously as mentioned

in Chapter 1). The space limitations of placing a motor in the hub region mean that the

size and power limits will need to be considered when designing the prototype. With these

potentially problematic areas in mind, basic design goals can be laid out.

The prototype design must satisfy the following design requirements:

ˆ Constant tip radius (based on test loop diameter).

ˆ Constant axial length (chord varying in radial direction).

ˆ Constant thickness blades.

ˆ Circular arc camber lines.

ˆ Integrated motor power and size constraints.

ˆ Target pressure ratio of 1.1 with 75 % efficiency.

This was accomplished through use of a simple design tool created in Microsoft Excel.

The results from the Excel design were used as inputs for creating a 3D model for CFD

analysis and verification. The results of the CFD analysis served to indicate if any physical

phenomena like flow separation should be problematic for the design at hand. Once satisfac-

tory results were obtained in CFD, these would be the reference to which the experimental

results are compared. The 3D CFD model was then exported to a Computer Aided Drafting

(CAD) file for use in the creation of the mandrel.

39

2.3.2 Impeller Manufacturing

There were two major factors that contributed to the limitations of the initial iterations of the

novel wound composite impeller manufacturing. The use of the simple cylindrical mandrel,

while certainly inexpensive, limited the possible geometries that could be manufactured to

the exotic star pattern designs. The fiber layup method of wet winding appeared to have

added difficulty to the winding process. The prepreg method, while easier to layup, also left

something to be desired as can be seen in Figure 1.5b that the individual bundles of prepreg

fibers are still visible which indicates that they may not bond together as cohesively as other

methods.

In an effort to improve the surface finish of the impellers, it was proposed to first attempt

dry winding the fiber around the mandrel and then infusing it with the matrix through a

vacuum assisted resin transfer molding process. A new type of mandrel design was needed

in order to allow for construction of a more traditional axial compressor impeller with radial

blades extending from the hub to allow for better realizing the performance potential. For the

prototyping process many iterations of mandrels were 3D printed (in plastic) and sacrificed

during the process. The best mandrel design achieved with 3D printing was then made from

casting machinable wax to begin the investigation of other potential methods for mandrel

manufacturing. This method could be a more suitable choice for commercial production

because the wax can be melted down and reused to save on cost and time. Initially, impellers

were wound by hand, where care must be exercised in trying to maintain consistency of fiber

tension and placement so that a high quality composite composition can be achieved, though

it is thought that eventually the process can be automated with a 5 axis winding machine.

2.3.3 Experimental Testing

Once an impeller has been manufactured, the resulting geometry should be analyzed and

compared to the original design in an attempt to quantify how successfully the new man-

ufacturing method worked. Measurements of some of the basic geometric features like the

40

axial length, hub and shroud diameters, blade and shroud thicknesses can more easily be

physically measured using hand tools. However, physically measuring quantities such as the

blade angle can prove more challenging so it was proposed to investigate the use of 3D scan-

ning technology for overlaying the scanned model onto the original CAD model to better

determine how close the results are to the design.

Next, the impeller needs to be mounted to the integrated motor and balanced before

any testing can be carried out. The balancing is performed on a BalanceMaster HB-20-SS

dynamic (two-plane) horizontal balancing system and this process is detailed in Appendix A.

Preliminary aerodynamic testing of a next generation wound composite impeller with inte-

grated hub motor was undertaken.

2.3.4 Spin Pit Testing

Since the original motivation for this novel manufacturing process arose out of the need for

higher tip speeds (and thus higher strength impellers), it is necessary to begin trying to

test the tip speed limits that can be achieved with these impellers. This was accomplished

through spin pit testing. To ensure safety, this type of testing is performed in a specially

designed large reinforced vacuum chamber commonly referred to as a spin pit. The spin

pit located at Michigan State University currently can allow failure testing at speeds up to

40,000 rpm [56].

41

Chapter 3

Compressor Design

3.1 Axial Compressor Design

3.1.1 Introduction

The benefits of first performing a simple analytical design is that in this method only a

set of algebraic equations are needed (versus partial differential equations) which therefore

renders the calculations to be much less complex. When dealing with compressor design,

the freedoms that remain for the designer are finely connected to what is given or specified

in the design problem. This web of relationships can easily cause the design to become

over constrained if care is not exercised. In order to construct a simple design tool, a basic

strategy needs to be determined which will be used as guide to ensure that the design remains

within the scope of this work as well as making certain that the design does not become over

constrained. Figure 3.1 highlights the basic areas that these constraints originate from. It

also serves to illustrate for each area what is to be specified to meet the design requirements

while not violating any of the constraints that arise based on these design requirements.

42

1D/2D Design Constraints

Fluid

Inlet

Properties

Conditions

Geometric

Relationships

Aero./Thermo.
Relationships

Mixture

Total

Constant Tip

Conservation

Composition

Properties

Radius

Of Mass

Equation Of

Mach Number

State

Constant

Axial Chord

Conservation

Of Energy

Flow Angle

Circular Arc

Blade

Conservation
Of Momentum

Constant

Compressible

Blade Thick.

Flow

Free Vortex
Condition

Figure 3.1: Analytical design constraint map

3.1.2 Axial Compressor Basics

The goal of a compressor is to increase the pressure of the working fluid by performing work

on the fluid. The work is performed by the rotating blades of the compressor impellers. In

order to be able to describe and understand what is occurring during this process a num-

ber of equations are necessary. Several of these are derived using a control volume analysis

alongside basic aerodynamic and thermodynamic laws. The main thermodynamic laws used

are the ideal gas law, the conservation of mass (continuity), the first law of thermodynamics

(conservation of energy), Newton’s second law of motion (conservation of momentum) and

the second law of thermodynamics. The compressible flow relationships are derived using

these laws along with some algebraic manipulation. The velocity component vector relation-

ships are determined from geometric and reference frame considerations. A simple radial

equilibrium condition was used to expand the calculations from hub to tip.

The fluid is treated as an ideal gas meaning that it follows the ideal gas law, which is

43

given as:

p∀ = mRT , p = ρRT

(3.1)

The ideal gas law is an equation of state that provides an approximate relationship between

the pressure, temperature, mass, and volume of a gas at low densities (high specific volumes).

Gases deviate from the ideal gas behavior the most in the neighborhood of the critical point.

The pressures and temperatures typically dealt with in compressors are relatively mild when

compared with the values necessary for the ideal gas equation of state to start breaking

down. When using the ideal gas law it is also common to assume constant specific heats

(cp and cv) when the temperature variation is small.

The thermodynamic property specific enthalpy h and specific total enthalpy h0 are used

to describe the energy the moving fluid contains in terms of both the internal energy u and

the flow energy pv. In analyses where an ideal gas assumption is utilized the enthalpy can

be expressed purely as a function of temperature.

h = u + pv = cpT

h0 = cpT0

(3.2)

(3.3)

When dealing with the properties of moving fluids, there are two versions of each property:

the local or static and the total or stagnation. The local property is the value seen as if

the reference frame was traveling with the fluid while the total property (denoted by the

subscript 0) accounts for the local property and includes the effect of the fluid velocity as

well. Through use of the isentropic relationships it is possible to obtain the compressible
flow versions of these equations in terms of Mach number M and critical Mach number M∗.

The main thermodynamic equations of interest are summarized below and their derivations

a =(cid:112)γRT

44

can be found in [57].

(3.4)

(cid:114) 2γ

a∗ =

M =

V
a

T
T0

=

1
1 + γ−1
2 M 2

1

(cid:0)1 + γ−1
(cid:0)1 + γ−1

2 M 2(cid:1)γ/(γ−1)
2 M 2(cid:1)1/(γ−1)

1

p
p0

ρ
ρ0

=

=

RT0

V
a∗

γ + 1
M∗ =
= 1 − γ − 1
(cid:16)
1 − γ − 1
(cid:16)

1 − γ − 1

γ + 1

γ + 1

γ + 1

=

=

M∗2

M∗2(cid:17)γ/(γ−1)
M∗2(cid:17)1/(γ−1)

(3.5)

(3.6)

(3.7)

(3.8)

(3.9)

In order to operate efficiently, compressors are typically operated under steady state

conditions after the transient startup period is completed. Under steady state conditions,

the continuity equation states that the mass flow rate into the device equals the mass flow

rate out of the device:

˙m = ρinAinVin = ρoutAoutVout

(3.10)

Where for an axial turbomachine Vin,out is the velocity in the axial direction Vax and Ain,out

is the cross-sectional area defined as:

Aax = π(r2

t − r2

h) − (rt − rh)Nbtb

The mean radius is determined using the root mean square of the hub and tip radii:

(cid:115)

rm =

r2
t + r2
h

2

(3.11)

(3.12)

The conservation of energy equation is derived by applying the first law of thermody-

namics to a steady flow through a control volume. Utilizing the concept of enthalpy, the

conservation of energy equation can be expressed as:

˜e = ∆h0 = cp∆T0 = q − w

(3.13)

45

The heat transfer per unit mass (specific heat transfer) between the system and its environ-

ment is given as q. The work done per unit mass (specific work) is given as w and ∆h0 is the

change in specific total enthalpy. For many turbomachines it is safe to assume they operate

adiabatically, or with no heat transfer to or from the environment (q = 0). This means that

the change in energy of the fluid is only dependent on the work input by the shaft.

Applying the conservation of angular momentum to the control volume of a generalized

turbomachine (as shown in Figure 3.2), the shaft torque τ can be expressed as:

Figure 3.2: Control volume for a generalized turbomachine [5]

τ = ˙m(r2Cθ,2 − r1Cθ,1)

(3.14)

Considering the conservation of angular momentum equation, another expression for the

specific work is obtained by multiplying both sides of Equation 3.14 with the angular speed

ω (where rω = U ), and then dividing by ˙m. Equating these two expressions for the specific

work results in Euler’s turbomachinery equation.

˜e = ∆h0 = U2Cθ,2 − U1Cθ,1

(3.15)

46

Inspection of this equation indicates that the work input is directly related to the blade speed

U and the difference in the circumferential velocity of the fluid. Since in an axial machine,

the blade speed does not typically vary much (if at all) between the inlet and outlet, that

means the larger the change in Cθ, the more work is done on the fluid and hence larger

pressure gains can be realized. The power of the compressor can then be found by:

P = ˙m˜e

(3.16)

In order to be able to quantify how effective a compressor is at compressing the given

fluid an ideal reference process is needed. Application of the second law of thermodynamics

is used to provide this ideal scenario with which to compare against. For adiabatic flow the

ideal process would be isentropic with entropy s remaining unchanged. Figure 3.3 shows the

actual compression process requiring more energy input than the isentropic process. Through

this the definition of total-to-total isentropic efficiency can be provided as:

ηisen,t-t =

h02,s − h01
h02 − h01

(3.17)

The isentropic efficiency defined above does not provide the entire picture when compar-

ing turbomachines with different pressure ratios. To illustrate this, consider the compression

process as a large number of small stages instead of a single stage. If the efficiency of each

small stage is the same then the isentropic efficiency of the whole stage will not be equal

to this. The magnitude of the difference between these depends on the pressure ratio of

the whole stage. This is because at constant temperature the slope of the constant pressure
lines on the h − s diagram is the same but at a higher temperature (higher h) the slope of

the constant pressure line is larger (Figure 3.4a), where this divergence and thus pressure

ratio causes the difference between the small stage (polytropic) efficiency and the isentropic

efficiency as shown in Figure 3.4b. The formal definition of the polytropic efficiency is:

47

Figure 3.3: h − s diagram for compression process [5]

(a) h− s diagram with example small stage
process substeps

(b) Effect of pressure ratio on difference be-
tween isentropic and polytropic efficiency at
constant γ

Figure 3.4: Isentropic vs polytropic compression process [5]

ηpoly,t-t =

γ − 1
γ

ln(p02/p01)
ln(T02/T01)

(3.18)

48

Where the total pressure ratio and total temperature ratio are often given the symbols:

Πt-t =

Γt-t =

p02
p01

T02
T01

(3.19)

(3.20)

Considering again the concept of torque and angular momentum, it is desirable to be able

to determine the forces acting on the compressor blades. This is accomplished by analyzing

the flow in a two-dimensional cascade. The two-dimensional cascade analysis gives rise to

the concept of velocity triangles which describe the velocity vectors in both the relative and

absolute frames of reference. General velocity triangles for the inlet and outlet of a rotating

stage are shown in Figure 3.5. The absolute components of velocity at the outlet of one

α

β

Cax = Wax

C

Cθ

Wθ

(a) Inlet

W

U

α

β

C

Cax = Wax

W

Cθ

Wθ

U

(b) Outlet

Figure 3.5: Velocity Triangles

stage are equal to the absolute components of velocity at the inlet of the following stage.

It is often desirable to have an expression for calculating the total pressure ratio of the

compressor as a function of the specific work, efficiency and inlet total temperature. Through

use of Equation 3.13 and the isentropic relations the specific work can be related to the total

pressure ratio:

Πt-t =

(cid:16) ˜eηisen,t-t

cpT01

(cid:17)γ/(γ−1)

+ 1

In order to allow for the comparison of machines of differing geometries and/or operating

49

(3.21)

conditions it is customary to use non-dimensional numbers/dynamic scaling. These can take

different forms depending on the type of turbomachine being considered (i.e. a pump, where

the fluid is incompressible, vs. a compressor). Two parameters that deal with aerodynamic

scaling are the flow coefficient and the stage loading coefficient. As shown in Equation 3.15,

the amount of work input to the flow depends partially on the blade speed U as well as the

amount of flow through the stage. Therefore, the flow coefficient ψ is defined as the ratio of

axial velocity to the blade speed. Selecting a value for this quantity serves to fix the relative

flow angles.

φ =

Cax
U

(3.22)

The other parameter of importance is referred to as the stage loading and is based on

the specific work or stagnation enthalpy change (i.e. Equation 3.15) where it has been

nondimensionalized by the blade speed squared.

ψ =

∆h0
U 2

(3.23)

According to Cumpsty, values of the flow coefficient typically fall in the range of 0.3–0.9

[58]. It has been shown that for a fixed stage loading, increasing the flow coefficient reduces

the required flow turning, suggesting higher values of flow coefficient could be advantageous

as long as care is taken to avoid choking/shock formation [5]. The stage loading provides a

measure of the flow turning and thus work input of the stage. Typically the stage loading is

limited to values of 0.4, but some highly loaded stages can have values up to 0.75 [5].

Two other important values for analyzing a cascade are incidence and deviation, com-

paring the blade angles to the actual flow angles. The first is called incidence i, which is

the difference in the flow direction into the blade and the tangent on the camber line at the

leading edge (where

(cid:48)

denotes the blade angle from the flow angle):

i = αin − α

(cid:48)
in

50

(3.24)

When dealing with moving blades, this definition is altered to refer to the relative flow angle

β and the blade angle:

i = βin − β

(cid:48)
in

(3.25)

For the angle convention used in this research, it is referred to as negative incidence if the

incidence causes the flow to experience more turning than due to the blade angle alone and it

is called positive incidence if it causes less turning. The flow coefficient essentially determines

the incidence for a particular stage (i.e. blade angles and ω/U ) because an increase from

the design flow coefficient provides a larger Cax which makes the flow angle larger (with

respect to the circumferential direction) while a smaller flow coefficient results in a smaller

Cax and a smaller flow angle compared to the fixed blade angle [58]. Lieblein determined

that there is an optimal range of incidence where the losses remain small, however outside

of this range the losses start increasing rather quickly due to stall [59]. As Mach number

increases compressors typically become more sensitive to the effects of incidence and the

acceptable range of incidence becomes narrower.

Applying this same concept to the outlet of a blade row results in what is referred to as

the deviation, δ given as:

δ = αout − α

(cid:48)
out

(3.26)

Again, when dealing with rotating blades, the definition is changed to reflect the relative

flow angle compared to the blade angle:

δ = βout − β

(cid:48)
out

(3.27)

The angle convention used here dictates that the deviation will be negative (i.e. the flow

angle is smaller than the blade angle because of being referenced to the circumferential

direction instead of the axial direction as shown in Figure 3.6). This value quantifies how

well the flow follows the turning of the blade. The flow will never follow the blade perfectly

so there is always some amount of deviation present. Too-high a deviation angle means the

51

blade is trying to turn the fluid farther than is possible and as a result the flow will separate.

Figure 3.6: Incidence and deviation definition with respect to the circumferential direction
for a rotating reference frame

There are also several important geometric parameters used in the design process. One

is the blade solidity σ which is the ratio of the blade chord c to the pitch s. Acceptable

values of blade solidity can range from 0.8–2 [5] [58].

σ =

c
s

(3.28)

The next parameter relates the difference between the tip radius and the hub radius, the

blade height H, to the blade chord length, essentially determining the compressor length

[60]:

AR =

rt − rh

c

=

H
c

(3.29)

This influences the blade losses and stability margin and commonly has values of 1–2 [5] but

52

middle and rear stage compressor rotors have had success in the range of 0.67–1 [61] [62].

Since the blade chord length is a result of the design process in this work an axial aspect

ratio is used to specify the axial chord length:

ARax =

H
ac

(3.30)

The other is the thickness-chord ratio which is the ratio of blade thickness to blade chord

length tb/c. Values of the thickness-chord ratio typically range from 5–10 %, where thin-

ner blades with thickness-chord ratio of 5 % have shown better performance at high Mach

numbers compared to thicker blades [5].

The pressure coefficient is another useful parameter for analyzing the performance of

airfoils and compressor blades. It is the difference of the static pressure nondimensionalized

by division with that is given as simply a pressure difference divided by a reference pressure

difference (usually the dynamic pressure, or total minus static). For compressors, this is

referenced to the inlet location:

Cp =

p − p1
p01 − p1

(3.31)

Inspection shows that Cp is negative when p < p1 and positive when p > p1. It is possible

to plot the pressure coefficient along the pressure and suction sides of the blade to gain a

better understanding of the flow occurring over the blade.

As mentioned above, the difference in Cθ is mostly responsible for the work input and

thus the increase in pressure. Logically then, to maximize pressure ratio the blade should

provide maximum turning to give the largest difference in Cθ. However in reality, if the

blades attempt to turn the flow too far, flow separation can begin to occur on the suction

side of the blade. Flow separation can lead to compressor stall and large reductions in

efficiency. Therefore, it is important to ensure that this does not occur. Over time several

parameters were devised that can be easily determined and provide an indication on when

separation and then potentially stall would occur. The simplest has been the de Haller

53

number developed by de Haller in 1953 [63]:

dH =

W2
W1

(3.32)

It gives an indication of the amount of deceleration the flow experiences through the

compressor stage. The simplistic nature of this factor makes it quick and easy to use as an

indicator of the blade performance, however it is strictly based on the relative flow velocities

at the inlet and outlet and does not include any geometric parameters. Also in 1953, Lieblein

devised the Diffusion Factor DF [59], which is similar to the de Haller number but with some

slight modification. He included de Haller’s deceleration ratio and added a term to emphasize

the flow turning in respect to the blade solidity as a geometric parameter for the capability

of the blades to guide the flow:

(cid:21)

(cid:20)

1 − W2
W1

|∆Wθ|
2σW1

+

DF =

(3.33)

The absolute value of the relative turning in the added term is used to account for counter-

rotation as suggested by Petralanda in [64]. Typically dH must be larger than 0.7–0.72. For

the DF , values of 0.6 or greater will usually indicate compressor stall and flow separation

which leads to a sharp drop in efficiency, while most compressors have values in the range of

0.35–0.55.

In order to expand the design consideration in the radial direction as the 3rd dimension,

a simple theory of radial equilibrium is introduced with the goal to balance the pressure

forces with the centrifugal forces acting on the rotating fluid so that the radial component

of velocity remains approximately zero. Through application of an elemental force balance

and omission of the higher order of smallness terms (i.e. O(cid:0)dx2(cid:1)) the radial force balance

equation becomes:

1
ρ

dp
dr

=

C2
θ
r

54

(3.34)

By assuming that Cax, h0 and s are constant radially and using thermodynamic relations

this can be simplified to:

d
dr

(cid:0)rCθ

(cid:1) = 0 , rCθ = constant

This is often referred to as the free vortex design and a more detailed derivation can be found

in [5] or [65]. With this, it is then possible to find values of Cθ at all locations of known

radius. This assumption of no radial velocity component also makes the meridional velocity

Cm equal to Cax. While this simple relationship might not be ideal for more sophisticated

designs it will allow for ease of initial design for CFD.

It is common to define the corrected mass flow with respect to a reference temperature

and pressure (typically atmospheric pressure, 101.325 kPa, and temperature at sea-level,

(3.35)

(3.36)

288.15 K) as:

˙mcorr = ˙m

(cid:113)

T0/T0,ref
p0/p0,ref

This dimensional quantity is the mass flow that would pass through the device if it was

operating with atmospheric inlet conditions.

