KT Thesis July2009 Final Print

8/6/2019 KT Thesis July2009 Final Print

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Contents

1 Introduction 12

1.1 Background . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12

1.2 Objectives and scope . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 16

1.3 Thesis outline . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 17

2 Literature review 19

2.1 Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 19

2.2 Compressor characteristic curve . . . . . . . . . . . . . . . . . . . . . . . 19

2.3 Compressor instability - surge and stall characteristic . . . . . . . . . . . . 23

2.4 Surge control techniques . . . . . . . . . . . . . . . . . . . . . . . . . . . 28

2.4.1 Surge avoidance . . . . . . . . . . . . . . . . . . . . . . . . . . . 28

2.4.2 Surge suppression . . . . . . . . . . . . . . . . . . . . . . . . . . 29

2.4.3 Active surge control . . . . . . . . . . . . . . . . . . . . . . . . . 30

2.5 Compressor surge test rig overview . . . . . . . . . . . . . . . . . . . . . . 33

2.6 Summary of literature review . . . . . . . . . . . . . . . . . . . . . . . . . 35

3 Engineering analysis and experimental setup 38

3.1 Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 38

3.2 Test Rig Overview . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 39

3.3 Motor bearings improvement . . . . . . . . . . . . . . . . . . . . . . . . . 41

3.4 Piping system design . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 48

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CONTENTS 2

3.4.1 Piping supports modal analysis . . . . . . . . . . . . . . . . . . . . 50

3.5 Compressor rotor modification . . . . . . . . . . . . . . . . . . . . . . . . 59

3.5.1 Rotordynamics analysis . . . . . . . . . . . . . . . . . . . . . . . 59

3.5.2 Unbalance force analysis . . . . . . . . . . . . . . . . . . . . . . . 65

3.6 Chiller for high speed motor . . . . . . . . . . . . . . . . . . . . . . . . . 70

3.7 Precision Alignment . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 72

3.8 Assembled compressor test rig . . . . . . . . . . . . . . . . . . . . . . . . 73

4 Testing and experimental results 76

4.1 Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 76

4.2 Commissioning . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 77

4.2.1 Motor solo run . . . . . . . . . . . . . . . . . . . . . . . . . . . . 77

4.2.2 Maximum speed testing and mechanical assessment of compressor . 77

4.3 Characteristic curve and instability identification . . . . . . . . . . . . . . 80

4.3.1 Characteristic curve . . . . . . . . . . . . . . . . . . . . . . . . . 80

4.3.2 Uncertainty analysis . . . . . . . . . . . . . . . . . . . . . . . . . 84

4.3.3 Analysis of surge observations . . . . . . . . . . . . . . . . . . . . 86

4.3.4 Discussions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 92

4.4 Effect of impeller tip clearance . . . . . . . . . . . . . . . . . . . . . . . . 95

5 Conclusions 100

5.1 Conclusions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 100

5.2 Implications . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 103

5.3 Recommendations for future work . . . . . . . . . . . . . . . . . . . . . . 104

5.3.1 Develop surge controller . . . . . . . . . . . . . . . . . . . . . . . 104

5.3.2 Speed tachometer . . . . . . . . . . . . . . . . . . . . . . . . . . . 105

5.3.3 Accelerometers . . . . . . . . . . . . . . . . . . . . . . . . . . . . 105

5.3.4 Acoustic measurement for surge detection . . . . . . . . . . . . . . 106


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CONTENTS 3

5.3.5 Additional pressure tappings in scroll . . . . . . . . . . . . . . . . 106

5.3.6 Vaned diffusers . . . . . . . . . . . . . . . . . . . . . . . . . . . . 106

5.3.7 Bearing force measurement . . . . . . . . . . . . . . . . . . . . . 107

5.3.8 New variable speed drive . . . . . . . . . . . . . . . . . . . . . . . 107

5.3.9 Hole pattern seals . . . . . . . . . . . . . . . . . . . . . . . . . . 108

5.3.10 Spare parts . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 108

A Precision alignment procedure 115

A.1 Prealignment . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 115

A.2 Alignment . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 115

B Procedure to check impeller tip clearance 121

C Chiller - Model HCV 1500 PR 123

D Inlet Air Filter Details 124

E Orifice Flow Meter Details 126

E.1 Flow meter selection . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 126