3.1.3 Axial Compressor Design Methodology

Due to the rather unusual set of design specifications, an unconventional design methodol-

ogy needed to be developed which could accommodate for this while being able to quickly

generate potential designs. First, a “0D” design was generated to determine all the prop-

erties at the inlet and outlet of each guide vane or rotor. These results were then used as

the inputs for expansion in the axial direction (“1D”) and with the free vortex design the

necessary information for constructing a 3D model of the design. CFD analysis of the 3D

model could then be used to ensure that the design operates without any major issues like

flow separation and to provide more realistic performance expectations than those provided

by the simplified design calculations.

The 0D design process considered the inlet and outlet station of the IGV, R and CR (and

55

Outlet Guide Vane OGV if desired) shown schematically in Figure 3.7. It determines the

geometry and properties at the hub, tip, and mean locations for each of these stations. With

Figure 3.7: Schematic of meridional view without optional OGV

the R (or R-CR pair) as the focus, the design began at the rotor inlet station and then, if

they are desired, the guide vanes were matched to this. Since the size of the prototype rotors

was mainly limited by the test stand and integrated motor dimensions, through specification

of the desired air gap and shroud thickness the inlet annulus became fully defined. By then

defining the desired number of blades and blade thickness the inlet cross-sectional area was

known.

With initial testing to be performed at ambient temperatures, by choosing the working

fluid (air or water vapor) and the inlet total pressure the inlet total properties were fully

defined. Because the design constraints required constant thickness circular arc blades,

traditional loss correlations were foregone in favor of a generalized loss through specification

of an approximate polytropic efficiency for each stage similar to the method used by Wilcox

in where “no direct consideration was given to the presence or absence of shock waves,

boundary layer, flow separation, wakes, and so forth, but an allowance for the effect of such

disturbances was made by assuming values of adiabatic efficiency” [66]. Through the use of

the velocity triangle relationships introduced in Figure 3.5, all of the fluid properties and

flow conditions can be determined at the inlet station. The common simplifying design

assumption of constant axial velocity throughout the compressor was utilized to bridge from
the inlet to the outlet [67] [68]. Then, by varying α, M∗

W at the inlet and β and
h/t at the outlet, the specific work was maximized while ensuring the many constraints were

C and M∗

56

satisfied.

With the blade angles at the inlet and outlet of the impeller known and by specifying a

constant axial length (per impeller) as well as circular arc blades, all blade angles are known

throughout the impeller. With this information it is possible to step through the rotor

satisfying continuity by solving for the requisite hub profile. The first step that needed to be

taken was to discretize the domain axially. The axial length or axial chord (ac) is controlled

through specification of the axial aspect ratio (ARax) that is used in the denominator while
rt − rh is the numerator. Each domain was discretized equally into n = 10 elements or
n + 1 = 11 nodes. Therefore, the axial position of the inlet of each element is obtained with

ac divided by n:

Zi = Zi−1 +

ac
n

(3.37)

With the axial discretization completed it was then possible to use this in conjunction

with trigonometric relationships to determine the blade angle at each axial node. For a

circular arc blade, the inlet and outlet angle are related by the camber angle:

θ = (βin − βout)

(3.38)

Figure 3.8 illustrates that for a circular arc blade under the special condition that βin is

equal to 90 degrees (with respect to the circumferential versus 0 degrees if it was referenced

to the axial direction), which applies to the IGV. Notice that for βin = 90° the stagger angle

is ζ = θ/2. The radius of the circular blade curvature can be determined with:

rb =

ac
sin θ

c = 2rbsin

2

With rb and θ it is possible to determine the chord length with:

(cid:16) θ

(cid:17)

57

(3.39)

(3.40)

Figure 3.8: IGV circular arc blade definition

Now solving Equation 3.39 for θ and substituting Equation 3.37 for ac.

θi = sin−1(cid:16) Zi

(cid:17)

rb

Finally utilizing this in Equation 3.38 and rearranging results in:

βi = βin − θi

(3.41)

(3.42)

While above can be used for the guide vanes with one of the blade angles being 90 degrees,

the procedure becomes more complex if neither of the specified blade angles is 90 degrees

(comparing Figure 3.8 and Figure 3.9).

Four geometric relationships with four unknown parameters are formulated in Equa-

tions 3.43-3.46 to fully determine the blade angle everywhere and are given in (note there

58

Figure 3.9: R-CR circular arc blade definition

are 4 equations and 4 unknowns).

cos(90 − βout) =

x + w

rb

sin(βin) =

x
rb
x

tan(βin) =

ac + y

tan(90 − βout) =

y

x + w

(3.43)

(3.44)

(3.45)

(3.46)

This system of equations was solved using linear algebra in the form of coefficient matrix

59

inversion of Equation 3.47.



(cid:124)

cos(90 − βout)

sin(βin)

0

0

−1

0

0

(cid:123)(cid:122)
tan(90 − βout)

A

−1
−1
−1

tan(90 − βout)

tan(βin)

=



(cid:125)

rb


(cid:124) (cid:123)(cid:122) (cid:125)x

w

x

y

·

0

0

−ac tan(βin)

0

(cid:123)(cid:122)

b

0

0

−1



(cid:124)

(3.47)



(cid:125)

The important result is shown in Equation 3.48, which can then be used in conjunction with

Equation 3.42 to obtain the blade angle at every location.

(cid:32)

(cid:33)

θi = sin−1

ac − Zi + y

rb

(3.48)

The domain was discretized radially into n = 15 equally sized elements with n + 1 = 16

nodes. This allows for the determined properties to be appropriately mass or area averaged

for comparison with the CFD results.

In this approach, the mass flow rate is a calculated value that is controlled mainly by

the specification of the inlet pressure, temperature, Mach number and geometry. For the

remaining axial nodes, with the mass flow rate and axial velocity constant throughout, it
was possible to provide an initial guess of h/ti ≥ h/ti−1 at each node and refine it in brief
iteration to satisfy continuity throughout. This also determined the hub contour. In fact,

for a design, if the inlet temperature and Mach number are held constant then adjusting

the inlet pressure adjusts the density,mass flow rate and power while the physical geometry

remains the same which is desirable for being able to test a prototype over a range of

60

conditions without having to manufacture a new impeller. The geometry of the design only

changes (to satisfy continuity) if the design inlet temperature or Mach number is changed.

Due to the linear relationship between pressure and density for constant temperature (from

Equation 3.1), the density,mass flow rate and power vary linearly with changing design inlet

pressure.

With the assumption of fluid total properties and axial flow area being constant through

a guide vane it was then possible to match the IGV to the rotor inlet conditions by adjusting
the IGV M∗

C value at each axial location to ensure that continuity was satisfied. In the case
where an OGV was desired this same process could be applied to match the rotor outlet

conditions.

If a R-CR setup was desired, there are two small differences between the treatment of

a R and a CR that serves to account for the counter-rotation. First, the absolute velocity

triangle at the R outlet is the same as at the inlet of the CR. With Cax, CU and α known at
the CR inlet, iteration with M∗

W fully determines the velocity triangle, and therefore U (and
ω) and βin can be calculated while ensuring continuity is satisfied. The combined specific

work and thus pressure ratio could then be maximized.

With the R and CR fully defined and solved for, now the IGV can be matched to the R

inlet. Since the IGV is stationary, the blade angle affects the absolute velocity components

(represented by C) rather than the relative velocity components (represented by W ) like in

the rotating impellers. This can be expressed as αR,in = βIGV,out. Now the IGV blade angles

are fully specified throughout, and the remaining properties can be solved for by iterating
on M∗

C to satisfy continuity at each node beginning at the outlet of the IGV and progressing

towards the inlet.

Up until this point, the design process has assumed inviscid flow to simplify the calcula-

tion process. This is typically a good assumption everywhere except near the hub, shroud,

and blade surfaces where in reality viscous effects lead to boundary layer formation. This

boundary layer effectively reduces the cross-sectional area of the compressor, which directly

61

affects how much mass flow can fit through (Equation 3.10), creating a discrepancy between

the mass flow of the analytical design compared to the 3D CFD simulation which accounts

for viscous effects. A simple method was used to determine approximately how much, on av-

erage, the boundary layer reduces the cross-sectional area. Through application of boundary

layer theory detailed in Schlichting [69], the boundary layer displacement on each surface was

estimated at each axial node. A new value of the cross-sectional area was then determined

that included the boundary layer thicknesses. With this area, an approximate, adjusted ˙m

was calculated. This was used when specifying boundary conditions for the CFD analysis.

If the original fully inviscid value was used for specifying the boundary conditions, the CFD

simulation would typically become choked and diverge.

To get a better estimate of the true performance of the design before the CFD analysis

has been performed, a reduction factor was employed. The ratio of the boundary layer

adjusted mass flow rate to the inviscid mass flow rate was utilized as the reduction factor.

By applying this to quantities of interest in the appropriate manner, it was possible to obtain

values more representative of what could be expected from CFD. The quantities ˜e,

˙Q, φ and

ψ can be adjusted through multiplication with the reduction factor while the pressure and

temperature ratios need to be raised to the power of the reduction factor. The power can be

calculated from the adjusted values of ˙m and ˜e while the torque can then be found from the

adjusted value of power and the angular speed. It should be noted that while this simplified

method worked satisfactorily to provide a quick way to estimate results to be expected from

CFD analysis, though a more thorough investigation of various geometries, fluids and flow

conditions is suggested to determine the limitations on this approach.

62

3.2 Rotor Design

3.2.1 Design Criteria

In order to begin the design process, the necessary geometric and thermodynamic assump-

tions and specifications need to be quantified. The size of the testing facilities was the major

geometric constraint by limiting the size of the impeller. Table 3.1 provides a summary of

the properties of the 10 in schedule 40 PVC pipe that is being utilized for the closed test

loop.

Table 3.1: Properties of 10 in schedule 40 PVC pipe

Property

Value

ID
OD
Tolerance
Minimum ID
Wall thickness

10.00/254.00
10.75/273.05
±0.015/±0.381
9.985/253.619
0.365/9.270

Units

in/mm
in/mm
in/mm
in/mm
in/mm

To ensure that the rotor does not come into contact with the wall of the pipe, the

minimum pipe ID was used as the limiting diameter. An air gap was specified to provide

enough space to handle the radial expansion of the rotor and the thickness of the outer rotor

shroud was specified. These two values were then subtracted from half of the minimum pipe

ID to obtain the flow tip radius. The flow hub radius at the inlet was specified based on

limitations imposed by the integrated hub motor requirement. The first iteration of the hub

motor was designed for 1 kW with plans to increase this to 15–20 kW in the future. This

power limitation means that the compressor will need to have a relative small mass flow rate,

since P = ˙m˜e and ˜e is essentially constant for a given pressure ratio. Keeping h/t high helps

to keep the mass flow down in conjunction with a low inlet pressure. The blade thickness,

number of blades, and axial aspect ratio were specified. The number of blades and aspect

ratio could then be adjusted as needed to assist with obtaining satisfactory DF values.

Preliminary aerodynamic testing was performed using air as the working fluid under the

63

Table 3.2: Geometric assumptions/specifications

Property

Value Units

Air gap
Shroud thick.
rt,in
rh,in
tb
h/t
Nb
ARax

mm
2.60
3.00
mm
121.21 mm
100.00 mm
mm
2.00
−
0.825
−
19
−
0.85

conditions specified in Table 3.3. The limits utilized for the floating and constraint variables

Table 3.3: Thermodynamic assumptions and specifications

Property

Value Units
−

Working Fluid Air
Rair
T0,in
p0,in
ηpoly,t-t

286.99
298.15
5
80

J/(kg K)

K
kPa
%

can be found in Table 3.4. This was implemented into an Excel program that was used to

Table 3.4: Optimization Constrains Summary

Property Limits

C

α, β
M∗
M∗
W
h/t
dH
DF

15 ≤ α, β≤165
0 ≤ M∗
C≤0.7
0 ≤ M∗
W≤0.8
h/tout ≥ h/tin
dHmean≥0.72
DFmean≤0.45

Units

°

−
−
−
−
−

solve for the rotor design based on maximizing the specific work within the framework of

these specifications and constraints.

64

3.2.2 Results

Figure 3.10: Resulting rotor design

The 3D geometry that resulted from using the analytical results as inputs into the model-

ing software can be seen in Figure 3.10. The large h/t ratio makes this rotor characteristic of

a later stage of a multistage compressor as can be seen by the significant amount of available

space in the hub region which will provide the room necessary for integrating the motor.

A summary of the main geometric results can be found in Table 3.5. The average values

refer to the radially area averaged values. The solidity of 1.047 falls within the acceptable

range of 0.8–2, discussed above. The thickness-chord ratio is approximately 5 % and is in

the common range of 5–10 %. The aspect ratio is lower than the traditional range of 1–2

but, again, values of less than 1 for later stage rotor designs has been shown to work well.

Table 3.6 provides a summary of the original, inviscid results. Due to the inviscid nature

of the calculations the performance is over-predicted. Since the motor was limited to 1 kW

initially and if the design was actually going to need 1.7 kW then it would not be able

to operate correctly. Through application of the reduction factor method, more realistic

values were obtained. These adjusted values are given in Table 3.7 alongside their original

counterpart. The power is now much closer to the target of 1 kW. Here the value of 0.54

65

Table 3.5: Geometric results summary

Table 3.6: Performance results - original

Property
Machine

rt,in;rt,out
rh,in
rh,out
h/tin
h/tout
ac
Aax,in
Aax,out
cave
save
σave
tb/c
AR

Flow Angles

αave,in
αave,out
βave,in
βave,out

Value Units

Property Value Units

Ω
˙m
˙Q
˜e
P
τ
Πt-t
Γt-t
ηpoly,t-t
ηisen,t-t
dH
DF
φ
ψ

15,705
0.10
1.85
16.96
1.70
1.03
1.165
1.057
80.00
79.60
0.72
0.45
0.72
0.50

rpm
kg/s
m3/s
kJ/kg
kW
N m
−
−
−
−
−
−
−
−

121.21 mm
100.00 mm
102.04 mm
−
0.825
−
0.842
mm
24.95
m2
0.0139
m2
0.0127
40.35
mm
mm
38.57
−
1.047
−
4.96
−
0.56

102.86
65.20
31.88
47.21

°

°

°

°

Table 3.7: Performance results - adjusted

Property

Value
Original Adjusted

Units

˙m
Reduction factor
˙Q
˜e
P
τ
Πt-t
Γt-t
φ
ψ

0.1001
0.78

1.85
16.96
1.70
1.03
1.165
1.057
0.72
0.49

0.0779
−
1.44
13.20
1.32
0.63
1.123
1.044
0.56
0.38

kg/s
−
m3/s
kJ/kg
kW
N m
−
−
−
−

66

for the adjusted flow coefficient is within the typically accepted range of 0.3–0.9. Values of

flow coefficient around 0.5 are typically chosen for initial designs because the compressor

is more stable to flow disturbances and off design operation than those with higher flow

coefficients. The value of 0.37 for the loading coefficient is in the middle of the typical range

0.3–0.5. Equations 3.15 and 3.23 indicate that a larger stage loading requires more turning

of the flow. This needs to be balanced with the blade solidity to keep the diffusion factor

under its limit resulting in the suggested range of solidity values mentioned previously with

stage loading typically limited to 0.4 [5]. Due to the simple reduction factor technique, these

values will typically be slightly less in CFD than the adjusted values, so it is acceptable if

they are slightly larger than the target value.

67

Chapter 4

Numerical Investigation

4.1 Introduction to CFD

In order to gain a more comprehensive understanding of how the compressor design will

perform in reality, a CFD analysis that accounts for the viscous effects was the next step

of the design process. Caution must be exercised however, because CFD results are highly

sensitive to the mesh quality as well as the simulation setup (physics, boundary conditions,

etc.). The CFD analysis was performed with the ANSYS 15.0 line of software through

the use of ANSYS Workbench, BladeGen, TurboGrid, and CFX modules, Workbench is

ANSYS’ project management interface that unites the standalone modules. First, BladeGen

was used to convert the before produced geometry data into an actual 3D model. Second,

the simulation domain (fluid flow path) was meshed using TurboGrid. Next, the simulation

was setup and executed in the CFX solver that employs a finite volume approach. Lastly,

the results were post processed in CFD-Post.

4.1.1 Governing Equations

Chapter 3 introduced the governing equations used in the context of performing the simplified

analytical design work. More general forms of them are used in the CFD analysis. Again,

these conservation laws are:

1. The conservation of mass.

2. The conservation of momentum.

3. The conservation of energy.

68

This set of equations are what constitute the Navier-Stokes equations that govern the motion

of fluids. Each of these equations can be expressed in integral or derivative form. Applying

them to a fixed finite control volume directly leads to the integral form whereas applying

them to a fixed infinitesimally small volume directly leads to partial differential form. The

latter are used in CFX where the control volume is split into many small volumes, which is

called meshing and is discussed in more detail further below.

The general form of the conservation of mass equation for a compressible fluid is then

given in tensor and regular notation:

+ ∇ · (ρU ) = 0

∂ρ
∂t

∂ρ
∂t

+

∂ρui
∂xi

= 0

(4.1)

(4.2)

where the indices i, j, k represent spatial directions, ρ is the fluid density, t is time, U is

the three dimensional velocity vector, xi is the i-th spatial dimension and Ui is the i-th

component of velocity. The general form of the conservation of momentum equation is:

∂(ρU )

∂t

+ ∇ · (ρU ⊗ U ) = −∇p + ∇ · τ + SM

∂(ρui)

∂t

+

∂(ρuiuj)

∂xj

= − ∂p
∂xi

+

∂τij
∂xj

+ SM,i

(4.3)

(4.4)

where SM is the momentum source vector, SM,i is the momentum source in the i-th direction

and τ is the stress tensor that relates to the strain rate by:

τ = µ

∇ · U + (∇ · U )T − 2
 ∂ui

− 2
3

∂uj
∂xi

3

τij = µ

+

∂xj



(4.5)

(4.6)

δ∇ · U



δij

∂uk
∂xk

where µ is the dynamic viscosity of the fluid and δ is the Kronecker delta function that is

69

equal to 1 for i = j and 0 otherwise.

The general form of the conservation of energy equation is given as:

∂(ρh0)

∂t

− ∂p
∂t

+ ∇ · (ρU h0) = ∇ · (λ∇T ) + ∇ · (U · τ ) + U · SM + SE

(4.7)

(cid:18)

(cid:19)

∂(ρh0)

∂t

− ∂p
∂t

+

∂(ρujh0)

∂xj

=

∂
∂xj

λ

∂T
∂xj

+

∂(uiτij)

∂xj

+ uiSM,i + SE,i

(4.8)

where h0 is the total specific enthalpy, λ is the thermal conductivity, T is the local tempera-
ture, ∇· (U · τ ) is the viscous work, U · SM is the work due to external momentum sources
and SE is the energy source vector.

4.1.2 Reynolds-Averaged Navier-Stokes Equations

In general, the Navier-Stokes equations introduced in the previous section describe both

laminar and turbulent flows, where it depends on if the viscous or inertial forces are more

significant in the flow to categorize it as one of the other. In laminar flow the fluid viscosity

provides a significant enough damping effect to keep the flow orderly, while in turbulent

flow the inertial forces dominate and lead to chaotic fluctuations of the flow properties. To

put the contributions of these two forces into relation for a given flow, the dimensionless

Reynolds number is defined as the ratio of the inertial forces to the viscous forces:

Re =

ρU L

µ

(4.9)

where ρ is the density, U is the fluid velocity, L is the characteristic length scale of the flow

being considered (e.g. blade chord length) and µ is the dynamic viscosity of the fluid. It is

apparent that laminar flow occurs at low Reynolds numbers and turbulent flow occurs at high

Reynolds numbers. The transition from laminar to turbulent flow does not occur or start

to occur at a specific Reynolds number as it is highly complex process that is not yet fully

understood, but the extensive research on the topic has allowed scientists to specify ranges

70

of Reynolds numbers for specific flow situations that represent the laminar, transitional and

turbulent flow regimes. In turbomachinery applications, the flows are typically turbulent.

Turbulence is characterized by fluctuations, in both space and time, of the flow field and

occurs over a very large range of length and time scales which leads to the complex nature

of turbulent flows and processes.

The foundations of turbulent analysis are based on two major concepts. First, is the idea

of the energy cascade postulated by Richardson in 1922 [70]. It states that turbulence can be

considered to be composed of different sized eddies (turbulent motions). The large eddies are

inherently unstable and break up, transferring their energy to smaller eddies which in turn

break up and transfer their energy to even smaller eddies. This process continues until the

Reynolds number of these eddies becomes small enough that the motion can become stable

due to the viscosity effectively dissipating the kinetic energy. The Kolmogorov hypotheses

proposed by Kolmogorov in 1941 [71] postulate that eventually the small scale motions

become locally isotropic, which determines the lower limit of the scales of length, time and

velocity (called the Kolmogorov scales) needed to fully account for turbulent motions. Due to

the nonlinearity of the Navier-Stokes equations they typically cannot be solved analytically

except in the most simple of cases and therefore are solved numerically. The Direct Numerical

Simulation (DNS) approach requires that all of the applicable length and time scales be

resolved, which can become extremely computationally expensive for even relatively simple

turbulence problems. To overcome this limitation, a turbulence modeling based approach is

used in conjunction with the Reynolds-Averaged Navier-Stokes (RANS) equations to obtain

approximate solutions.

ANSYS CFX utilizes a RANS solver, which employs Reynolds decomposition to break in-

stantaneous velocities into their time-averaged and fluctuating quantities. For incompressible

flow and subsonic flow time averaging is used and for compressible flow with large changes in

density correlations involving density fluctuations may be introduced which can be difficult

to handle. Using a mass weighted time averaging technique called Favre time averaging can

71

help to avoid some of these complications in transonic and supersonic flow.

If Φ(x, t) is any dependent variable, then decomposing this into the mean Φ(x) and

fluctuating Φ(x, t)(cid:48) parts gives Φ = Φ + Φ(cid:48). Where the mean is defined as:

(cid:90) t+∆t

Φ =

1
∆t

t

Φdt

(4.10)

and the time average of the fluctuating component is given as:

(cid:90) t+∆t

(cid:90) t+∆t

Φ(cid:48) =

1
∆t

t

Φ(cid:48)dt =

1
∆t

t

(Φ − Φ)dt = 0

(4.11)

To obtain the RANS equations each dependent variable is decomposed and substituted into

the original conservation equation. Then the time averaging technique is applied to this to

obtain the mean equation. Subtracting the mean equation from the original equation with

the decomposed parts gives the fluctuating equation. The fluctuating component of density

is typically very small and so is omitted giving ρ = ρ. The resulting RANS equations are

then:

∂ρ
∂t

+

∂(ρui)

∂xi

∂(ρui)

∂t

+

∂(ρuiuj)

∂xj

= − ∂p
∂xi

+

(cid:18)

(cid:17)

+ SM,i

= 0

(cid:16)

∂
∂xj

(cid:19)

τij − ρu(cid:48)
iu(cid:48)
(cid:104)
(cid:16)

j

(4.12)

(4.13)

+ SE,i

(4.14)

(cid:17)(cid:105)

∂(ρh0)

∂t

− ∂p
∂t

+

∂(ρujh0)

∂xj

=

∂
∂xj

λ

∂T
∂xj

− ρu(cid:48)
jh

+

∂
∂xj

τij − ρu(cid:48)
iu(cid:48)

j

ui

The mean total enthalpy is then represented as:

h0 = h +

1
2

uiui + k

(4.15)

2 u(cid:48)
iu(cid:48)
where k is the turbulent kinetic energy, k = 1
i. Comparing Equation 4.4 and 4.13
it can be seen that they are almost identical in form except Equation 4.13 contains the
extra term ρu(cid:48)
j. These are referred to as the Reynolds stresses. The extra term ρu(cid:48)
iu(cid:48)
jh

72

comparing Equation 4.8 and 4.14 is referred to as the Reynolds flux. The introduction

of these extra terms represents the turbulence closure problem in which turbulence models

must be provided for the computation of the Reynolds stresses and fluxes allowing the RANS

equations to be solved.

4.1.3 Turbulence Modeling

Over the years many different models have been proposed. The complexity of these models

can vary significantly both in implementation and computational costs. Ideally, for a tur-

bulence model to be considered useful it should be applicable to a wide range of problems,

accurate, and computationally economical. There are generally two types of turbulence

models: eddy viscosity and Reynolds stress models. The Reynolds stress turbulence models

are based on using seven additional transport equations to solve for the components of the

Reynolds stress tensor and dissipation rate. Solving all of these additional equations can

quickly become very computationally expensive and therefore this is typically not utilized.

All eddy viscosity turbulence models are based on the Boussinesq hypothesis which assumes

that the momentum transfer caused by turbulent eddies can be modeled with an eddy vis-

cosity analogous to how the momentum transfer caused by molecular motion in a gas can

be described by a molecular viscosity. This allows the Reynolds stresses to be assumed

proportional to the mean velocity gradients in the form:

 ∂ui

∂xj

+

∂uj
∂xi

−ρu(cid:48)
iu(cid:48)
j = µt

 − 2

3

δij

ρk + µt



∂uk
∂xk

(4.16)

where µt is the turbulent eddy viscosity that needs to be modeled to fully close the system

of equations. Similarly, the Reynolds flux can then be modeled as:

−ρu(cid:48)

jh =

µt
P rt

∂h
∂xj

(4.17)

with P rt being the turbulent Prandtl number.

73

There are three main categories of eddy viscosity turbulence models:

1. Zero-equation (algebraic)

2. One-equation

3. Two-equation

The zero-equation model utilized in CFX uses a simple algebraic equation to find a constant

value of µt for the entire flow domain, ignoring the transport of turbulence. This method

does not work well when the turbulent length scale varies, such as flows with separation or

circulation, and is therefore limited to very simply flows.

The one-equation models are based on adding an additional transport equation of either

the turbulent kinetic energy or the turbulent eddy viscosity and prescribing a turbulent

length scale. This accounts for the transport of some turbulence but is still limited to

specific types of flow such as attached wall-bounded flow.

For two-equation models an additional transport equation is solved in addition to one

for the turbulent kinetic energy. The purpose of this additional equation is to model the

turbulent length scale, which the previous models have not done. Because of this, two-

equation models account for the convection and diffusion of turbulent energy. This makes

them advantageous for problems involving more complex geometries and flow fields. The
k − ε and the k − ω models are the two main approaches.
In the k − ε model, the ad-

ditional transport equation models the turbulence eddy dissipation ε and then relates the

turbulent eddy viscosity to the turbulent kinetic energy and eddy dissipation by a constant

of proportionality:

µt = ρCµ

k2
ε

(4.18)

The k − ε model is relatively simple to implement and can provide reasonable predictions
for many flows such as free shear type flows. For the k − ω model the additional transport

equation is used to model the turbulent specific dissipation (or turbulent frequency) ω and

74

then assumes a relationship to the turbulent eddy viscosity by:

µt = ρ

k
ω

(4.19)

The k − ω model is similar to the k − ε model in implementation but it performs well for

wall bounded flows.

The Shear Stress Transport (SST) turbulence model developed by Menter was chosen as

the closure method for this study because it gives highly accurate predictions of onset and

amount of flow separation under adverse pressure gradients [72]. This turbulence model is
a hybrid of the k − ε model that is better suited for free turbulence and the k − ω model

that is better suited for wall bounded turbulence. Menter proposed a method of combining

the two models by first transforming the ε equation to be in terms of ω, giving one equation

for ω in the free stream and one for the near wall. Then, using a blending function, the

desired contribution can be obtained depending on the location. This was initially called
the baseline k − ω model, but this overpredicted the turbulent eddy viscosity, preventing

proper prediction of onset and amount of flow separation. The final transport equations for

the SST model are then:

∂(ρk)

∂t

+

(cid:19) ∂k

(cid:21)

∂xj

µt
σk3

+ Pk − β(cid:48)ρkω

(4.20)

(ρujk) =

µ +

∂
∂xj

(cid:20)(cid:18)

(cid:20)(cid:18)
(cid:21)
(cid:19) ∂ω

∂
∂xj

∂xj

∂(ρω)

∂t

+

∂
∂xj

(ρujω) =

∂
∂xj

µ +

µt
σω3

+ (1 − F1)2ρ

1

σω2ω

∂k
∂xj

∂ω
∂xj

+ α3

ω
k

Pk − β3ρω2
(4.21)

where Pk represents the production of turbulent kinetic energy. The α’s, β’s and σ’s are

correlation coefficients with the subscript referring to which model it originated from (1 =
original k − ω, 2 = transformed k − ε, 3 = baseline k − ω and SST, while (cid:48) applies to all).

The coefficients of the baseline and SST models are assumed to be a linear combination of

75

the corresponding coefficients of the underlying models:

φ3 = F1φ1 + (1 − F1)φ2

(4.22)

Menter then proposed a modification to limit the value of the turbulent eddy viscosity near

wall boundaries by changing the definition of the turbulent eddy viscosity to:

µt =

ρa1k

max(a1ω; SF2)

(4.23)

where S is an invariant measure of the strain rate and F1 and F2 are blending functions, the

values of which can be found in the ANSYS CFX-Solver Theory Guide [73].

4.2 Compressor Modeling

The BladeGen module was used in conjunction with the 2D results to generate the 3D mod-

els for the simulation. The basic meridional profile of the blade is input into BladeGen along

with the desired mode (specifying angle and thickness in this case). Then, the desired blade

thickness and number of blades are specified before entering the main BladeGen interface.

From here it is possible to input or adjust the geometric properties to construct the desired

blade geometry. The full rotor was shown previously in Figure 3.10. To reduce the com-

putational load, the periodic nature of turbomachines is utilized to reduce the simulation

domain to the fluid path around a single blade as can be seen in Figure 4.1.

76

Figure 4.1: Simulation domain

4.3 Meshing of Fluid Domain

To solve the governing equations using the finite volume method, the physical domain needed

to be meshed or split into many small finite volumes, the mesh elements. TurboGrid uses

hexahedral shaped elements, where each element has 8 nodes or vertices which are shared

with the adjacent elements. All fluid properties and solution variables are then stored at

each node. The actual control volume that the conservation equations are applied to is then

constructed around each mesh node by joining the centers of the edges and element centers

surrounding the node. In 3D, this results in each hexahedral element being divided into four

smaller element sectors and referred to as an element-based finite volume method.

Due to the complex geometries of turbomachinery and how vital a quality mesh is for

obtaining accurate results, ANSYS has developed a meshing tool specifically designed for

the meshing of turbomachinery domains. By linking the TurboGrid module to the BladeGen

module in WorkBench, TurboGrid automatically extracts the relevant fluid domain. With

77

the meshing capabilities of ANSYS, TurboGrid can quickly and automatically generate a

relative high quality hexahedral mesh following general recommendations for mesh quality

such as avoiding high grid stretching (aspect ratios), jumps in grid density, poor grid angles

(skewness and orthogonality) and using a finer, more regular grid in critical regions (i.e.

large gradients). TurboGrid added the Automatic Topology and Meshing option (ATM

Optimized) which provides the user the ability to refine the mesh in important areas, such

as in the boundary layer and at the leading and trailing edges then, once the specifics of the

mesh have been set, it is possible to increase or decrease the mesh density through a scaling

factor while maintaining the general form of the mesh which makes transitioning to finer

meshes for mesh independence studies a quick and easy process.

For the turbulence model to perform well the mesh resolution in the boundary layer

needs to be sufficient. The boundary layer is generally divided into three regions, the vis-

cous sublayer (region very close to the wall where viscous forces dominate), the logarithmic

layer (region where turbulence dominates) and the buffer layer (region between the viscous

sublayer and logarithmic layer where both effects are prevalent). Figure 4.2 provides an

illustration of these regions. This is modeled using the log law relationship originally devel-

Figure 4.2: Turbulent boundary layer flow regimes [74]

oped by the works of Prandtl and von K´arm´an and summarized in more detail in Pope [71]

78

and the ANSYS CFX-Solver Theory Guide [73]. The mathematical formula that describes

the relationship between mean velocity and the distance from the wall is given through the

following equations:

y+ =

uτ =

ρ∆yuτ

µ

(cid:17)1/2

(cid:16) τω

ρ

u+ =

Ut
uτ

=

1
κ

ln(y+) + C

(4.24)

(4.25)

(4.26)

where y+ is the dimensionless distance from the wall based on viscous lengths, u+ is the

dimensionless near wall velocity, uτ is the friction velocity, Ut is the known velocity tangent

to the wall at a distance ∆y from the wall, τω is the wall shear stress, κ is the von K´arm´an

constant and C is a log-layer constant that is dependent on wall roughness.

There are two approaches for modeling the flow near walls that are commonly employed.

One is through use of a wall function, like that which was just introduced, for modeling

the near wall region without resolving the boundary layer, which reduces the computational

requirements. The other is the Low-Reynolds-Number method which resolves the boundary

layer profile through use of very fine mesh elements in the direction normal to the wall, which

requires significantly more computational resources than the wall function method. CFX has
an automatic wall treatment method for k−ω based turbulence models, which automatically

transitions between wall function and low-Reynolds number grids without a loss in accuracy

[74]. It is based on the existence of analytical expressions for ω in both the viscous sublayer

and the logarithmic region. Through use of a blending scheme it is then possible to obtain

the proper contributions from each expression based on the near wall mesh density.

The Reynolds number based on the blade chord length, average relative velocity and

density was determined from the analytical design to be 25,500 which was used for specifying

the near wall element size in TurboGrid using the y+ method. For a wall function approach

y+ values for the first cell of 20 has been found to work well, but in general smaller values

79

give a better resolution but at increased computational cost. The mesh also needs to have

enough resolution along the blade surface to be able to capture any potential separation

phenomenon and so the target element count along the blade in the streamwise direction

was to be at least 100 elements. At least 20 elements were used around the leading and

trailing edges to provide the necessary resolution in these regions of interest. The element

size growth ratios were kept below 1.2 to ensure smooth transition between element sizes

when moving away from the walls [75].

4.4 CFX Solver

The ANSYS CFX 15.0 solver utilizes an element-based finite volume method to solve the

RANS equations.

It employs an implicit, coupled (pressure-velocity), compressible flow

solver. CFX makes use of a multigrid accelerated Incomplete Lower Upper (ILU) factoriza-

tion (matrix preconditioning) technique in order to solve the discrete system of linearized

equations. When performing a steady-state analysis in CFX, the time-step acts as a pseudo

transient under relaxation factor that helps to guide the approximate solutions toward the

steady-state solution and helps to reduce the number of iterations necessary to reach con-

vergence.

The governing equations must be discretized in order to be applied to each control volume.

This is done by first integrating the RANS equations over each control volume and then

applying Gauss’ divergence theorem to transform the volume integrals (containing divergence

or gradient operators) to surface integrals. These equations are then discretized such that the

volume integrals are split among each element sector (these are essentially sub-elements) and

then accumulated to the control volume in which the sector belongs. The surface integrals

are discretized with respect to the center of each sector’s surface (where the midpoints of

the surface edges are connected to the center of the surface and then the midpoints of these

new lines are referred to as the integration points). The integrals are then evaluated using

numerical integration. A discretized form of the conservation equation for a passive scalar

80

φ is given as:

(cid:18) ρφ − ρoφo

∀

∆t

(cid:19)

+

(cid:88)

ip

( ˙mφ)ip =

(cid:18)

(cid:88)

ip

(cid:19)

ip

Γef f

∂φ
∂xj

∆nj

+ Sφ∀

(4.27)

where ∀ is the control volume, ∆t is the time step, Γef f is the effective diffusion coefficient,
∆nj is the outward unit normal vector of the surface, the subscript ip denotes evaluation at
an integration point, and the superscript o refers to the previous time level.

The advection term (i.e.

the term containing ˙m in Equation 4.27) in the discretized

equations needs the integration point values of φ which need to be interpolated from the

nodal values. An upwind scheme is used for this:

φip = φup + β∇φ∆(cid:126)r

(4.28)

where φup is the value at the upwind node, (cid:126)r is the vector from the upwind node to the
integration point, and β and ∇φ are dependent on the type of upwind scheme chosen. A high
resolution scheme was used in determining β and ∇φ and are detailed in the CFX-Solver

Theory Guide [73].

A non-coupled or segregated solver solves the momentum equations first based on a

guessed pressure in order to obtain an equation for the pressure correction. This leads to

a potential for a large number of iterations required to obtain satisfactory convergence. A

coupled solver solves the mass-momentum equations as a single set. Advantages of this

method are its robustness, efficiency, generality and simplicity while the main disadvantage

is the large number of coefficients that must be stored [73].

Without the multigrid approach, the ILU factorization preconditioning technique perfor-

mance rapidly decreases as the number of mesh elements increases. Therefore, the multigrid

approach is utilized to significantly improve this limitation. For a given mesh size, iterative

solving techniques tend to only be efficient at reducing the errors that have a wavelength on

the order of the mesh spacing. This means that errors with longer wavelengths can take a

81

long time to disappear while those with shorter wavelengths disappear rather quickly. The

multigrid approach overcomes this by performing initial iterations on a fine mesh and later

iterations on a progressively coarser “virtual” mesh. This is done by merging the original

mesh elements to form larger ones. A system of discrete equations for the coarse mesh is

formed in a similar way by simply summing the fine mesh equations. The solution results

are then transferred from the coarsest mesh back to the original mesh to obtain an accurate

solution with much faster convergence rates.

There are two types of convergence that should be monitored to determine if the solution

error has been reduced to an acceptable level. The first is computational convergence which

is evaluated by monitoring quantities called residuals. If the solution to the linear system of

equations was exact then the result would be:

[A]{ϕ} = {b} or {b} − [A]{ϕ} = 0

(4.29)

where [A] is the coefficient matrix, {ϕ} is the solution vector and {b} is the right hand side.

However, since the solution vector is an approximate solution Equation 4.29 is not equal to

zero. This imbalance is called the raw residual as is given of the form:

{r} = {b} − [A]{ϕ}

(4.30)

It is common to employ some form of residual scaling or normalization procedure and then

monitor the normalized residuals for use as a convergence criteria. In CFX the normalized

residuals are of the form:

[˜rϕ] =

[rϕ]
ap∆ϕ

(4.31)

where ap is representative of the control volume coefficient and ∆ϕ represents the range of

the variable in the domain. It should be noted that because of the normalization procedure

used in CFX that the normalized residuals are independent of the initial conditions. This

82

means that in general for most steady state simulations that Root Mean Square RMS residual
values of 1 × 10−4 is relatively loose convergence, values of 1 × 10−5 is typically sufficient
for most engineering applications and values of 1 × 10−6 or lower are considered very tight

convergence. Ensuring the residuals have been reduced beyond a certain level is typically

not enough to consider the simulation converged though.

It is also necessary to monitor

relevant parameters of interest to ensure that they are also converged to steady state values.

The second type of convergence is known as grid convergence ensuring that the results

are independent of the mesh density. To determine if the results are mesh independent the

simulation is repeated using progressively more dense meshes. When the results stop varying

with increasing mesh density then the results are deemed mesh independent.

4.5 Simulation Setup

With the fluid domains constructed and meshed, the next step was to setup the simulation.

This was done by importing the mesh into CFX through CFX-Pre which is the ANSYS tool

used to fully define the simulation physics models, boundary conditions, initial conditions,

desired outputs and other solver options. A steady state analysis was performed with adia-

batic (no heat transfer), no slip walls. The fluid was specified to be air treated as an ideal

gas. For steady state turbomachinery simulations a good range of timesteps for estimating

the relevant timescales is between 0.1/ω to 1/ω [76]. To ensure the physics were captured

properly, the smaller end of this range was selected, which corresponded to a time-step of
6.1 × 10−5 s.

Specifying proper boundary conditions is critical to achieving reliable simulation results.

They must be specified to represent the desired physical situation as closely as possible

to achieve accurate results. For this simulation domain aside from the periodic boundary

conditions used to reduce the computational domain, there is one inlet and one outlet for

which conditions need to be specified. A robust choice for subsonic turbomachines not

operating near stall or choked conditions (i.e. near design conditions) is to specify the total

83

pressure and temperature at the inlet boundary and a mass flow rate at the outlet. This

pair of boundary condition choices also works well for extracting the necessary data from

the analytical design results as they are constant values at their respective locations rather

than data profiles. By specifying a total pressure and temperature inlet the static pressure

and velocity profiles are then determined by the solver. For the mass flow rate outlet, the

distribution is directly related to the mass flow just upstream while the specified total value is

enforced at every timestep. This is done by starting with the local mass flow rate distribution

and using this to estimate the total mass flow rate through the outlet. A scaling factor is

then determined by comparing the specified value to the estimated value which is used to

adjust the local mass flow rate values to ensure they sum to the specified total mass flow

rate. The pressure distribution is then unconstrained by the boundary condition and is an

implicit result [74]. The total temperature is specified as 298.15 K, the total pressure as

5 kPa and the mass flow rate outlet for one sector, or 1/19th of the total, as 0.0041 kg/s.