F Procedure for testing surge on compressor 131


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List of Figures

1.1 Examples of barrel and horizontal split multistage centrifugal compressor

(Siemens) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13

1.2 Examples of overhung and integrally geared compressors (Siemens) . . . . 13

1.3 Compressor characteristic curve . . . . . . . . . . . . . . . . . . . . . . . 15

2.1 (A) Velocity vector triangle at high flow; (B) Velocity vector triangle at low

flow . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 21

2.2 Illustration of gas particle path across diffuser [1] . . . . . . . . . . . . . . 24

2.3 Photos to illustrate damage on impeller from compressor surge [1] . . . . . 26

2.4 Illustration of inlet guide vanes to suppress surge [24] . . . . . . . . . . . . 31

2.5 Illustration of air injection technique for compressor stabilization [23] . . . 32

2.6 Illustration of experimental setup for Spakovszky test rig [29] . . . . . . . . 33

2.7 Simulated surge occurrence without any surge controller [3] . . . . . . . . 36

2.8 Simulated surge control result with impeller tip clearance actuation [3] . . . 36

3.1 Compressor test rig cross sectional drawing . . . . . . . . . . . . . . . . . 40

3.2 Exploded view of compressor test rig . . . . . . . . . . . . . . . . . . . . . 41

3.3 Comparison between SKF 7005 CD bearing versus SKF hybrid bearing . . 46

3.4 Piping layout to show the three discharge throttle valve positions . . . . . . 50

3.5 Frequencies of several types of aerodynamics flow instabilities by Willems

[40] . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 51

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LIST OF FIGURES 5

3.6 Schematic of the equivalent model to represent compressor system . . . . . 53

3.7 Flow area to be considered when calculating the Helmholtz frequency . . . 53

3.8 Inlet piping support and discharge piping supports . . . . . . . . . . . . . . 55

3.9 Modal analysis results for inlet pipe support - 1st and 2nd modes . . . . . . 57

3.10 Modal analysis results for discharge piping support - 1st and 2nd modes . . 58

3.11 Locations of 4 lead pieces on the impeller face . . . . . . . . . . . . . . . . 61

3.12 Location of spacer ring behind the impeller . . . . . . . . . . . . . . . . . 62

3.13 Courtesy of Kobe Steel - Finite Element model showing deflection of impeller 64

3.14 Rotstab rotor model with spacer ring added behind impeller . . . . . . . . . 65

3.15 Free-free mode shape of rotor with space ring included . . . . . . . . . . . 66

3.16 Schematic to illustrate force analysis on spacer ring of compressor . . . . . 68

3.17 Show an offset spacer ring could contribute to unbalance force . . . . . . . 69

3.18 Frictional force analysis - analogous analysis to clamped spacer ring . . . . 69

3.19 Air cool chiller skid for the high speed motor . . . . . . . . . . . . . . . . 72

3.20 Inlet filter and exhaust for compressor test rig . . . . . . . . . . . . . . . . 74

3.21 Assembled compressor test rig for surge testing . . . . . . . . . . . . . . . 74

3.22 Assembled compressor test rig for surge testing . . . . . . . . . . . . . . . 75

4.1 Motor amperage versus compressor speed curve . . . . . . . . . . . . . . . 78

4.2 Overload protection card in variable frequency drive - damaged red diode . 79

4.3 Orbital plots for impeller end and motor end of compressor rotor at 10000rpm

at full flow(Throttle valve is 100% open) . . . . . . . . . . . . . . . . . . . 81


at minimum flow (Throttle valve is 31% open) . . . . . . . . . . . . . . . . 81

4.5 Orbital plots for impeller end and motor end of Compressor Rotor at 17000rpm

at full flow(Throttle valve is 100% open) . . . . . . . . . . . . . . . . . . . 82


at minimum flow (Throttle valve is 31% open) . . . . . . . . . . . . . . . . 82


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LIST OF FIGURES 6

4.7 Characteristic curve at different operating speed (Impeller tip clearance at

23 mils) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 84

4.8 Plot at 16292 rpm to characteristic curve with error bar . . . . . . . . . . . 86

4.9 Pressure profile fluctuations at compressor discharge plenum, compressor

casing and compressor inlet at 16000 rpm . . . . . . . . . . . . . . . . . . 88

4.10 Magnified pressure profile plots at discharge plenum, casing and inlet at

the initiations of instability at 16000 rpm - showing 21 pressure peaks . . . 89

4.11 Magnified pressure profile plots at discharge plenum, casing, inlet at insta-

bility at 16000 rpm - showing 7 pressure peaks . . . . . . . . . . . . . . . 90

4.12 Waterfall plot to show dominant frequency change as throttle valve is closed

- 10000 rpm . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 92


- 15000 rpm . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 93


- 16000 rpm . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 93

4.15 Sanadgols simulated result to show influence on characteristic curve with

impeller clearance adjustment [3] . . . . . . . . . . . . . . . . . . . . . . . 97