4.6 Convergence Results

For the analysis performed in this study the mass, momentum, and energy (not shown
here) equations converged to values of at least 1 × 10−5 within 225 timesteps on the highest

density mesh shown in Figure 4.3. The efficiencies converged to constant value (within 0.2 %

difference) after approximately 160 timesteps shown in Figure 4.4.

Initially, the stage was meshed with 201,201 elements as shown in Figure 4.5a. The

number of elements was then increased incrementally and the simulation rerun with each

mesh. The finest mesh simulated containing 1,553,052 elements can be seen in Figure 4.5b.

With the efficiencies as the main parameters of interest that required the highest number of

iterations to converge, they were chosen for monitoring grid convergence. Through inspection

of Figure 4.6 the values appear to have stopped varying above an element count of 1,000,000

elements. This has been shown to be a reasonable mesh size to be able to resolve the potential

flow features that occur in an axial compressor [75].

84

Figure 4.3: Governing equation RMS residuals convergence on the mesh with the highest
density

Figure 4.4: Efficiency convergence on the mesh with the highest density

85

(a) Coarse mesh - 201,201 elements

(b) Fine mesh - 1,553,052 elements

Figure 4.5: Comparison of coarsest and finest meshes used in grid convergence study

Figure 4.6: Mesh independence study results

86

1.646.48.811.213.616x 1057072747678808284868890ηNumber of ElementsGrid Convergence − Efficiency  ηpoly,t−tηisen,t−t4.7 Results

A range of mass flow rates were simulated to check the accuracy of the predicted operating

point with respect to the best point found through CFD. This is shown in Figure 4.7 where

the machine mass flow rates have been converted to the corrected mass flow rate through

Equation 3.36 and atmospheric reference conditions. The analytical design point is high-

lighted with red stars and indicates that the analytically predicted design point was indeed

the best operating point (or at least very close to it).

Figure 4.7: CFD compressor mapping results - Πt-t and ηpoly,t-t vs ˙mcorr

A summary of the relevant CFD results compared to the unadjusted, inviscid analytical

results is provided in Table 4.1. The CFD results satisfy the power constrain put forth

by the integrated motor. The design also meets the pressure ratio target of 1.1 as well as

the efficiency target of greater than 75 %. The original analytical results are also provided

along with the percent difference. As expected, the original analytical results overpredict

the performance of the rotor. The power (and indirectly torque) differs the most due to

being the product of the ˙m and ˜e which compounds the overprediction. The rotor did,

however, achieve a higher efficiency than the assumed value used for estimating the losses in

the analytical design.

87

1.41.51.61.71.81.921.041.051.061.071.081.091.11.111.121.131.147677787980818283Compressor MapTable 4.1: Performance results - original compared to CFD

Property Value

Units %Diff.

Original CFD

˙m
˙Q
˜e
P
τ
Πt-t
Γt-t
ηpoly,t-t
ηisen,t-t
φ
ψ

0.1001
1.85
16.96
1.70
1.03
1.165
1.057
80.00
79.60
0.72
0.49

0.0779
1.33
12.90
1.01
0.61
1.100
1.037
82.77
82.55
0.55
0.38

24.85
kg/s
m3/s
32.61
27.16
kJ/kg
kW 51.22
N m 51.22
−
5.71
−
1.91
−
3.41
−
3.64
−
26.10
−
26.01

Table 4.2 shows a comparison of the adjusted analytical results with the CFD results.

The percent difference values for the adjusted analytical results are much better than for the

original values, where the analytical results only slightly overpredict the performance.

Table 4.2: Performance results - adjusted compared to CFD

Property Value

Units %Diff.

Adjusted CFD

˙Q
˜e
P
τ
Πt-t
Γt-t
φ
ψ

1.44
13.20
1.32
0.63
1.127
1.044
0.56
0.39

1.33
12.90
1.01
0.61
1.100
1.037
0.55
0.38

m3/s
7.86
kJ/kg
2.30
kW 2.31
N m 2.31
−
2.36
−
0.69
−
1.21
−
1.10

Because the aim of the design process was more focused on achieving the specific geo-

metric goals pertaining to the novel composite impeller manufacturing method rather than

pushing the performance boundaries, the simple design approach that was implemented pro-

88

vides a quick and easy way to obtain impeller designs while having a reasonable prediction

of the performance. The development of the novel method of composite impeller manufac-

turing is still in it’s relative infancy. As more of the potential strengths and weaknesses of

the manufacturing process are discovered and quantified, the design space will expand and

this will allow for enhanced performance.

A plot of the pressure coefficient can be seen in Figure 4.8 for a span of 0.5. The

upper portion of this plot represents the pressure side of the blade while the lower portion

corresponds to the suction side of the blade. The large spikes at the leading edge are caused

Figure 4.8: Blade loading at midspan

by the streamlines of the flow adjusting from the free stream to following the contour of

the blade. This leads to the streamlines curving which generates a pressure gradient that

faces outward in the normal direction from the center of the radius of curvature. On the

suction side this is facing away from the blade and causes the negative spike seen on the

lower portion of the plot while the opposite is the case for the pressure side. The thickness

of the blade as well as the shape of the leading edge both contribute to the shape of this

spike. Typically, the spiked region is narrower, but in this case the flow streamlines require

89

00.10.20.30.40.50.60.70.80.91−5−4−3−2−1012345CpStreamwise (0−1)Blade Loading Chartmore distance to adjust to the contour of the blade due to the thickness requirements of the

composite manufacturing process. The suction side curve has a small indent in the range of

0.05–0.2 of the normalized streamwise direction that indicates there could be a small region

of flow separation (and reattachment) which can be verified in Figure 4.9 by the small region

of at or near zero velocity [58].

Figure 4.9: Leading edge velocity distribution with small separation bubble

Despite having satisfactory design values of both the dH and DF , it is possible to still

have some minor flow separation occurring. Plots of velocity created near the tip region were

used to investigate this further. It can be seen in Figure 4.10 that a small region of separation

develops from a span of approximately 80 % and beyond. The size of this separation can

be determined with contour plots of relative Mach number on a stream surface placed near

the trailing edge. It can be seen in Figure 4.12 that this region with relative Mach numbers

close to zero is isolated to just the corner of the suction side and shroud from approximately

80 % and above confirming findings from Figure 4.10.

Spanwise distributions of α, β, average incidence, and average deviation are shown in

90

(a) 70 % span

(b) 80 % span

(c) 90 % span

Figure 4.10: Blade-to-blade velocity distributions

Figure 4.11: Relative Mach number stream surface contour LE

Figures 4.13 and 4.14 at both the leading and trailing edge locations. The analytical values

have been included for comparison and there is good agreement between them. The effects of

boundary layer can clearly be distinguished on the CFD curves near the hub and shroud (span

0 and span 1 respectively). At the trailing edge the effects of the region of separation can be

seen again at spans above 0.9 in the plot of βCFD.

In this design approach the analytical

values for β become the actual blade angles for the CFD simulation so by comparing these

with the CFD results for β which are the actual flow angles it is possible to estimate the area

91

Figure 4.12: Relative Mach number stream surface contour TE

Figure 4.13: Analytical vs CFD comparison of spanwise distribution of absolute and relative
flow angles including average incidence - LE

averaged incidence and deviation angles. The same can be done with β at the trailing edge

to obtain an estimate for the average deviation and these results are given in Table 4.3. β

values above span 0.9 were omitted due to the flow separation skewing the values as shown

previously in Figure 4.14.

In Figures 4.15 and 4.16 contour plots of Cm and p in the meridional plane are shown.

92

0122436486072849610812000.10.20.30.40.50.60.70.80.91SpanFlow Angles at the LEFigure 4.14: Analytical vs CFD comparison of spanwise distribution of absolute and relative
flow angles including average deviation - TE

Table 4.3: Comparison of analytical and CFD blade and flow angles

Property

Value Units

Design Flow and Blade Angles

αave,LE
αave,TE
βave,LE
βave,TE
CFD Flow Angles

αave,LE
αave,TE
βave,LE
βave,TE
iave
δave

102.86
65.20
31.88
47.21

92.21
63.63
35.03
42.05
3.15
−5.16

°

°

°

°

°

°

°

°

°

°

The values of Cm accelerate a small amount as the flow moves from the inlet to outlet and the

cross sectional area reduces (both geometrically and due to boundary layer growth). Near

the shroud at the trailing edge there is a region of lower velocity caused by the separated

93

0122436486072849610812000.10.20.30.40.50.60.70.80.91SpanFlow Angles at the TEflow but overall it does not disturb the core flow excessively. The pressure contour plot

indicates the values of pressure increase relatively uniformly as the flow moves from inlet to

outlet as the contour lines are mostly vertical. These contour plots help to confirm the flow

is well behaved everywhere except for the aforementioned pocket of separation in the suction

side/shroud corner.

Figure 4.15: Circumferentially area averaged Cm contour in the meridional plane

Figure 4.16: Circumferentially area averaged static pressure contour in the meridional plane

With the rotor having met the design goals and because the focus was not on design

optimization at this stage, it was decided that this small amount of separation was acceptable.

94

If there was enough separation occurring to cause difficulty obtaining convergence (along with

noticeably reduced performance) then there would be need to return to the analytical design

and adjust certain constraints (such as dH or DF ) to be slightly more conservative before

generating a new design.

With the geometric limitations imposed on the design process for manufacturing pur-

poses, the performance of the rotor was deemed satisfactory for this preliminary case. The

approximate target pressure ratio and efficiency were achieved in CFD. These performance

baselines will be used for comparison with the future experimental results.

95

Chapter 5

Composite Manufacturing

5.1 Introduction

There is still significant room to further expand the potential as highlighted in Section 1.4

for the novel method of composite impeller manufacturing initially proposed by M¨uller that

has proven to be a promising new technology. The modifications to the novel manufacturing

process proposed in this thesis are centered around two main goals:

1. To enable the manufacture of more traditional axial turbomachine designs.

2. To provide high surface quality on the finished impeller.

In order to use this process to obtain an axial impeller that more closely resembles tradi-

tional designs additional fiber guidance than provided by a simple slotted cylinder mandrel

was to be provided. Previously wet winding and prepreg layup were used. For this study

it was proposed to layup the fibers while dry and then to infuse with resin. This served to

make the layup process cleaner and also to obtain a good surface finish. The eventual goal

is to refine this process into a modified form of RTM or VARTM to reduce the amount of

excess resin wasted and time spent extracting and cleaning up the impeller. It should be

possible to design the mandrel in a way that it is reusable and manufacture it from durable

material such as aluminum or stainless steel. However, for prototyping, to keep costs low

while providing flexibility for improvements to the design, the mandrels were 3D printed out

of plastic.

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5.2 Mandrel Modeling

Once an axial compressor has been designed that meets the criteria laid out in Section 2.3.1

the next step was to design a mandrel that allows this to be manufactured. The first step

was to export the design from BladeGen into a CAD file. This provided a model of the

periodic flow path that was simulated in CFD. With this it was then possible to extract the

exact fluid path between two blades (rather than the flow path centered around one blade).

This single flow path was then duplicated and patterned around the axis of rotation to give

a “negative” of the impeller. A disk was added which joined each individual flow path on

one side (leading or trailing edge) in order to fix these in place.

(a) CFD domain

(b) Flow path

(c) Basic mandrel design

Figure 5.1: Mandrel modeling progression

Many of the initial mandrel design choices were guided by either the geometric require-

ments of the impeller or for 3D printability. For example, if the impeller has a low h/t ratio

and therefore a larger change in hub radius from inlet to outlet, it can be preferable to place

the structural disk on the leading edge side in order to be better suited for 3D printing the

mandrel, so the 3D printer can build up from the largest area to the smallest area reducing

potential for sagging. If the impeller was going to be driven by a shaft rather than having

an integrated hub motor it could be necessary to make use of hub “fingers” to help the fiber

stay in this region as can be seen in Figure 5.1c. When an integrated hub motor is included

in the design the amount of space for fiber winding in the hub region is greatly reduced so

the use of these hub “fingers” might no longer be required.

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5.3 Mandrel Manufacturing

5.3.1 3D Printed

There are a wide range of 3D printing technologies available on the market today, with new

developments occurring all the time. They range from entry-level desktop machines that

print inexpensive plastic all the way to production-grade machines that print specialized

materials with higher quality. The quality of the printed part is therefore dependent on the

brand and type of printer (and printer settings), the quality of the CAD model and the

material being printed.

Initially, the mandrels were manufactured on smaller tabletop style printers utilizing a

Fused Filament Fabrication (FFF) technique. There are many different settings available to

control the feed rate, temperatures, layering, infill patterns, etc. and the optimal settings

can vary depending on both the material as well as the geometry of the part being printed.

With these entry-level machines, even with good feed rate and temperature settings, longer

print time seemed to correlate with an increased chance of print failure or reduced quality.

This is part of the reason for having the open space in the center of each flow path that

needed to be filled before infusing. The other goal was that this additional filler material

would be easier to remove in the post processing than the thermoplastic or cured resin.

The first materials used were Polyactic Acid (PLA) and Acrylonitrile Butadiene Styrene

(ABS) as they are the most common types of thermoplastic filament and provided enough

rigidity and toughness to withstand the winding and infusing process for the most part. On

some of the designs with the hub “fingers”, since they were rather thin there was occasionally

issues with breakage where there was poor layer bonding as seen in Figure 5.2. Another

example of the types of print errors that were experienced can be seen in Figure 5.3.

High Impact Polystyrene (HIPS) filament is another filament similar to ABS and PLA

which is also compatible with the tabletop style printers. It can be dissolved with limonene,

and so seemed to be an excellent option. However, the print quality attainable with these

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Figure 5.2: Printed mandrel with broken “finger”

Figure 5.3: Example of region of low print quality

entry-level printers was deemed insufficient. Because printing the entire model at 100 % infill

(i.e. as a solid) would increase the print times by a factor of 1.7 compared to printing at 50 %

99

infill leading to much higher chances of print failure, particularly on these hobbyist printers.

For this reason, there were voids introduced inside of the mandrel as shown in Figure 5.4.

The relatively low print quality resulted in resin seeping into these voids during the infusion

Figure 5.4: Printed infill at less than 100 %

process performed under vacuum to ensure good resin fill around the carbon fiber. A close

up of this infused inside of the mandrel is shown later in Figure 5.34 as highlighted the

dashed black box. This could have occurred because either minute holes existed in the shell

of the print or if there was poor layer bonding in the shell, the force of the vacuum pressure

could have caused a small opening to form. This occurred with all mandrels printed on the

entry-level printers due to quality limitations and was less noticeable until attempting to

remove the soluble material.

To prevent the resin from seeping into the voids inside of the mandrel as well as improve

the surface finish of the mandrel, a mandrel was printed on a production-grade photopoly-

mer printer, the Stratasys Objet260. These printers operate on a different principle than the

previously used 3D printers employing thermoplastic filament extrusion. By using a pho-

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Figure 5.5: Polyjet mandrel

topolymer liquid that is cured instantly using UV light smoother, higher quality parts can be

obtained without issues like warping, poor layer bonding or % infill. Comparing Figures 5.3

and 5.5 shows the higher quality of the Polyjet material developed by Stratasys Ltd.

[77].

This increase in quality comes with a significant increase in cost however.

Stratasys Ltd. also offers a Fused Deposition Modeling (FDM), their patented FFF

processes, line of production-grade printers that can print in either ABS or a soluble support

material. More details of this technology can be found in the technical application guide

distributed by Stratasys [78]. This was done on a Fortus 250mc by printing the entire mandrel

out of only the soluble support material, the process of extracting the cured impeller should

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be made significantly simpler. Even though these printers are using materials similar to the

entry-level versions, they are significantly more robust and precise, providing the reliable

operation needed for a production quality machine. As a result, they can print larger parts

with a much lower chance of failure or print errors occurring. This allowed the mandrel to

be successfully printed with out the empty space in the middle of the flow paths and at

100 % infill without failure. This mandrel, for manufacturing the impeller whose design has

been detailed in this work, is shown in Figure 5.6. By inverting the typical ABS/soluble

Figure 5.6: Soluble core dissolvable mandrel

support configuration, the gray material between the flow paths is actually ABS being used

as support for soluble material. It needs to be removed before the mandrel is coated with

a layer of sealant to prevent any irregularities from forming at the interface between the

composite resin and the soluble core material. Lastly, a coating of release agent was applied

to help ease the removal process.

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5.3.2 Cast Wax

An alternative to 3D printing the mandrel is to use wax castings to create the mandrel

in pieces. While this approach has more upfront costs involved with manufacturing the

mold, base plate and hub, the potential for a high quality finish makes it appealing. The

wax castings can be easier to remove without damaging the impeller and can be melted

down and recast. This can save on cost if many of the same impeller design were to be

manufactured as the machinable wax is inexpensive.

To produce each flow path out of wax, a three piece mold was manufactured out of

aluminum. Alignment pins are used to hold the pieces in place as well as to leave holes in

base of the wax casting for aligning the castings on to the base plate of the mandrel. Freeman

Blue Machinable Wax was used in creating these castings. The wax becomes soft at around
100 ◦C and achieves a low enough viscosity for pouring into the mold at approximately 165 ◦C

[79]. One challenge of manufacturing the wax castings relates to the coefficient of thermal

Figure 5.7: Mold of fluid path for wax casting

expansion (1.62 × 10−4 mm/(mm ◦C)) and how the molten wax cools during and after the

pouring process, which needs to be taken into account for high quality surface finish. By
preheating the mold before pouring the wax to approximately 100 ◦C on all inner surfaces

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the wax will cool more slowly and maintain its shape better. To ensure there are no air

(a) Wax mandrel base

(b) Positioning pins

Figure 5.8: Wax mandrel aluminum base

bubbles, especially in the bottom corners, it is necessary to stir the molten wax with a piece

of wire quickly after it has been poured into the mold. In this iteration the top of the mold

was left uncovered and the wax cooled the fastest at this location (as opposed to the heated

mold surfaces) and therefore this area experienced the most contraction. This contraction

effect can be seen on the top of the castings where the wax sunk in to form some voids. In

future attempts the top can be closed off and insulated to force the wax to cool more slowly

and evenly. Use of a wax sprue going into a small opening in this insulating cover could help

ensure a more complete fill of the mold cavity as the wax cools.

The hub portion of the mandrel was manufactured in four sections out of a more durable

high temperature plastic as shown in Figure 5.9a. This allows it to be removed and reused

after a little cleaning. Figure 5.9 shows that several small features have been added to assist

with positioning the hub pieces.

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(a) Hub pieces with position features

(b) Hub positioning features in base

Figure 5.9: Wax mandrel hub region

(a) Cast wax mandrel

(b) Wax mandrel close up

Figure 5.10: 1st cast wax mandrel

5.4 Fiber Winding Process

The novel method of manufacturing a composite impeller out of a continuous fiber allows for

different winding patterns to achieve the desired geometric and structural properties. This

was shown previously in Table 1.2 about the original star pattern designs. The carbon fiber

employed is a carbon fiber ribbon (tow) of 12 thousand filaments. The properties of the 12K

tow are provided in Table 5.1.

Initially, impellers were wound with either a single or double 12K carbon fiber tow. To

limit fiber bunching when winding on the new mandrel designs a carefully planned out wind-

ing pattern is essential because the domain is much more restricted than in the previous star

pattern designs. If it is wound upon a shaft and needs the hub “fingers” then this extra wind-

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Table 5.1: Carbon fiber properties [80]

Property

ρ
Tensile strength 4902
Tensile modulus
230

1.8

Value Units
g/cm3
MPa
GPa

ing step must also be accounted for in the pattern planning. Figure 5.11 shows the winding

pattern utilized for impellers with 12 blades attached to a shaft. The extra step needed to

(a) Step 1

(b) Step 2

(c) Step 3

(d) Step 4

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(e) Step 5

(f) Step 6

Figure 5.11: Basic winding steps

add fiber to the hub region can be seen in red in Figure 5.12. To prevent fiber bunching

at the start and finish of the hub step, the starting position should be alternated around

with each successive base loop. For example, assume step 1 has almost been completed but

rather than going between flow paths 11 and 12 a hub winding step is completed before

winding between flow paths 11 and 12 and then completing steps 2–6. Now on the next

base loop the hub winding step would be shifted to start at the end of step 2 before winding

between flow paths 9 and 10. In this way it is possible to keep the extra fiber occurring from

entering and exiting the hub winding step distributed throughout rather than building up

at one location. This process is continued until the mandrel is full of fiber with the fiber tow

exiting the mandrel radially at the shroud. Extra fiber is then able to be wound around the

outer radius of the mandrel to ensure the shroud has sufficient fiber content.

If a permanent magnet motor is to be integrated into the hub region then a negative of

the motor housing geometry is mounted in the hub region. Due to the size requirements of

the integrated motor, the radial space between the flow paths and the motor is significantly

reduced when compared to the drive shaft design. Therefore, in this configuration, the hub

“fingers” are omitted along with the additional hub winding step. In order to ensure good

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Figure 5.12: Hub winding step

Figure 5.13: Side view of winding in progress

fiber content around the motor, the hub motor negative was wound with fiber first and then

mounted into the mandrel as seen in Figure 5.14. This made it more convenient to begin the

winding pattern from the hub rather than the shroud, but in general the same basic winding

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(a) Unassembled

(b) Assembled

Figure 5.14: Mandrel with motor housing

pattern was utilized. This basic technique/pattern can be easily adapted to the desired

number of blades for a given application. Care must be exercised when winding around the

shaft or motor negative to not wind too far around the radius before continuing to the next

blade or this could lead to fiber bunching in this region.

Due to the large h/t of the design to be manufactured and tested for this research, the

new mandrel incorporated the area into which the hub motor will integrate as a part of the

mandrel. For the current iteration there is some flexibility in sizing the space for winding

between the hub and motor region. Because of this, a two part winding pattern was utilized

to help provide good fill in this region. Before starting loop 1, the central hub ID was pre-

wound until it was fully covered in fiber as shown in Figure 5.15. Figures 5.16 and 5.17

show only the first 6 steps of each loop but the number of steps in each loop is equal to the

number of blades (i.e. 19 for this rotor).

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Figure 5.15: Pre-winding of hub region

(a) Step 1

(b) Step 2

(c) Step 3

(d) Step 4

(e) Step 5

(f) Step 6

Figure 5.16: Integrated motor first 6 winding steps of loop 1

The current spacing with fiber can be seen in Figure 5.18. The fiber appears to fill the

available space well, however there is a small triangular space left behind by the crisscrossing

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(a) Step 1

(b) Step 2

(c) Step 3

(d) Step 4

(e) Step 5

(f) Step 6

Figure 5.17: Integrated motor first 6 winding steps of loop 2

nature of the fibers on the hub side of the fluid paths. This could be reduced by making

the space for winding in this area slightly smaller, but unless the structural testing indicates

this is a problematic area this does not seem necessary at the current stage. It would only

serve to make the winding process more challenging by providing less space to work in.

In the current process getting good fill on the open, trailing edge side of the mandrel can

be challenging. The approach is to extend the fluid paths outward following the curvature

of the blade surfaces. The goal is to provide extra space to essentially wind past where the

blade should end so when the impeller is extracted, there is good fill at the trailing edge.

Currently the mandrel was extended a few extra millimeters (4 mm in direction of blade

curvature). As can be seen in Figure 5.19 there is always a small portion of mandrel that the

fiber will begin to slip off of, if further winding is attempted. This problem was exacerbated

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Figure 5.18: Hub winding region of soluble core

with the first iteration of the wax mandrel due to the voids at the trailing edge in many of

the wax pieces.

Figure 5.19: Axial winding progress

5.5 Resin Infusion/Curing Process

After the fiber is wound onto the mandrel, it needs to be infused with resin and cured. The

resin being used is the SC15 resin and hardener developed by Applied Poleramic Inc.

It

is a high performance low viscosity resin that is well suited for VARTM processes. It is a

two-phase toughened epoxy that is cycloaliphatic amine cured. Some of the properties of

this resin are [81]:

1. Good damage resistance (in structural and ballistic applications)

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2. High resin and composite modulus, strength and toughness

3. High interlaminar shear strength

The SC15 epoxy was mixed with the hardener at a ratio of 70 % epoxy to 30 % hardener

which provided approximately 2 hours of work time before gelling begins to occur. This

is an exothermic reaction and in order to not overheat the plastic mandrels to the point

of deformation the curing cycle used in this work was lower than those proposed by other

research into this epoxy [82]. The curing was performed in a curing oven by ramping up
the temperature at 7 ◦C/min and holding at 57 ◦C for 5 hours. The final properties of the

resin are dependent on the curing cycle though so epoxy selection and different curing cycles

could be explored in the future.

To prepare for the infusion process, if the flow path had open centers then each of them

needed to be filled with caulk to prevent resin from filling these spaces. The idea being that

the caulk would be easier to remove post infusion than solid print material. This can be

seen in Figure 5.20. However, when using the dissolvable or wax mandrels this step was not

necessary. Next, the small diameter mandrels were fitted into a cylinder with an internal lip

to support the mandrel as shown in Figure 5.21, in order to contain the resin as it infuses

into the outer shroud region. The large diameter mandrels were placed in a bucket for the

infusion process.

The preliminary attempts at infusing were performed using a small vacuum chamber.

The mandrel assembly was placed into a small plastic bucket inside of the vacuum chamber.

Another small bucket containing the resin was left outside of the vacuum chamber. Polyethy-

lene vacuum tubing was used to connect this supply of resin to the vacuum chamber (and

subsequently the mandrel assembly inside) as seen in Figure 5.22. Through evacuation of

the chamber the resin is drawn into the mandrel. Holding the chamber at a low pressure

for an extended period helps to facilitate the degassing of the resin as well as improving the

dispersion of the resin thus helping to reducing the presence of voids in the impeller.

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Figure 5.20: Flow paths filled with caulk

Figure 5.21: Infusing cylinder

Figure 5.22: Vacuum chamber used in infusion process

5.6 Impeller Extraction Process

Once an impeller has been infused and cured it needs to undergo post processing. Due to

the primitive infusion process the initial impellers required significant work to extract them

from the excess resin. An example of a cured impeller that still needs to be extracted can

be seen in Figure 5.23.

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(a) Bottom side

(b) Top side

Figure 5.23: Post infusion

Progress of the extraction process is illustrated in Figure 5.24. At the current stage of

development of the manufacturing process, the “fingers” of the mandrel for a shafted rotor

remain embedded in the cured impeller.

The extraction process of the wax mandrel rotor is detailed in Figures 5.25 5.26 and 5.27.

The trailing edge face had to be machined down a small mount before the 4 piece hub region

shown previously in Figure 5.9a could be removed. Through inspection of the blades, the

total amount of material that needed to be removed was determined by identifying where

there was little fiber as highlighted in red in Figure 5.26. After this material was removed,

the rotor was mounted onto its hub adapter and the excess resin at the OD was turned down

on a lathe as shown in Figure 5.27.

Currently this process requires more machining than desired for manufacturing compos-

ites. Ideally, no machining would be required because when the tooling cuts the material it

breaks fibers and also exposes them to the surroundings which can provide starting points

for delamination or separation of fiber and matrix, which can causes part failure. Extra care

was needed in order to ensure that the impeller was not damaged during the post processing.

115

(a)

(b)

(c)

(d)

This was particularly important when working near the blades due to their relative small

thicknesses. By making improvements to the mandrel design and infusion process it will be

possible to achieve a VARTM manufacturing process that requires much less post processing

to produce a finished impeller. Suggestions for improving upon this process to reduce or

eliminate the need for machining are proposed in Appendix B.

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(e)

(f)

Figure 5.24: Extraction process of Shaft-v1.2

Figure 5.25: Extracting IM-v2.1 - trailing edge partially finished

117

Figure 5.26: Close up of TE region where winding became difficult due to wax piece voids

Figure 5.27: Removing excess resin at OD

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5.7 Results

5.7.1 Shaft Based Designs

The first attempts at using this improved novel composite manufacturing technique resulted

in crude impellers. These impellers were approximately 101.6 mm in diameter and 21 mm in

axial length. The initial mandrel was printed in pieces rather than as a single piece. It was

immediately clear that the small variations in print quality made it difficult for the pieces to

fit together well. In the future this should be more feasible with larger designs with reusable

mandrels that are manufactured out of something such as aluminum. Despite the quality of

(a) Leading edge

(b) Trailing edge

Figure 5.28: Shaft-v1.0

the initial mandrel, it was used to manufacture the first impeller Shaft-v1.0 to gain insight

into potential problems. It can be seen in Figure 5.28a that there was not enough fiber laid

up axial as the hub region is almost entirely resin. This indicated that the mandrel should

be extended further in the axial direction to allow the fiber to be wound up farther without

slipping off the mandrel while attempting to fill the rotor geometry. It can also be seen in

Figures 5.28a and 5.28b that the blades themselves exhibit a waviness and have excess resin

buildup on the surfaces. This was due to using the first mandrel despite it’s deficiencies.

In the assembly of this mandrel the pressure and suction sides of the flow path did not fit

119

together well with the hub and shroud pieces. Only the hub and shroud pieces were used

and connected with packaging tape. The lack of rigidity in the packaging tape led to the

waviness in the blades as well as the excess pockets of resin build up. These challenges were

avoided in subsequent iterations by moving to the one-piece mandrel design.

Figure 5.29: Shaft-v1.0-Large (10 in OD)

A larger version of this rotor Shaft-v1.0-Large was also simultaneously manufactured

to gain experience dealing with larger geometries. The outer diameter was scaled to ap-

proximately 10 in. The larger geometry was easier to work with but, as expected, took

significantly longer to lay up the fiber. The quality appears improved in Figure 5.29 as com-

pared to Figure 5.28 simply due to the ease of working with the larger geometry. However,

as previously mentioned the multi-piece 3D printed mandrel design limited the quality that

could be achieved.

Rotor Shaft-v1.1 was manufactured on a one piece mandrel. It can be seen in Figure 5.30

that the one piece mandrel design prevented the issue of waviness and excess resin pockets

in the blade region that the first rotors experienced. However, it can be seen that there are

120

places where the resin did not seem to infuse fully which contributes to a poor surface finish

in areas. On several of the blade’s leading edges it can be seen that there are regions where

(a) Leading edge

(b) Trailing edge

Figure 5.30: Shaft-v1.1

the fiber did not properly fill the blade path which resulted in portions of only resin that

would easily chip or break. This was thought to be caused by the blades being too thin in

combination with poor print quality on the blade surface in these regions. This effectively

prevented the fiber from being able to reach the bottom of the mandrel near the disk’s

surface. There is a portion of the mandrel structural disk remaining in the hub region on

the leading edge side, that was unable to be removed without damaging the leading edge of

the blades. With less than 100 % print infill, any tiny gaps between the layers allowed for

resin to seep in during the infusion process which added to how rigidly this portion of the

mandrel was attached to the rotor.

The next rotor Shaft-v1.2 was manufactured on a mandrel where the disk was located on

the trailing edge side of the flow paths rather than the leading edge. The mandrel quality was

significantly improved by having this mandrel printed out of the Polyjet material. This made

a significant difference in the surface finish of the blades as seen in Figure 5.31. Because of

the solid infill of the Polyjet mandrel, there was significantly less flexing than in the previous

mandrels, especially towards the end of the winding process and away from the structural

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support provided by the disk. This resulted in an overall more uniform blade thickness.

(a) Leading edge

(b) Trailing edge

Figure 5.31: Shaft-v1.2

This difference in thickness is highlighted by comparing the leading and trailing edges of the

blades in Figure 5.30 to those in Figure 5.31. The blade thickness of this model had not yet

been increased resulting in these blades being thin; however, because of the higher precision

and quality of the Polyjet mandrel the blades did not have the same issues as those in Shaft-

v1.1. It can also be seen in these that the shroud needed more additional winding on the

outer radius to obtain a better fiber volume ratio (this is especially clear in Figure 5.31a).

It was determined at this time that the increased quality of the Polyjet mandrel was not

currently worth the significant increase in cost.

The next impeller, Shaft-v1.3 shown in Figure 5.32, was manufactured on a PLA mandrel

with the structural disk on the trailing edge of the mandrel. It can be seen in Figure 5.32a

that there were small areas of low fiber content where the blades meet the shroud on the

leading edge. These likely resulted from the additional hub winding step contributing to

fiber bunching in the hub region which would explain why these only occur at the shroud.

The mandrel design could have been extended further axially so as to give more space to

layup fiber to help prevent these resin heavy spots in these structural important regions.

The quality of print layers between the blades, in conjunction with the blade thickness being

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small, led to difficulty in getting the fiber between the flow paths without some fiber breakage

in some locations. However, the fiber layup process went smoother on this mandrel than for

that of Shaft-v1.1 and the difference in fiber fill quality and overall infusion process be seen

by comparing Figure 5.30 and Figure 5.32.

(a) Leading edge

(b) Trailing edge

Figure 5.32: Shaft-v1.3

Due to the success of the previous rotor in terms of quality, another rotor Shaft-v1.4

was made on a nearly identical mandrel. The blade thickness was doubled to alleviate the

problems with winding the fiber through the thin blade passage. This way if there were print

errors in the blade passage it would be easier to fit a tool in to remove any plastic “hairs”

that the fiber could get stuck on. This had the added benefit of increasing the strength of the

rotor due to having thicker blades. The blades on the rotor from the Polyjet mandrel, which

had the most precision geometrically, seemed too thin to provide the structural strength

needed. Shaft-v1.4 was the best result of any of the iterations to this point. Through visual

inspection it appeared to have minimal locations with excess resin or poor fiber content (aside

from inside the hub region between the “fingers” and shaft which was the same for all the

shafted designs). The surface finish was better than any of the previous rotors made on the

inexpensive mandrels which could be attributed to applying several coatings of mold release

agent to the mandrel first. Because the surface finish of the mandrel is essentially imprinted

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(a) Leading edge

(b) Trailing edge

Figure 5.33: Shaft-v1.4

on to the surface of the rotor, none of the rotors created from FDM printed mandrels were

as smooth as the Polyjet mandrel which can essentially print smooth surfaces.

Rotor Shaft-v1.3, was of similar quality to Shaft-v1.4 except for the shroud where it can

be seen in Figure 5.34 that more fiber winding around the shroud was needed to prevent

the outer resin heavy layer as indicated by the red arrow. The difference in blade thickness

can also be seen here. These two rotors have a gloss finish because a coating of aerosol clear

coat was applied to make handling them safer by covering potentially sharp edges. This was

done to several of the rotors that were likely to be handled frequently.

Rotor Shaft-v1.5 was the first rotor manufactured on a HIPS mandrel with a single 12K

fiber tow (all previous rotors were also wound with a single fiber tow). As briefly mentioned

previously, the HIPS did not dissolve well. The appeared to be due to resin seeping into the

inside of the mandrel during the infusion process. This might be improved if the mandrel

was printed at a 100 % infill but this led to many problems with print failure, at least when

printing on the entry-level 3D printers. There are some locations where the blade meets the

shroud on the leading edge that appear to have low fiber content but overall the quality of

this rotor showed significant improvement from the first attempts and some consistency with

respect to Shaft-v1.4.

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Figure 5.34: Shaft-v1.3 v1.4 comparison

(a) Leading edge

(b) Trailing edge

Figure 5.35: Shaft-v1.5

A second impeller Shaft-v1.6 was also manufactured on a HIPS mandrel but was wound

with two 12K fiber tows at once. This helped to reduced the time needed to complete the

fiber layup process by approximately half. It also appears to have reduced the occurrence of

the small localized regions of low fiber content. Again, not considering the region between

the hub and shaft because this is bound to have lower fiber volume fraction due to the current

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mandrel design and winding scheme.

(a) Leading edge

(b) Trailing edge

Figure 5.36: Shaft-v1.6

Due to the complex geometry of turbomachine impellers it is difficult to measure quan-

tities such as the blade angles. To compare the resulting rotor geometry with the original

design, a NextEngine 3D Scanner Ultra HD was used to generate a scanned model of the

rotor. Figure 5.37 shows that this method of analysis is not without its own challenges. The

3D scanner works by mounting the object to be scanned into the scanner’s rotating part

holder. The scanner then shines lasers at the object and measures what is reflected back

from the part in order to generate the digital model. With this being a light based approach,

the surface type can have a significant effect on the quality of the scan [83]. Therefore, ideally

the surface would be a matte white color which means that the semi-glossy black surface of

the rotor was not well suited for the 3D scanning process. Without coating the impeller in

matte white paint, it proved difficult to achieve more complete scan results of inside the flow

passages but the general geometric properties are represented well enough for comparison

with the original CAD model of the impeller.

The original CAD model had the scanned model aligned over top of it with modified

hub region to allow the scanned model to be seen more clearly. In general, it appeared that

the resulting geometry was consistent with the design as shown in Figure 5.38. The axial

126

(a) Leading edge

(b) Trailing edge

Figure 5.37: 3D scan - Shaft-v1.5

length of the manufactured rotor did not quite reach the full design length because of the

issues related to winding the “open” side of the mandrel. The blade shapes and hub profile

appear to match the design well considering that this is the preliminary stage of developing

the manufacturing process.

127

(a) A

(b) B - LE

(c) C - TE

Figure 5.38: Impeller scan overlaid with CAD model for visual geometric comparison

5.7.2 Integrated Motor Based Designs

When manufacturing an impeller with an integrated motor instead of a shaft there were a

few differences which were introduced in Section 5.4. The main difference was that instead

of winding around a shaft the fiber tow was wound around a negative of the space for

mounting the motor. This provided a better fiber volume ratio in the hub region when

compared to those impellers constructed around a shaft which also helped to reduce issues

of fiber bunching in the hub region associated with the additional hub winding step. In order

to accommodate the integrated hub motor the model used for the Shaft-v1.4-1.6 was scaled

128

to 240 mm in outer diameter and 60.33 mm in axial length. The winding was performed with

two 12K fiber tows but using more could have saved time and helped the overall quality.

This could also be accomplished by using a single 50K tow for example.

The first prototype manufactured for use with an integrated motor IM-v1.0 can be seen

in Figure 5.39 after the extraction process. There was still a need for additional post-

processing in order to clean up the rotor, particularly on the blade trailing edges. Inspection

(a) Leading edge

(b) Trailing edge

Figure 5.39: IM-v1.0

of Figure 5.39b shows the off-white color that implies the impeller has a low fiber volume

ratio in this region. This corresponded with the open side of the mandrel. Since this was

a relatively consistent issue on the open side of the mandrel, this is an area to focus on

improving in future iterations. Due to the low fiber content, the structural strength was also

low here which led to chipping which can be seen on the edges of several of the blades as

well as the shroud.

The next version of the integrated motor design IM-v2.0 (aerodynamic design summarized

in Chapter 3), was manufactured on the soluble core mandrel. The larger h/t provided

its own set of challenges for pairing with an integrated hub motor. As can be seen in

Figure 5.40 the available space in the hub is significantly larger than compared to the low

129

h/t characteristic of a front stage of a multistage compressor like IM-v1.0.

In IM-v2.0 the main region that is lacking fiber is near the hub flow surface regions. In

Figure 5.41 bubbles left some undesirable voids in several of the hub flow surfaces. How-

ever, many of the hub flow surfaces did not have these defects as shown in Figure 5.42.

Improvements to the infusion process can minimize or eliminate the formation of these voids

in the future. The imprinted texture from the 3D printed mandrel is also shown clearly in

these figures. With the continuous fiber winding pattern through the blade and blend radii

at the hub, there may be enough reinforcement present to provide high strength in these

critical regions. However, further testing is needed to determine if these resin rich regions

are actually problematic with respect to the structural properties and failure limits of the

rotors. Independent of this efforts to improve the fiber content of these regions should be

considered in future work.

(a) Leading edge

(b) Trailing edge

Figure 5.40: IM-v2.0

The rotor manufactured with the wax mandrel, IM-v2.1, can be seen in Figure 5.43. This

rotor, as expected, had a much smoother surface as a result of the wax mandrel. In general,

the hub regions in this rotor were mostly free of the voids seen in IM-v2.0 except for 2–3

defects left by small bubbles. A close up of one of these is shown in Figure 5.44. It is thought

130

Figure 5.41: Voids in hub flow surface (LE) - IM-v2.0

Figure 5.42: Hub flow surface with no visible voids (TE) - IM-v2.0

that the reduction in voids is a result in the difference between the surface finish of the 3D

printed mandrel compared to the wax mandrel. The ridges in the 3D printed mandrel may

have provided more resistance for bubbles trying to move to the surface causing them to

remain trapped inside the composite leading to these voids. The smooth surface of the wax

did not provide this resistance allowing the bubbles to be successfully purged. Comparing

Figures 5.41 and 5.42 with 5.44 and 5.45, it can be seen that modification of the winding

pattern provided improved fiber content in the hub region. However, there is still potential

further to improve upon this in future work.