4.16 Characteristic curve at 14900 rpm when thrust disk is statically moved in

the axial direction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 98

4.17 Characteristic curve at 16287 rpm when thrust disk is statically moved in

the axial direction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 99

A.1 Final soft foot readings on motor and amount of shims used . . . . . . . . . 119

A.2 Setup for face and rim alignment on between compressor and the motor . . 120

A.3 Alignment fixtures for fine adjustment of the motor . . . . . . . . . . . . . 120

E.1 Illustration of a typical setup for an orifice flow meter [35] . . . . . . . . . 127


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LIST OF FIGURES 7

E.2 An illustration of the orifice flow meter by Lamdasquare with flow straight-

eners upstream . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 129


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List of Tables

3.1 ndm - Speed limit calculation for compressor test rig at possible running

speeds . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 43

3.2 Material construction for ball bearings [47] . . . . . . . . . . . . . . . . . 45

3.3 Deep groove versus angular ball bearings design properties for 25 mm inner

diameter bearings . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 47

3.4 Descriptions for hybrid bearing 6005-2RSLTN9/HCFC3WT . . . . . . . . 48

3.5 Helmholtz frequency estimated comparison table . . . . . . . . . . . . . . 53

3.6 Comparison of pipe supports predicted frequencies versus Helmholtz fre-

quencies at each valve location . . . . . . . . . . . . . . . . . . . . . . . . 56

4.1 Results of motor solo run up to maximum speed of 10000 rpm . . . . . . . 77

4.2 Comparison of amount of axial position fluctuations at 16000 rpm . . . . . 91

A.1 Table of key items required to perform alignment between motor and com-

pressor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 116

A.2 Final alignment readings between motor and compressor . . . . . . . . . . 119

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Nomenclature

H Flow angle at high flow

L Flow angle at low flow

Flow Angle

Coefficient of friction

H Helmholtz resonator frequency [Rad/s]

Shaft rotational speed [rad/s]

Diameter of coupling hub [inch]

Ac Flow area at the eye of the impeller [m2]

A Axial distance between motor front and back leg

a Speed of Sound [m/s]

bf Bottom face alignment reading [mils]

bf Bottom rim alignment reading [mils]

BL Motor back leg alignment adjustment [mils]

C Damping Matrix [lbs/inch]

c Speed of sound in air under ideal conditions

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LIST OF TABLES 11

V2r Absolute velocity radial component

V2t Absolute velocity tangential component

V2 Absolute velocity vector of gas particle

Vd Velocity downstream of Orifice

Vp Plenum Volume [m3]

vs Maximum discharge flow velocity [m/s]

Vu Velocity upstream of orifice flow meter

W2 Impeller blade tip exit velocity vector

x Displacement vector


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Chapter 1

Introduction

1.1 Background

Centrifugal compressors are used extensively in many industries such as oil refining

industries, chemical industries and the upstream oil and gas industries. They are particu-

larly well suited for processes that require a very wide performance range as they can be

designed with different mechanical configurations for the specific process needs [1].They

are commonly used for pressurizing different types of gases and moving them in flow vol-

ume from different physical locations or from different containments. These compressors

are dynamic machinery and work on the principle of using motion to transfer energy from

the compressor rotor to the process gas. Compression of the gas is achieved by means of

blades on a rotating impeller for a single stage machine, or a set of impeller blades for a

multi stage machine. This rotary motion of the gas results in an outward velocity due to the

centrifugal forces. The tangential component of this outward velocity is then transformed

to pressure by means of a diffuser. Centrifugal compressors come in many different con-

figurations. Figure 1.1 and Figure 1.2 show four typical types of centrifugal compressors

commonly found in the industry today. Barrel design multi-stage compressors are used in

high pressure services. Horizontally split multi-stage compressors are typically applied in

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CHAPTER 1. INTRODUCTION 13

Figure 1.1: Examples of barrel and horizontal split multistage centrifugal compressor

(Siemens)

Figure 1.2: Examples of overhung and integrally geared compressors (Siemens)

high volume medium pressure applications [1]. Its horizontally split design provides good

accessibility and on site maintenance to the internal components of the compressor. Over-

hung centrifugal compressors are generally applied in low pressure high volume services.

Their economical cost is an attraction for a wide variety of general purpose air compression

applications. Integrally geared centrifugal compressors are commonly applied in industrial

gas industry or instrument air services where its small foot print gives advantage to use a

low speed coupling and short rigid rotor design to provide the advantage for such applica-

tions [2].