Geometric properties that could be measured with a dial caliper are compared to their

design values in Table 5.2. For rotor IM-v2.0, the resulting values for shroud OD, rt, rh,in and

rh,out are all in very good agreement with % error values of less than 0.5 %. The axial chord

131

(a) Leading edge

(b) Trailing edge

Figure 5.43: IM-v2.1

Figure 5.44: A small void in hub flow surface (LE) - IM-v2.1

length deviated from its design value further at 5.51 % error. However, this was actually

only shorter than the design specification by 1.37 mm. This is due to the way the winding

process currently finishes at the open, trailing edge side of the mandrel as mentioned at the

end of Section 5.4. To prevent having areas of little to no fiber content on the trail edge of

the blades, the impellers were machined down slightly farther than their design specifications

to remove these areas. By adjusting how far the mandrel is extended axially, this dimension

can be matched with better accuracy in future work. The blade thickness at the leading

and trailing edges are both larger than the design values with the leading edge 11.34 % error

132

Figure 5.45: Hub flow surface with no visible voids (TE) - IM-v2.1

and the trailing edge 17.02 % error, but because the blades are thin the actual differences

are only 0.23 mm and 0.34 mm respectively. This can be partially be attributed to the 3D

printed mandrel geometric tolerances in addition to a small amount of flex in the fluid paths

which allowed the blade regions to spread apart as fiber was packed into the slots.

For rotor IM-v2.1 the resulting values for shroud OD, rt, rh,in and rh,out are similarly

all in very good agreement also having 0.5 % error or lower. The blade thicknesses and

axial chord for this rotor differed more significantly than rotor IM-v2.0. The wax mandrel

manufacturing process requires trial and error to get the resulting geometries to match the

specified values due to the wax shrinking up to 7 % as it cools to room temperature [79]. As

expected, this effect manifested itself in the blade thickness measurements which had the

largest values of % error at 56 % and 54.52 % at the leading and trailing edges respectively.

These are equivalent to over 1 mm of thickness. The axial chord was off by 18.82 % from the

design value which was the most significant absolute difference at 4.70 mm. This is due to

the initial attempts at casting the fluid paths which resulted in some bubbling at the trailing

edge side as mention in Section 5.3.2 which limited how far fiber could be wound axially.

The reduced axial chord length is the main contributing factor to the 26 g difference in mass

between the two rotors. Again, the mandrels should be extended further, axially, to better

achieve the desired geometry.

133

Table 5.2: Geometry comparison

Property

Value
Design Measured Diff.

Abs. Units

(%)
Error

-
0.19
0.40
0.02
0.01
0.49
0.23
0.34
1.37

-
0.24
0.25
0.60
0.01
0.46
1.12
1.09
4.70

g

-

mm 0.07
mm 13.50
mm 0.01
mm 0.01
mm 0.48
mm 11.34
mm 17.02
mm 5.51

g

-

mm 0.10
mm 8.23
mm 0.49
mm 0.01
mm 0.45
mm 56.00
mm 54.52
mm 18.82

IM-v2.0

Mass
Shroud OD
Shroud thick.
rt
rh,in
rh,out
tb,in
tb,out
ac

IM-v2.1

Mass
Shroud OD
Shroud thick.
rt
rh,in
rh,out
tb,in
tb,out
ac

-
248.42
3.00
121.21
100.00
102.00
2.00
2.00
24.95

-
248.42
3.00
121.21
100.00
102.00
2.00
2.00
24.95

384
248.23
2.60
121.22
100.01
102.49
2.23
2.34
23.58

358
248.18
3.25
120.61
100.01
101.54
3.12
3.09
20.25

134

Chapter 6

Experimental Testing

6.1 Introduction

Before aerodynamic testing the electric motor needed to be integrated into the hub region

of the rotor. The rotor assembly then needed balanced to ensure the vibration levels were

below a safe threshold to prevent damaging the testing facility or the rotor assembly. The

rotor assembly would then be mounted inside the compressor section of the test rig. The

necessary electrical connections could then be made between the motor and VFD, as well

as the Data Acquisition (DAQ) system and sensors. Finally, after ensuring all the sensors,

monitors, and controls are working properly, testing could commence.

6.2 Electric Motor

6.2.1 Integrated Motor - IM-v1.0

The prototype was driven by an integrated Permanent Magnet Synchronous Motor, or

PMSM, (custom designs by McCleer Power Inc.) located inside the hub region of the rotor.

In a typical permanent magnet motor for driving a shaft, the stator and windings are located

at the outer radius while the permanent magnets are affixed to the shaft inside the stator

assembly. In this configuration, the setup was inverted where the stator was mounted onto a

stationary shaft and the permanent magnets were located at the outer radius of the motor.

The VFD then powers the motor.

Theoretically, using arc shaped permanent magnets would allow the motor to produce

more power but would be more costly than purchasing commercially available permanent

magnets. With a wide range of sizes rectangular permanent magnets available off the shelf,

135

an array of these stock magnets were used initially. It should be noted that a trade-off was

made here. In this case, no back iron is used while back iron was used with the custom arc

magnets in the second motor. This required that the bar magnets be thicker than the arc

magnets with back iron in order to achieve similar motor power. In general, use of back

iron in a PMSM design allows for smaller magnets and can provide more torque but also

decreased efficiency.

The rectangular magnets required a nonmagnetic spacer to affix them in the correct

positions. The stator and permanent magnet mounting spacer can be seen in Figure 6.1.

Figure 6.2 shows a sectioned view of this configuration without the stator or shaft.

Figure 6.1: TE view of integrated motor CAD design with magnet slots and stator

The stator functions as a single piece, but is actually manufactured as a set of laminations.

These are thin planar axial slices which are then stacked to meet the length requirements

of the motor design. It is done in this way for two main reasons: First, depending on the

size of the motor, it could simply not be feasible to manufacture the stator as a single piece

136

Figure 6.2: Sectioned view of integrated motor CAD design with motor cap

(a) Before taping

(b) After taping

Figure 6.3: Kapton tape application to stator laminations

while achieving the desired material properties and dimensions. Secondly, by using thin

laminations, each of which has an insulating coating, it is possible to prevent axial currents

from flowing in the core. These are driven by the magnetic field and cause considerable

heating and reduced efficiency if allowed to occur [84]. The stator laminations are made out

of a HF-10 electrical steel which is designed for use in high speed motors and generators.

137

(a) Start

(b) Finish

Figure 6.4: Stator winding process for 4 pole, 3 phase wye configuration PMSM

Permanent magnet motors operate with alternating current and therefore it is beneficial

to wind the stators with Litz wire. This is a special kind of wire that is designed to reduce

losses associated with transmitting alternating current. In the first integrated hub motor

design, the wires are fed through the hollow, stationary shaft (a simple piece of steel tubing)

to the stator laminations as seen in Figure 6.4a. The wire was carefully wound around the

stator before being fed back into the shaft and through the other side. The completed stator

and shaft assembly is shown in Figure 6.4b.

A model of the IM-v1.0 rotor was 3D printed to allow for the motor assembly to be

mocked up while the actual composite rotor was in the process of being manufactured. The

assembled prototype is shown in Figure 6.5.

Once rotor IM-v1.0 was finished being manufactured, it was then possible to install the

integrated motor. A small spacer was required to affix the magnets into position. This was

3D printed out of a composite PLA filament containing chopped carbon fiber for additional

strength. This piece fit into the space left behind by the motor housing insert shown pre-

viously in Figure 5.14. The shaft and stator assembly was then carefully installed into the

rotor. This process is shown in Figure 6.6. The completed assembly can then be seen in

Figure 6.7 with the leading and trailing edge motor caps securing the shaft, bearings and

138

(a) TE

(b) TE close up

(c) Assembly

Figure 6.5: Integrated motor mock up in 3D printed rotor

motor into place. When compared to Figure 5.39, it can be seen that the rotor has been

fully post-processed from the infusion process.

139

Figure 6.6: Installing integrated motor into IM-v1.0

(a) Leading edge

(b) Trailing edge

Figure 6.7: IM-v1.0 assembly with integrated motor installed

6.2.2 Integrated Motor - IM-v2.0

Integrating the motor with rotor IM-v2.0 required a different scheme to position it within the

space shown previously in Figure 5.40. For this iteration, custom neodymium arc magnets

and back iron were utilized. Several other parts also needed to be designed and manufactured

140

including the shaft, end caps, and spacers. These parts can be seen in Figure 6.8. Figure 6.8b

(a) Full view

(b) Sectioned view

Figure 6.8: CAD model of PMSM for integration with IM-v2.0

shows that the magnets are mounted to the cylindrical back iron which has a flange on one

side that connects to the motor end cap. On the other side is simply a spacer to keep the

rotor and hub adapter centered axially over the magnets and stator. The shaft has a channel

in it for passing the stator winding connections out and through for connection to the VFD.

The other portion of the shaft was left hollow for a potential cooling system addition in

future work.

Figure 6.9 shows the full CAD model with the integrated motor mounted into the hub

adapter with rotor. A total of 4 dowel pins (2 per side, offset 180°) were press-fit into the

motor caps and used to align the assembly with respect to the axis of rotation.

A hub adapter was designed to integrate the rotor with the electric motor. The outer

diameter was matched with the shape of the rotor inner diameter with lobes for torque

transfer and a lip on the leading edge side to prevent forward axial movement of the rotor.

The center region was designed to accommodate the 1 kW motor and shaft assembly. Other

design considerations were made for securing the trailing edge side of the rotor and reducing

the mass. Balancing was accomplished through a mass-removal approach as detailed in

141

(a) Full view

(b) Sectioned view

Figure 6.9: CAD model of IM-v2.0 with integrated motor

Appendix A. It was manufactured out of 6061 aluminum and is shown with the rotor and

without the motor in Figure 6.10.

Figure 6.11 shows the trailing edge side of the full assembly. The rotor is secured against

the lip on the leading edge with set of bolts and washers where the washers extend over the

composite rotor securing it in position. The motor leads exit the end of the shaft to allow

for connection with the VFD. The ring with radial etch marks is one of the balancing rings

required for using the mass-removal approach covered in Appendix A. It is possible to see

where mass has been removed in several places during this process.

142

Figure 6.10: IM-v2.0 mounted to the hub adapter - no motor

143

Figure 6.11: IM-v2.0 mounted to the hub adapter integrated motor assembly

144

6.2.3 Variable Frequency Drive

A variable frequency drive is needed to provide and control the power to the electric motor.

The custom high speed permanent magnet motors require a high frequency output VFD to

achieve their design speeds. For PMSMs, the VFD output frequency (in Hz) is related to

the Ω (in rpm) and the number of motor poles through the following:

Ω =

f ∗ 120
# of poles

(6.1)

From this relationship, it is clear why PMSMs are referred to as synchronous motors since

the rotational speed is directly proportional to the VFD output frequency. Most standard

VFDs have the maximum output frequency limited to between 400–600 Hz by their software,

but some models are able to have this limit increased through a software upgrade. This was

a major reason the Yaskawa A1000 series was chosen. This line of VFDs come limited

to 400 Hz maximum output frequency but can have the software flashed to increase the

maximum output frequency to 1000 Hz.

The specific model number of the drive chosen was the 4A0004. This drive is rated for

1.5 kW at 4.1 A. Among the myriad of configuration settings provided, the Yaskawa A1000

also includes a set of output terminals and corresponding functions which can be used in

conjunction with the DAQ system to monitor desired parameters. These include important

parameters such as the output frequency (and thus rotor speed), output current, and output

power.

Due to their inherent design, it is common for high speed motors to have a low impedance

which can lead to problems such as excessive peak current, low speed cogging, increased

torque ripple, or increased motor temperature. Because of this, it is often necessary to

employ a load or line reactor between the VFD and the motor. The line reactor increases

the impedance of the system and reduces the peak current (which prevents the VFD from

faulting due to drive overload) [85]. An MTE RL-403 line reactor was selected which was

145

designed for a 3 phase, 4 A, 2 hp (approximately 1.5 kW), and 5 % impedance at 480 VAC.

6.3 Assembly Balancing

When an object rotates there are inertial forces generated that act outward in the radial

direction which are referred to as centrifugal forces. When unbalanced, these forces lead to

unwanted vibrations which can cause a range of issues including noise, increased wear and

fatigue failure, and even catastrophic failure. It is therefore imperative that steps are taken

to ensure that these forces are kept below a minimum threshold to ensure the compressor can

operate safely. It is impossible to manufacture a rotating part that has these forces perfectly

balanced and therefore rotating parts and assemblies must usually always be balanced to

some degree before use. A summary of the balancing equipment and procedures used in this

work can be found in Appendix A.

6.4 Testing Facility

The experimental testing was to be performed on the test rig that was developed at MSU by

Pohl [52]. The goal was to provide a flexible and cost effective testing facility. Commercially

available parts were utilized when ever possible to keep costs low. Since it is to be a closed

system, the PVC pipe, sensors, and other attachments all need to be capable of withstanding

a range of pressures without failing or leaking significantly. The physical realization of the

schematic in Figure 2.1 is shown in Figure 6.12.

A vacuum system was needed in order to perform testing at conditions below atmospheric

pressure. There were two vacuum pumps available in the MSU Turbomachinery Lab to

handle this task. A larger unit, the Edwards E1M40, for quickly reaching lower pressures

and a smaller unit, the Robinair Vacumaster 15500, which can be used for maintaining

a constant pressure despite leakages. The E1M40 unit is also capable of being utilized

with water vapor which could be desirable for future testing. A summary of the pertinent

146

Figure 6.12: Aerodynamic testing facility [52]

specifications for these vacuum pumps is found in Table 6.1.

Table 6.1: Vacuum pump specifications [86] [87]

Property

Value

Model
Power
Flow Rate
Ultimate Vacuum Rating
Connector

Vacumaster 15500 E1M40
249
8.5
4.6
3/4(cid:48)(cid:48), 3/8(cid:48)(cid:48) Flare

1100
50.5
5
ISO-KF40

Units

−
W

m3/h
Pa
−

There are several areas within the compressor section where data acquisition is desired.

A potential arrangement, which provides the possibility of incorporating guide vanes, that

illustrates the planes of interest is shown in Figure 6.13. By having traversable probes, it

would then be possible to obtain radial profiles of the relevant pressure and temperature

data.

There are a number of different sensors that were available to be employed for gathering

data and monitoring the compressor during operation. A summary of the currently available

147

Figure 6.13: Possible compressor section probe arrangement [52]

sensors is given in Table 6.2. These include sensors for measuring pressure and temperature

Table 6.2: Available sensors

Type

Model

Dwyer 5 series J type
Thermocouple
Omega PRCU 100 Ω
RTD
USC PDC type
Pitot-Static
Kiel
USC KCB type
Abs. Press. Trans. Omega Custom
Rel. Press. Trans. Allsensors 30INCH-D-MV 0–7465
0–100
Rel. Humidity
−20–80
0.5–10,000

Mitchell PCMINI52

IMI Sensors 603C1

Accelerometer

Range
−40–333
−50–260
± 10
± 48;± 45
5

°

°

Units Accuracy
◦C
◦C

±2.5 ◦C
± 0.25 ◦C
± 3 %,1 %
up to M =1.0
psia ± 0.03 %FSO
± 0.25 %FSO
Pa
< ± 2 %
%
± 0.2 ◦C
◦C
± 50 g’s
Hz

as well as the relative humidity to account for potential fluctuations of the atmospheric

conditions in the laboratory. The rotational speed can be determined directly through the

frequency of the VFD when using a PMSM as the output frequency and rotational frequency

are synchronous. It is also possible to deduce the power and torque being supplied to the

motor through the VFD by monitoring the voltage and the current.

A National Instruments (NI) CompactDAQ system is a modular setup that consists of

a chassis into which different modules can be inserted depending on the need. This was

used to interface the sensors and controls to a custom NI LabVIEW program that logs all of

148

the relevant data signals. A MATLAB code was created to process the measurement data

and assist with the analysis and presentation of the results. A detailed presentation of the

development, potential, and use of the testing facility and these systems and programs can

be found in the work by Pohl [52].

6.5 Compressor Integration

The compressor assembly was mounted into a 1.2 m length of PVC tube as shown in the

schematic in Figure 2.1. There were several important points to be considered. First, because

testing was to be performed under vacuum, it was critical to ensure that it was possible to

seal around whatever mounting features were devised. This included any wiring that needed

to be passed out of the loop, such as the motor leads. Second, there needed to be a way to

mount the hub annuli and bullet nose caps that serve to transition the flow from the fully

open pipe to match the rotor inlet and outlet hub diameters. Lastly, it needed to secure the

shaft ends into place.

Figure 6.14: Compressor assembly with hub annuli and mounting stanchions

149

Figure 6.14 shows the assembly that was to be mounted into the pipe section. There are

two sets of three stanchions spaced at 120° apart. Axially they are located approximately

47.5 mm from the centerline of the rotor in both directions. The ends of the stanchions that

face outward radially pass through compression fittings that are mounted into the pipe at

the corresponding angular positions. These seal around the stanchions and then provide a

surface for the securing nuts to tighten against. Through use of a hollow stanchion rod, it

is possible to pass any necessary wires through before sealing around them with caulk. The

Figure 6.15: Sectioned view - CAD model of compressor section of test loop

other end of the stanchions pass through the hub annuli where they were connected to the

shaft mounts as shown in Figure 6.15. The bullet nose caps were then mounted to the hub

annuli before the compressor section was installed in to the test loop.

6.6 Testing Results

6.6.1 IM-v1.0

Initial testing of IM-v1.0 was performed with the assembly placed inside of a vacuum cham-

ber which was held at a pressure of 698.5 mmHg vacuum (commonly referred to as coarse

vacuum) to reduce the aerodynamic load. An accelerometer was mounted to the shaft mount

150

at the trailing edge to allow for vibration monitoring. The quality of this rotor with inte-

grated motor, in terms of alignment and unbalance, was the main limiting factor in the speed

that was able to be reached. The maximum speed achieved before vibratory loads became

too large was 7140 rpm which corresponded to a tip speed of approximately 90.5 m/s.

A Fast Fourier Transform (FFT) of the accelerometer data provided some additional

insight into what vibrational frequencies were present in the system. The FFT analysis

was performed over a 0.5 s period while the rotor was spinning at 7140 rpm and the results

are shown in Figure 6.16. There are are peaks at 1X, 2X, 3X and 4X multiples of the

excitation frequency (shaft rotational speed). This is typically indicative of shaft and/or

bearing misalignment and mechanical looseness. This is logical because of the steel tube

that was used as a cost effective shaft for this first prototype was not precision ground for

straightness or proper bearing fit [88] [89].

Figure 6.16: FFT of 0.5 s of accelerometer data during steady state operation at 119 Hz
excitation frequency

151

050100150200250300350400450500Frequency (Hz)00.10.20.30.40.50.60.70.80.91AmplitudeFFT at 119 Hz Excitation Frequency6.6.2 IM-v2.0

6.6.2.1 Aerodynamic Results

The rotor assembly was balanced at 1/10 of the operational speed to within a G6.3 tolerance

prior to installation into the test loop. For the first aerodynamic test of IM-v2.0, the test loop
was evacuated down near its design operating pressure of 5 kPa (4325 Pa and at 20.9 ◦C).

The VFD was then switched on, which allowed the motor to begin ramping up to the

design speed of 15,705 rpm which corresponds to a tip speed of 200 m/s. The total-to-total

differential pressure was measured across the rotor which can be seen in Figure 6.17. Due to

the proximity of the test stand to the VFD the original data had significant noise shown in

blue. To improve readability, the data was averaged using a locally weighted linear regression

model shown in orange. The total pressure difference measured from startup until shutdown

Figure 6.17: Total-to-total pressure difference measured across rotor with averaged (orange)
data overlaid onto original (blue) data

was approximately 200 Pa. This equates to a pressure ratio of 1.046. Despite efforts to

balance the assembly to a satisfactory level, the vibration levels were high enough at the

maximum operational speed (as discussed further in the next section) warrant shutting off

152

470502534566598630662694726758790Time (s)-200-160-120-80-4004080120160200Total-to-Total Differential Pressure Monitorthe VFD before any potential catastrophic failure took place.

6.6.2.2 Vibration Analysis

While performing the aerodynamic test, two accelerometers were used to monitor and record

vibration signals throughout the test to ensure that the vibratory loads were safe while

during operation. FFT analysis was then used to determine which frequencies were present

providing insight into potential problems with the assembly.

The International Organization for Standardization (ISO) standard 10816-3 (for indus-

trial sized machines 15 kW or larger with integrated driver) provides general guidelines for

evaluation vibration but these estimates might vary for specific machines and installations

[90]. The standard divides these guidelines into two based on if it has a rigid or flexible

mounting system. A mounting system is classified as rigid when the amplitude of the low-

est natural frequency exceeds the main excitation frequency amplitude by more than 25 %

otherwise it is considered flexible. For the operating speed of 15,705 rpm, the 4 zones of

operation based on this standard are shown in Table 6.3 for both rigid and flexible mounting

configurations. While this standard is technically for machines of larger power than the 1 kW

device tested here, it can still provide an idea of acceptable ranges.

Table 6.3: ISO 10816-3 vibration severity zones at approximately 15,700 rpm [90]

Zone Description

Range
Rigid

Flexible

A
B
C
D

New machine equivalent
Acceptable for long-term operation 2.33–4.75
4.75–7.5
Acceptable for limited operation
Potential for damage
7.5+

0–2.33

0–4
4–7.5
7.5–12
12+

Units

g(cid:48)s
g(cid:48)s
g(cid:48)s
g(cid:48)s

A sample of 0.5 s of accelerometer data during full speed operation was analyzed from each

accelerometer (one at the inlet side and one at the outlet side - mounted on the stanchions)

can be seen in Figures 6.18 and 6.19. The amplitudes peak around 10g’s. The inlet channel

153

had noticeably higher vibration levels with most peaks over 10g’s and some peaks over

13g’s. However for the outlet, the majority of the peaks are under 10g’s with only a few

peaks crossing the 10g threshold. Even without yet knowing if the mounting scheme fell into

the rigid or flexible regime, this level was not insignificant. In order to prevent damage to

the machine, it was not run for very long with these vibration levels. This was despite the

rotor assembly being well balanced for its operating conditions, which indicated that there

was most likely some problem with the shaft mounting system leading to such high levels of

vibration.

Figure 6.18: Inlet-accelerometer data during 0.5 s period of operation at design speed

Again, FFT analysis was performed on the accelerometer readings to provide insight into

the health of the machine during operation. Two plots for each set of accelerometer data

were created. Figures 6.20 and 6.21 show the same FFT of the inlet accelerometer data, but

the first indicates the locations of the higher harmonics while the second shows the locations

of the NX/2 subharmonics. Figures 6.22 and 6.23 show these for the accelerometer data at

the outlet.

Inspection of these plots also indicates there the main excitation frequency is by far the

most dominant frequency without any lower natural frequencies of comparable amplitude.

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00.050.10.150.20.250.30.350.40.450.5Time (s)-15-10-5051015Acceleration (g)Accelerometer at InletFigure 6.19: Outlet-accelerometer data during 0.5 s period of operation at design speed

Figure 6.20: Inlet-FFT of accelerometer signal with first 6 higher harmonics labeled

This means that the support system can be considered flexible which is common for turbo-

generators and compressors [90]. From Table 6.3 this means that it potentially could be safe

to operate for short periods with vibrational magnitudes of this load, but again these are

only general guidelines so caution should be exercised.

155

00.050.10.150.20.250.30.350.40.450.5Time (s)-15-10-5051015Acceleration (g)Accelerometer at Outlet02004006008001000120014001600Frequency (Hz)00.511.52AmplitudeFFT of Accelerometer Signal at Inlet - Higher HarmonicsFigure 6.21: Inlet-FFT of accelerometer signal with NX/2 harmonics labeled

Figure 6.22: Outlet-FFT of accelerometer signal with first 6 higher harmonics labeled

As can be seen in both sets of FFT plots, the largest amplitude spike occurs at a frequency

of approximately 261.75 Hz, which is the rotor rotation frequency (this is not the same as

the VFD output frequency which is related to the rotational speed through Equation 6.1).

In Figure 6.22 there are also noticeable spikes at the 2–6X harmonics. The presence of a

very large 1X harmonic can typically be caused by the rotor imbalance and/or mechanical

156

02004006008001000120014001600Frequency (Hz)00.511.52AmplitudeFFT of Accelerometer Signal at Inlet - Half Harmonics02004006008001000120014001600Frequency (Hz)00.511.52AmplitudeFFT of Accelerometer Signal at Outlet - Higher HarmonicsFigure 6.23: Outlet-FFT of accelerometer signal with NX/2 harmonics labeled

looseness[88]. Mechanical looseness can also lead to the presence of many harmonics, both

integer and non-integer, as can be seen in [89]. This could have been caused by a small

amount of play in one of the shaft bearings due to improper machining that did not match

the specifications. This was noticed during the assembly process as on one side the bearing

would slide on and off with ease rather than with resistance that would be characteristic of the

close tolerance that was specified. The shaft mounting system most likely also contributed

to the vibration issues since the rotor assembly was able to be satisfactorily balanced before

being installed into this system.

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02004006008001000120014001600Frequency (Hz)00.511.52AmplitudeFFT of Accelerometer Signal at Outlet - Half HarmonicsChapter 7

Spin Pit Testing

7.1 Introduction

It has been hypothesized that having the continuous fibers wound in the radial direction (the

primary direction of large centrifugal forces at high rotational speed) of the blade and then

integrated into the shroud provide added structural strength and rigidity when compared

to traditionally manufactured composite rotors. The traditional mounting method causes

undesirable stress raisers where the blades are mounted into the rotor hub. The free ends

can also be more susceptible to vibration when compared to the one-piece composite rotors

produced by the method detailed in this research. The focus of this chapter is to begin

exploring the structural and tip speed limits of these rotors through spin pit testing.

7.2 Spin Pit

To test the hypothesis that this new manufacturing technique results in high strength rotors,

it is necessary to perform proof or overspeed spin testing. Spin testing is critical for high

speed rotating machinery due to the large amount kinetic energy that can be released during

part failure. To get an idea of the magnitude of this kinetic energy, consider a steel disk 14 in
in diameter by 3 in thick, rotating at 27,000 rpm has a kinetic energy of about 3,000,000 ft − lb

which is roughly equivalent to five full-size automobiles traveling together at 60 mph [91].

This makes it clear that safety should be of utmost importance when spin testing. While the

rotational kinetic energy in composite rotors is typically lower than in a comparable metal

rotor due to the lower mass (and therefore lower moment of inertia), it is still significant

enough that special care must be taken when spin testing. For this reason, spin test systems

158

are designed to contain the fragments during a failure event. In addition, these containment

systems referred to as ”spin pits”, are typically located inside a blast proof test cell that

isolates it from the operator for additional safety.

A spin pit is a large heavy duty cylindrical vacuum chamber that has several layers

of ballistic material lining the inside diameter. Testing is almost always performed under

vacuum for several reasons. First, it reduces the aerodynamic load associated with rotating

in a fluid medium which can make it difficult and costly to drive the rotors to testing speeds.

Secondly, the fluid could also cause unsteady aerodynamic forces which could destabilize

the rotor leading to inaccurate test results. Lastly, the vacuum acts as a safety feature by

reducing the risk of explosion of metal, composite dust, or oil fog during a failure event [92].

The spin testing facility that has previously been developed at Michigan State University

is shown in Figure 7.1 and detailed further by Vander Klok [56]. The spin pit is isolated inside

a blast proof room for additional safety. It consists of a vacuum chamber that is 1829 mm

in diameter and is 1016 mm in height. Inside there are two layers of ballistic liners, a 63 mm

thick forged steel ring, and a layer of 200 mm thick ballistic rubber bricks installed around

the inner diameter for containing fragments during a bursting event. The drive turbine is

mounted to a 720 mm diameter removable bulkhead. There are nine total 305 mm diameter

view ports with one on the bottom and the remaining eight located in a circular array on

the top cover.

The spin pit is equipped with a BSi Model 6100 air turbine from Barbour Stockwell.

This turbine is rated for spinning a maximum test weight of 400 kg with a maximum power

output of 40 hp at 40,000 rpm. The bearings have an integrated oil lubrication and cooling

system in addition to squeeze film dampers which help to isolate the turbine from the forces

generated during testing. Barbour Stockwell supplies the turbine with the control system

and software necessary for operation and monitoring of parameters such as the rotational

speed, acceleration rate and valve position. Braking triggers, referred to as faults, can be

linked to critical parameters such as turbine oil level, oil pressure and oil flow rate, or trip

159

Figure 7.1: Spin testing facility

wires. The turbine requires an external air supply to power it, requiring 700 cfm to achieve

maximum performance. However, lower flow rate compressors can be used if maximum power

or torque is not required. A diesel compressor rated for 375 cfm was utilized and provided

ample flow to achieve 40,000 rpm while testing small, lightweight rotors.

7.3 Hub Adapter Redesign

The rotor needed to be mounted and coupled to the drive turbine. The shaft of the drive

turbine is 0.5625 in in diameter so the hub adapter design needed to be modified for it to

be compatible. Figure 7.2 shows the CAD model of the modified hub adapter mounted to

160

the turbine shaft. The centrifugal forces the hub adapter would experience at the maximum

Figure 7.2: CAD model of modified hub adapter and shaft for spin testing

speed of the air turbine will be 6.5 times larger than those experienced while operating at the

aerodynamic design rotational speed of 15,705 rpm. To allow the hub adapter to withstand

these forces, several other design changes were made.

The first change involved replacing the mass removal balancing approach to a mass ad-

dition where the array of smaller diameter holes is for the placement of balancing weights.

A larger hole array was added for high strength bolts to be used in securing the rotor to

the adapter. Radial reinforcing ribs were added to connect the inner and outer diameter

regions to provide increased stiffness. The high strength steel alloy 17–4PH stainless steel

treated to H900 condition was chosen because this is the precipitation hardened condition

which exhibits the highest strength properties. This was also selected because it comes in an

easy-to-machine solution-annealed condition, referred to as condition A. After machining,

the part is subjected to a mild heat treatment process to obtain high strength and corrosion

resistance properties with minimal shrinkage or warping. The resulting material properties

are comparable to grade 5 titanium except with respect to the density which is 1.76 times

larger than that of the titanium. While the lower density of grade 5 titanium would have

a smaller moment of inertia and thus rotational kinetic energy than a steel alloy, it is sig-

nificantly more expensive and difficult to manufacture. Titanium bolts, nuts and washers

161

were selected to secure the composite rotor to the hub adapter. The properties of these two

materials are given in Tables 7.1 and 7.2.

Table 7.1: 17–4PH H900 material properties
[93]

Table 7.2: Grade 5 titanium alloy material
properties [94]

Property
17–4PH H900

Value Units

Property
Grade 5 Ti Alloy

Value Units

7800
ρ
197
Young’s Modulus
0.272
Poisson Ratio
144
Bulk Modulus
77.44
Shear Modulus
Ten. Yield Strength
1.379
Ten. Ultimate Strength 1.448

kg/m3
GPa
−
GPa
GPa
GPa
GPa

4430
ρ
114
Young’s Modulus
0.330
Poisson Ratio
112
Bulk Modulus
42.86
Shear Modulus
Ten. Yield Strength
1.10
Ten. Ultimate Strength 1.17

kg/m3
GPa
−
GPa
GPa
GPa
GPa

7.3.1 Structural Analysis

7.3.1.1

Introduction to FEA

FEA was performed on the modified hub adapter design to determine how the assembly

would handle the large centrifugal loads. This was accomplished with the ANSYS Mechanical

module through a static structural analysis which determined the displacements, stresses,

and strains caused by the bolt clamping in addition to the centrifugal effects. The CAD

assembly needed to be simplified before beginning to setup the simulation.

The CAD assembly was reduced to contain only the hub adapter and the bolts, nuts, and

washers for securing the rotor in place. The assembly was then reduced further through the

axisymmetric conditions by considering only a single 90° segment. The bolt and nut threads

were removed to further simplify the simulation domain because ANSYS Mechanical has

a simplified bolt modeling tool that can provide nearly the same accuracy as true thread

modeling through specification of the desired thread details of mean pitch diameter, pitch

and half thread angle [95]. Figure 7.3 shows the model that was used for the structural

162

analysis.

Figure 7.3: Modified hub adapter - reduced geometric domain for FEA

After the geometry was imported, the material properties were applied to the appropri-

ate individual components. In this simulation, because the metals are assumed to exhibit

comparable strength properties in all directions, the corresponding material properties in

ANSYS were specified to be isotropic and taken from Tables 7.1 and 7.2.

Next, the cyclic region symmetry conditions were applied to the appropriate periodic

surfaces. The interfaces, or contact connections, were defined where every contact interface

was defined as a frictional interface with a specified coefficient of friction. Values for the

coefficient of friction between grade 5 titanium alloy acting on itself was found to be 0.35

for static conditions and 0.3 for sliding conditions [96]. However, values for grade 5 titanium

alloy acting against 17–4PH H900 were not readily available and had to be estimated from

other similar materials. Due to this a conservative value of 0.3 was chosen and applied to all

of the appropriate contact regions. The two primary forces that the bolt and nut assembly

will experience are the preload force and the centrifugal force. The main concern for failure

to occur was expected to be from shearing of the bolt shaft rather than tension loading which

163

could cause the threads to fail. Due to this, the simplified bolt thread modeling feature in

ANSYS Mechanical was deemed a good approximation because they should not undergo

large deformations which is when the model starts to break down. The threads are 5/16”-18

which corresponds to a mean pitch diameter of 0.2764 in and a pitch of 0.0556 in. The half

pitch angle was specified as 60° which is the standard for most common threads.

Each component was meshed, with a focus on providing a quality mesh in regions of

interest. These regions included all of the holes in the hub adapters as well as the bolt,

washer and nut interfaces. Specifically, the thread model required that the mesh in this

region have elements smaller than 0.25*pitch or 0.056 mm. The initial mesh had a total

of 113,819 nodes and 56,214 elements while the most refined mesh had 418,806 nodes and

263,818 elements.

Next, the simulation conditions were specified. This included several kinds of bound-

ary conditions consisting of support, load, and inertial types. The shaft hole was given a

cylindrical support boundary condition that fixed the cylindrical surface from moving or

deforming in the radial, axial or circumferential directions. This may not be a perfect rep-

resentation of what will be taking place physically, because in reality the central hole could

expand a small amount under the centrifugal load. However, due to the small radius of this

hole compared to the much larger radius where the main areas of interest are located this

simplifying assumption was deemed to be a reasonable approximation. A simple calculation

of the magnitude of the centrifugal force for a fixed mass and rotational speed at two radii

can assist with confirming this. The magnitude of the centrifugal force can be found from

[97]:

F = mω2r

(7.1)

For a fixed mass and rotational speed, the ratio of the centrifugal force at the radius of

the bolt hole in the hub adapter to the inner shaft hole radius is 15.4 times larger. When

factoring in that there is more mass at the outer radius, the centrifugal force is skewed even

farther towards the outer radius.

164

A load type boundary condition was applied to simulate the bolt preload condition. Based

on the torque specifications for 5/16”-18 for grade 5 titanium, a preload force of 20 kN was

specified [98]. The ratio of the ultimate tensile strength of the bolt to the magnitude of the

stress caused by the preload force was 2.05, while the ratio of the bolt shear strength to

the estimated shear force at maximum rotational speed was 1.02 confirmed that the most

likely method of failure for the bolt would be by shearing due to the centrifugal load. The

rotational velocity was accounted for through the application of an inertial load and was

initially specified as the maximum rotational speed of the air turbine (40,000 rpm). To

mimic reality, these different loads needed to be applied in a specific order, or load step. The

bolt pretension was applied on the first load step and then locked for the remaining load

steps. This is comparable to tightening down the bolts before rotating the assembly. The

inertial load was applied last.

The loads applied to the model induce stresses and strains into the materials. Stress, σ,

is defined as the ratio of the applied force, F , divided by the area of the cross-section on

which it is applied, A. The strain is the ratio of the change in length, δ, to the undeformed

length, L. The relationships for linear elastic isotropic materials are expressed through the

generalized Hooke’s Law which is given in matrix form as[99]:

E(1 − ν)

Eν

Eν

0



σxx

σyy

σzz

τxy

τyz

τzx



=



Eν

Eν

0

0

0

E(1 − ν)

Eν

0

0

0

Eν

0
E(1 − ν) 0

0

0

0

0

0

0

G 0

0 G 0

0

0 G

0

0

0

0






xx


yy


zz

γxy

γyz

γzx



·

(7.2)

where σij and τij represent the normal and shear stresses on the ith face in the jth direction,


ij and γij represent the corresponding normal and shear strains. E is known as Young’s

modulus or modulus of elasticity G is the shear modulus or modulus of rigidity and ν is

165

Poisson’s ratio. Young’s modulus is defined as the ratio of stress to strain and the shear

modulus is defined as the ratio of shear stress to shear strain. Poisson’s ratio is given as:

ν = −lateral strain
axial strain

= − 
yy

xx

= − 
zz

xx

(7.3)

If the materials were not isotropic, then there would be different values for E, G and ν

representing the differing directional properties. Isotropy also provides a direct relationship

between these properties:

G =

E

2(1 + ν)

(7.4)

The principle stresses and strains are able to be determined by solving for the eigenvalues

of the stress and strain tensors. The determinant of the stress tensor is given as:

σxx − λ1

γxy

γzx

γxy

σyy − λ2

γzx

γyz

γyz

σzz − λ3

(cid:12)(cid:12)(cid:12)(cid:12)(cid:12)(cid:12)(cid:12)(cid:12)(cid:12)(cid:12) = 0

(7.5)

(cid:12)(cid:12)(cid:12)(cid:12)(cid:12)(cid:12)(cid:12)(cid:12)(cid:12)(cid:12)

(cid:114)1

2

where λ1, λ2 and λ3 are the principle stresses. However, the naming convention for the

principle stresses is to set:

σ1 = max(λ1, λ2, λ3)

σ3 = min(λ1, λ2, λ3)

(7.6)

(7.7)

and σ2 is assigned to the remaining lambda value. The equivalent, or von Mises, stress can

then be determined through:

(cid:2)(σ1 − σ2)2 + (σ2 − σ3)2 + (σ3 − σ1)2(cid:3)

(7.8)

σe =

This allows for reducing complex stress state down to a single value which can easily be

compared to the material’s yield strength. The Factor of Safety (FoS) used for judging the

166

model is given as the ratio of the tensile yield strength of the material to the equivalent stress

and is also known as the von Mises, or maximum-distortion-energy, criterion. This criterion

has been shown to work well for ductile materials. For brittle materials, the ultimate strength

would be used instead of the yield strength, but because the ultimate strength is larger than

the yield strength, this would give a larger value for the FoS for the same equivalent stress

[99]. Therefore, the more conservative yield strength was used for the yield criterion.

ANSYS Mechanical solves the global system of equations of the form:

[K]{u} = {F}

(7.9)

where [K] is the global stiffness matrix, {u} is the global solution vector and {F} is the

global applied load vector. In order to handle the nonlinearities that arise from the contact

interfaces in addition to any geometric nonlinearities that could arise due to the chang-

ing geometry from deflection, the program utilizes a Newton-Raphson method to solve the

nonlinear problem of form:

[KT

i ]{∆ui} = {F a} − {F nr
i }
{ui+1} = {ui} − {∆ui}

(7.10)

(7.11)

where [KT
vector of applied loads, {F nr

i ] is the Jacobian (tangent) matrix, {ui} is the displacement vector, {F a} is the
i } is the vector of restoring loads (i.e. the element internal loads),
and the subscript i represents the current equilibrium iteration number. The right hand side
of Equation 7.10 is referred to as the residual or out-of-balance vector {R} that the Newton-

Raphson method is trying to reduce below the convergence criteria which is commonly set

to 0.01 for structural degrees of freedom [100]. This method consists of subdividing the load

into incremental loads and then at each point the solver evaluates the out-of-balance vector,

which is the difference between the restoring forces and the applied loads, and performs a

linear solution for the out-of-balance loads while checking for convergence of forces and mo-

ments. If convergence is not obtained, the out-of-balance vector is reevaluated, the matrices

167

updated, and a new solution is calculated. This process is repeated until convergence is

obtained. These substeps also served to apply the loads gradually, achieving a more realistic

response than if the loads went from 0–100 % instantaneously. The automatic load stepping

feature of the program utilizes a bisection technique to determine an appropriate number of

substeps that achieves good balance between accuracy and computing requirements through

automatic recovery from convergence failure. It reverts to the previous successful step and

halves the step size and solves again. This process is repeated until convergence is achieved

or until a minimum step size (if specified) is reached [101].

It is also important to ensure that the simulation results were converged with respect to

the mesh density. The parameters of interest that were monitored were the maximum shear

stress, maximum principle stress, and the von Mises stress. They were then plotted against

the number of mesh elements shown in Figure 7.4.

Figure 7.4: Grid convergence - Maximum shear stress, maximum principle stress and von
Mises stress vs element count

168

0.50.8751.251.62522.3752.75Number of Elements1050200400600800100012001400160018002000Stress (MPa)Grid Convergence - StressMaximum Shear StressMaximum Principle Stressvon Mises Stress7.3.1.2 Results

FEA analysis was performed at the maximum turbine speed to check for problematic areas.

The results for the FoS can be seen in Figure 7.5. The minimum value of 1.05 occurs at the

edge of the bolt hole and bolt in the same location. This value being so close to 1 indicates,

that in reality, there is a good possibility of failure of the hub adapter or bolt occurring at

40,000 rpm. Reduction of the speed to 35,000 rpm increased the FoS to 1.35 which is a more

reasonable for aerospace design but this could still be an over predicted value due to the

idealized nature of the simulation. Therefore, it was decided to limit the maximum testing

speed for the current configuration to 30,000 rpm which had a FoS of 1.40 to prevent any

chance of the hub adapter causing the assembly to fail.

Figure 7.5: FEA of hub adapter assembly at 40,000 rpm

Figure 7.8 shows the manufactured hub adapter fully assembled with the rotor and

mounted on the turbine shaft.

169

Figure 7.6: FEA of hub adapter assembly at 35,000 rpm

Figure 7.7: FEA of hub adapter assembly at 30,000 rpm

170

(a) Leading edge

(b) Trailing edge

Figure 7.8: IM-v2.1 assembly for spin testing

171

7.4 Testing and Results

The rotor (IM-v2.1) and shaft assembly was first balanced according to Appendix A before

being installed into the air turbine and can be seen hanging below the bulkhead in Figure 7.9.

The bulkhead was then lowered into place and secured. The chamber was then evacuated

before all of the auxiliary systems and the compressor were turned on. The chamber achieved

a steady state vacuum pressure of 736.6 mmHg.

Figure 7.9: Mounted assembly hanging under bulkhead

A Photron FASTCAM SA4 high speed camera was used to monitor the rotor during

testing and for documentation of a failure event for analyzing where failure first began and

how it propagated. The camera position can be seen in Figure 7.1 where it is aimed at a

mirror that allows it to see through the bottom view port and up at the underside of the

rotor assembly. The camera was set to record 6000 fps at a resolution of 512x512 pixels in

a continuous recording loop until triggered by a failure event.

The air turbine was controlled by the TC4 controller and SpinIV software from Barbour

Stockwell Inc. Figure 7.10 shows the setup for controlling, monitoring and recording the air

turbine and high speed camera data. It is possible to program the software to make running

a test achievable by the press of a single button, or it can be run manually. For this test the

turbine was controlled manually, where the valve opening % was gradually opened in small

172

Figure 7.10: SpinIV software and SA4 live feed

increments to let the speed ramp up slowly. It was first opened 5 % until the acceleration rate

slowed then increased to 10 %, again until the acceleration rate began to slow. Figure 7.11

illustrates this trend through plotting of the speed history data.

Increments of 5 % were

Figure 7.11: SpinIV monitor of air turbine speed vs time as the valve opening was adjusted

used until 20 % valve opening where the increment was then reduced for better control. A

173

00:0001:0002:0003:0004:0005:0006:0007:00Time (min)00.270.540.811.081.351.621.892.162.432.7Rotational Speed (rpm)104SPINIV - Turbine Speed vs Timemaximum valve opening of 25 % was reached. At speeds of 25,000 rpm and above, there

started to be some noticeable eccentricity in the rotation and audible noise. Therefore in

order to prevent damaging the air turbine bearings due to the direct drive setup used here,

the test was powered down. A peak speed of 26,168 rpm was achieved. This corresponds to

a tip speed of 340 m/s which is the highest tip speed reached by a wound composite impeller

by 140 m/s. After removing the rotor from the spin pit and performing a visual inspection,

there was no visible damage. These results are encouraging that with an improved test setup

higher tip speeds could be achieved. The vibrations could have been due to the rotational

speed approaching or crossing a resonance frequency of the assembly. Future work could

be done to determine the natural frequencies to ascertain if there are rotational speeds that

should be avoided. Isolating the rotor assembly on a separate shaft and bearing system that

is then joined with the turbine shaft by a sacrificial coupling would allow for testing up to

the 40,000 rpm limit of the air turbine. When a failure event occurs this part would shear

off due to the forces in the system and prevent damage to the air turbine.

174

Chapter 8

Conclusion and Future Work

8.1 Conclusions

The novel wound composite manufacturing process for producing scalable, modular, high

strength and cost effective axial turbomachines could have far reaching implications in a

number of industries. By building on the prior research on developing this technology, the

work presented here was able to make significant strides in adapting and improving upon

the customizable nature of this manufacturing process. It was demonstrated that through

new mandrel design and manufacturing techniques in addition to fiber layup patterning, it

is possible to create more complex 3D impeller geometries than the original “star” pattern

designs.

A quick and simple design approach was devised to allow for easy generation of aerody-

namic designs that could be validated through CFD before being used for creation of the

winding mandrels. 3D printing technology provides a fast and inexpensive method for ex-

perimenting with various mandrel designs and winding patterns. However, the use of cast

wax mandrels allows for a reusable mandrel design to be realized. This has a higher initial

cost to manufacture the casting mold and base. Further work is needed to improve how the

resulting cast geometry matches the design because of issues such as shrinkage and bubble

formation. However, once the technique is better established subsequent uses become very

inexpensive because the cost of wax is very low and can also be melted down and recast.

In the prior work fibers were wet-wound, which proved challenging to handle and intro-

duced a time constraint component to the process. Therefore, through dry-winding of the

fibers and infusion under vacuum, these difficulties were eliminated while facilitating im-

provements to the overall quality and geometric possibilities of the resulting rotors. It was

175

demonstrated that both front (low h/t) and rear (high h/t) stage axial compressor rotors

can easily be produced. The ability to manufacture shaft-driven configurations with this

alternative approach was demonstrated. In addition, integrating a PMSM motor at the hub

region was also shown to be possible through careful design considerations (as opposed to

at the shroud, which had been demonstrated in previous research). Integrating the drive

motors is an essential part of enabling modular design and it also provides an increased level

of control to each stage which can be driven independently, thereby allowing for more flex-

ibility in terms of potential operating range in a multistage machine with counter rotation

or during the start up period.

Preliminary aerodynamic testing of this “next generation” wound composite impeller with

integrated motor was able to reach the highest operational design speed to date at 15,705 rpm.

Despite high vibration levels limiting testing, a total-to-total pressure differential of 200 Pa

was demonstrated corresponding to an approximate pressure ratio of 1.046 at 4300 Pa inlet

total pressure. Preliminary spin pit testing was also performed with a maximum speed of

26,168 rpm reached, which corresponds to a tip speed of 340 m/s, surpassing the previous

maximum by 140 m/s.

8.2 Contributions

ˆ A method of streamlined design for manufacturing was developed that allows for the

quick generation of impeller designs that can be easily converted into mandrel models.

With the focus of this work on refining the manufacturing process the number of

adjustable parameters was reduced through simplifying geometric design assumptions

of constant blade thickness, constant tip radius, constant axial chord and circular arc

blades. Then, with the specification of a small number of preliminary geometric inputs

such as rt,in, rh,in, tb, Nb, AR, p0,in and T0,in the design approach maximizes the

pressure ratio within the confines of limiting parameters like Mach numbers, blade

angles, DF and de Haller number. A simple boundary layer displacement technique

176

was included to more accurately predict the mass flow rate. The ratio of this adjusted

mass flow rate to the inviscid mass flow rate was then used to adjust the resulting

performance parameters to more closely predict the CFD results.

ˆ The analytical design approach was numerically validated through CFD showing closely

matching performance parameters without major regions of separation or the formation

of shocks.

ˆ Compared to the previous cylindrical mandrel designs, this new mandrel manufacturing

approach, which fills the entire fluid path, allows for any 3D shaped axial turbomachine

designs and swirl distributions to be produced.

ˆ The new approach has been tested by producing wound composite impellers using

enhanced mandrel designs manufactured through 3D printing and wax casting. Dry

fiber layup and vacuum infusion technique was explored. Various modifications to the

winding patterns were experimented with to improve the fiber fill and overall quality.

Vacuum infusion, especially in conjunction with the wax mandrel, allowed for a smooth,

comparably high quality surface finish compared to wet wound or 3D printed mandrel

versions. The production of both front (low h/t) and rear (high h/t) stage rotors with

either shaft or integrated motor driving capabilities further demonstrated the flexible

nature of this manufacturing process.

ˆ The first experimental spin pit testing of “next generation” shrouded wound composite

impellers up to a tip speed of 340 m/s demonstrated the structural integrity.

8.3 Future Work

ˆ An investigation in which many differing geometries and flow conditions are compared

can provide additional insight into the use of the simple boundary layer reduction

factor approach for performance estimation.

177

ˆ The further improvement of mandrel designs will be important for continuing to in-

crease the quality of the impellers. The wax mandrel approach needs the casting

process to be further developed in order to produce more accurate geometries that are

free from voids. Other modifications could be incorporated to begin investigating the

production of airfoil shaped blade profiles to improve the aerodynamic potential.

ˆ Investigation of alternative winding patterns or techniques in conjunction with mandrel

modifications to improve the fiber volume fraction. Focus on improving the fiber fill

in areas such as the inner hub flow surface that currently has little to no fiber in

some regions should help realize the full potential of the structural properties that are

achievable with composites.

ˆ Investigate improving impeller quality through sophistication of the infusion process.

RTM/VARTM improvements such as those detailed further in Appendix B would help

reduce the need for post-processing.

ˆ Improve compressor section of aerodynamic testing facility and assembly to allow for

more extensive aerodynamic testing and validation of design specifications. Areas

to focus on improving include the shaft and bearing mounting scheme to improve

alignment and rigidity as well as outer diameter rotor sealing to reduce leakage flow.

ˆ A thorough mechanical characterization of the composite impellers. Proposed meth-

ods for determination of properties like density and void content are introduced in

Appendix B. Experimenting with Non-Destructive Testing (NDT) methods to be able

to deduce information about the quality of an impeller without the need to sacrifice

the part of interest.

ˆ More extensive failure testing in the spin pit through isolation of the composite impeller

shaft from the air turbine shaft to prevent damaging the turbine during a failure event.

Figure 7.9 shows a bolting interface already exists on the underside of the bulkhead for

178

mounting a separate shaft assembly that can then be connected through a sacrificial

coupling.

ˆ Other potential spin pit investigations could include mounting strain gauges directly to

regions of interest or using 3D Digital Image Correlation (DIC) analysis during testing

to provide further insight into structural behavior under high loading and during failure

events.

179

APPENDICES

180

APPENDIX A

Dynamic Balancing

The presence of unbalance in machines with rotating parts is of a major concern for both

safety and cost reasons.

It causes vibration to occur during operation which can lead to

catastrophic or fatigue based failure. It is especially important to address this in high speed

turbomachines for safety and reliability. Unbalance is the result of an uneven distribution

of mass (whether from design or manufacturing deficiencies) which causes the inertial axis

to be out of alignment with the axis of rotation causing the unwanted vibrations.

The ISO has defined a method of classifying this vibration level in terms of balance

quality grade which is expressed as a “G” number. The G number, given in Equation A.1,

is the product of the specific unbalance and the angular velocity of the rotor at maximum

operating speed. This means that a smaller G number has less imbalance present for a given

rotational speed. The application type typically provides an estimate of what G number is

required to keep the vibration level small enough, though it is possible there are cases that

could need a higher or lower G number than is recommended. These threshold levels are

detailed further in the ISO standard 1940-1 [102].

G = eper ∗ Ω = constant

(A.1)

The G number relates the permissible residual specific unbalance, eper, and the rotational

speed, Ω. Where the permissible residual unbalance, Uper, is equal to the specific unbalance

multiplied by the rotor mass, m:

Uper = eper ∗ m

(A.2)

A car tire would have a G40 which means that eper ∗ Ω is 40 mm/s which at a speed of
1000 rpm gives a eper =300 g ∗ mm/kg. Multiplying this value by the mass of the rotor

181

then gives the maximum permissible unbalance that must be achieved during the manufac-

turing/balancing process to ensure safe operation. As the speed/precision of the rotating

system increase, so does the need to reduce the permissible unbalance. Many turboma-

chinery applications use a grade of G6.3 for medium speed and G2.5 for most high speed

applications.

In order to balance the rotor assembly mass needs to be added or subtracted at strategic

locations to counteract the unbalance forces. The balancing of the rotor assembly is to be

performed on a BalanceMaster HB-20-SS dynamic (two-plane) horizontal balancing system

located in the Turbomachinery Lab at MSU. It has a 20 kg weight capacity and can handle

rotor sizes of up to 810 mm in diameter and up to 840 mm in axial length. The balancer

is equipped with a belt drive, two accelerometers, an optical speed sensor and a software

interface. The shaft and rotor setup for installation into the test loop is balanced in the

inboard rotor configuration (rotor assembly is between the bearings) while the setup for the

spin pit is balanced in the outboard or overhung configuration where the rotor assembly is

cantilevered to one side of the bearings. A two plane balancing approach is used to reduce

the dynamic unbalance below the desired threshold.

The EasyBalance 2.2 software was supplied with the balancer and provides a range of

configuration settings and tools used in the balancing process. The appropriate configura-

tion must be selected and the corresponding geometry and tolerance data input into the

setup screen shown in Figure A.1. Before the actual balancing process can take place it is

necessary to associate the physical angular orientation of the rotor with those displayed in

the software. This is accomplished by first recording the rotor assembly as is followed by

activating the electronic zeroing function. With this activated, a test mass is added and then

the balancer is run again. With these results it is possible to determine and account for the

relationship/offset between the display and the actual correction planes. The test weight is

then removed and the rotor assembly unbalance is measured again to determine how much

mass and at what locations it needs to be added (or removal would occur at 180 deg from this

182

Figure A.1: EasyBalance 2.2 Rotor Setup Screen [103]

location). Once the correction masses have been added or removed then the balancer is run

again. If done properly, the residual unbalance in each plane should have been noticeably

Figure A.2: EasyBalance 2.2 polar diagram display of rotor balanced to within G6.3 tolerance

183

reduced. This process is repeated as necessary until the residual unbalance values in both

planes fall under the limits determined by the specified G number and operation speed. This

means that they fall within the green circles on the polar diagram plot (where the radius

corresponds to magnitude of unbalance mass) as seen in Figure A.2.

Originally the balancing approach was to use removable disks (with angular markings

etched onto the surface) that could have mass removed from them at the desired correction

location which were shown in Figure 6.9. In Figure 5.41, post balancing, it is possible to see

where mass has been removed in several locations. This method of taking readings and then

removing the ring and determining how much material from the disk needed to be removed

proved challenging and tedious. Due to this, the balancing scheme was changed to a mass

addition approach which can be seen in Figure 7.9. By using an array of threaded holes and

set screws of varying mass it was possible to balance the spin pit assembly without needing

to permanently altering any parts of the assembly.

184

APPENDIX B

Composite Characterization

In Section 5.7 it was illustrated that there are essentially two different regions of interest on

the impeller where the composite properties could vary significantly. These are the blade

and shroud regions (where the fiber appears to be more tightly packed) and the hub region

(where there appears to be more potential for lower fiber volume fractions). Ideally, these

regions would have a comparable fiber volume fraction but work still needs to be done in

order to achieve this. Due to these differences, a value for the overall density of the composite

might not properly characterize the composite in where ever the region of interest might be.

Therefore, it is proposed to use several methods in conjunction to attempt to more fully

quantify the impeller properties.

The American Society for Testing and Materials (ASTM) standard D792 provides stan-

dardized guidelines for determining the density and relative density of composites through

use of the displacement method. However, this standard was designed for samples weighing

between 1–50 g which is well below the weight of an impeller. Therefore, several potential

approaches are proposed. One approach is to scale up the displacement measure setup to

accommodate the entire rotor. Another way would be to prepare samples from a complete

wheel by cutting it into smaller specimens for analysis. Depending on how the samples are

chosen it would be possible to get specimens that are representative of only the blade/shroud,

only the hub or both regions combined. This would allow for the samples to fall into the

1–50 g range for using the D792 standard to determine the density. The properties of the

matrix are somewhat dependent on the cure cycle so several samples of pure matrix should

also be prepared and tested in order to obtain an accurate measure of it’s density.

Standard D792 is based on the concept of Archimedes’ principle which states that for

any object that is partially or entirely immersed in a fluid, it is buoyed up by a force equal to

185

the weight of the fluid that the object displaces. By having a simple and precise method of

determining mass, it is possible to remove the need to measure volume. The steps necessary

to determine the density of the composite samples through use of this principle, which are

laid forth in standard D792 [104], are summarized as follows:

1. Measure and record the temperature of the distilled water in the immersion vessel.

2. Weigh the sample in air (referred to as a).

3. Attach a piece of fine wire sufficiently long to reach from the hook above the pan to

the support for the immersion vessel (the weight of the wire should also be noted).

4. Place the immersion vessel on the support and completely immerse the suspended

specimen (and sinkers if needed to completely submerge sample).

5. Ensure that there are no bubbles stuck to the specimen (or sinker).

6. Determine the mass of the suspended specimen and sinker/partially immersed wire

(referred to as b).

7. Weigh the sinker and wire immersed to the same depth to determine their apparent

mass (referred to as w).

8. Calculate the relative density, also known as specific gravity using SG = a/(a + w− b).

9. Calculate the density of the composite by multiplying the relative density by the ref-

erence density of the water.

The equipment setup needed to perform these steps is shown in Figure B.1. The immer-

sion vessel is supported above the weighing pan of the balance so that it does not contribute

to the reading. The sample holder is connected to the weighing pan so that the apparent

mass of the submerged specimen, sinker and wire can be measured. Due to the larger size

and mass of the impeller, it is not possible to satisfy the requirements of specimen mass

186

Sample Holder

Immersion Vessel

Immersion Vessel

Support

Weighing Pan

Balance

Figure B.1: Setup for determining density

or balance precision. The same techniques put forth in ASTM standard D792 can still be

employed. However, a larger immersion vessel and sample holder are needed in addition to a

higher capacity (and thus less precise) analytical balance in order to accommodate for this.

It is possible to approximate the fiber volume fraction once the densities of the composite

and its constituents are known. If it is assumed that there are no voids in the composite

then the fiber volume fraction can be estimated using the rule of mixtures. This is based on

the fact that the sum of the fiber volume fraction and matrix volume fraction must be equal

187

to the total volume of the specimen (∀f + ∀m = 1). The composite density is then given as:

ρc = ρf∀f + ρm∀m

(B.1)

By using the relationship between ∀f and ∀m to substitute for ∀m in Equation B.1 it is
possible to solve for the fiber volume fraction.

∀f =

ρc − ρm
ρf − ρm

(B.2)

The previous methods for determining the density of the composite or constituent mate-

rials all ignore the fact that there can be voids inside that can skew the results. The presence

of voids, like areas of low fiber content, can significantly affect the mechanical properties.

Therefore, it is also important to gain an understanding of the void content that results from

the current state of the novel manufacturing process. If there appears to be a consistently

high void content then some focus on improving the infusion process to reduce this will be

necessary to ensure the mechanical characteristics are more consistent.

The ASTM standard D2734 provides an approach, in conjunction with standard D3171,

to estimate the void content. First, the densities of the resin and the composite are deter-

mined individually. Since the density of the reinforcement provided by the manufacturer

is typically accurate it is not necessary to measure this separately. Then the resin content

of the composite is determined using the thermogravimetric analysis technique specified in

D3171. A sample of pure matrix is first carbonized through pyrolysis (essentially burning

without oxygen) to determine if any residue remains. The process is repeated on a composite

specimen. By comparing the initial and final masses it is possible to determine quantities

such as the carbonization ratio (of the pure matrix), the mass of the matrix, the reinforce-

ment content, the matrix content, and the void volume [105]. With the matrix content and

density known as well as the reinforcement content and density, it is then possible to deter-

mine a theoretical composite density. The void content can then be determined from the

188

theoretical and measured composite densities [106].

A main goal for the future of the manufacturing process should be to minimize the

need for post processing as much as possible. Machining composites is known to reduce

the mechanical performance through a number of potential mechanisms. These include

fiber pullout and fiber-matrix debonding which can lead to accelerated failure [107]. This

means that it could both save time and increase the mechanical strength and integrity of

the resulting impellers. A proposed solution is shown in Figure B.2. The inside of the

infusion chamber would first be coated in mold release. Then a mandrel with fiber already

wound around it would be placed inside the chamber and then the lid reinstalled. The resin

could then be infused using vacuum or positive pressure allowing experimentation with what

method produces the best results in this application. The underside of the lid could be

designed to mesh with the tops of mandrel fluid paths to help improve the quality of the

trailing edge side of the impeller.

Figure B.2: Proposed infusion chamber design

189

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