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The unique capability to apply centrifugal compressors in a broad range of services

had rendered them as the work horse in many of such industries. Hence it is important to

learn how to control these machines more effectively and in a pursuit to make them more

efficient, improve or optimize their working envelop. The attractiveness of centrifugal

compressors over other compressors such as screw compressors, reciprocating compres-

sors or axial compressors is mainly because they have fewer rubbing parts, are relatively

energy efficient, and give higher airflow than a similarly sized reciprocating compressor

(i.e, positive-displacement). The performance of a centrifugal compressor is identified by

its characteristic curve, which relates three main parameters of the compressor - mainly the

flow rate, the differential pressure head produced and the speed of the compressor. It is

important to note that changing any one of these three parameters will affect the other two

parameters, and hence the operating point on the compressor curve. The operating point of

the machine is determined by the intersection between the system resistance curve super-

imposed onto the compressor characteristic curve. The point of intersection is the current

operating point of a centrifugal compressor. The stable operating region on the compressor

curve is bounded by the surge line and the choke line. The surge line basically separates the

region of stable operation from the region of unstable operation. Figure 1.3 illustrates how

a typical compressor characteristic curve would look like. Surge is a system phenomena

in a centrifugal compressor and the surge point is defined as the peak head on the com-

pressor characteristic curve at the particular operating speed. It is the point at which the

compressor cannot add enough energy to overcome the system resistance [3]. This causes

a rapid flow reversal. As a result, high vibration, temperature increases, and rapid changes

in axial thrust can occur. These occurrences can damage the rotor seals, rotor bearings,

the compressor driver and cycle operation. Most turbo machines are designed to withstand

occasional surging. However, if the turbo machine is forced to surge repeatedly for a long

period of time or if the turbo machine does not have an adequately designed surge control

system, repeated surges could result in a catastrophic failure of the machine.


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Figure 1.3: Compressor characteristic curve

To prevent surge from occurring in centrifugal compressors, it is widely practiced in the

industry to limit the full operating envelop of a compressor by placing a safety margin line

of typically 10% away from the surge point of the compressor [9]. This limit line basically

completely prevents surge from happening, but at the expense at lower efficiency of the

centrifugal compressor and a smaller operating window. In todays world of centrifugal

compressors, as demand for machine efficiency and flexibility keeps increasing, maintain-

ing this practice of a 10% safety limit line may not be the best option. This is because such

margin would limits the operating range of the centrifugal compressors and can prevent it

from operating at maximum efficiency, which may lie at or close to the safety margin line

[22]. Researchers, users and manufacturers [6, 29, 25, 12, 23]are seeking more innovative

ways to stabilize the compressor or limit surge without significantly compromising com-

pressor performance and efficiency. Numerous research works [42, 16, 15, 12, 48]have

been attempted over the years to find an economically and practically feasible solution to

better control surge. These approaches are mainly classified into [20, 22]surge avoidance,

surge suppression and active surge control. Of the three approaches, active surge control

offers the most promising approach to really stabilize the compressor and hence allow it to


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operate at a wider operating envelop.

Sanadgol [3] developed a new method for active surge control for centrifugal compres-

sors with unshrouded impellers that would use a magnetic thrust bearing to modulate the

impeller tip clearance. Sanadgols simulation results showed that if the position of the com-

pressor shaft could be actuated with sufficient authority and speed, the induced pressure

modulation makes control of surge promising. The active control of the magnetic bearing

would allow realistic real time static and dynamic positioning of the rotor with precision.

This makes it possible to modulate the impeller tip clearance even when the centrifugal

compressor was in operation. Sanadgols theory and results had to be validated with exper-

imental data. A centrifugal compressor test rig that would be fully supported on radial and

axial magnetic bearings had to be built [4]. The fully operational test rig would serve as the

hardware required to perform experimental work and provide actual test data to validate

Sanadgols theory. The compressor test rig, the magnetic bearings and the control philoso-

phy had been developed over the years by Sanadgol, Nathan and Buskirk [3, 4, 5]. Various

major components of the test rig had been sponsored by industrial partners of Rotating

Machinery and Controls Laboratory.

The work for this thesis focuses on the improvement of mechanical components, and

assembling and commissioning the test rig. The test rig would also be tested into its insta-

bility region to demonstrate the effect of static axial modulation of impeller tip clearance.

1.2 Objectives and scope

The objectives and scope of study for this thesis is:

Perform engineering analysis and improvement on the mechanical components of thetest facility and its supporting systems to ensure that the test facility could be assem-

bled, commissioned and tested safely and reliably into high speed for experimental

work to be carried out.


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Experimentally determine the safe operating envelop and identify the potential limi-tations on the commissioned test facility. Characterize the performance of the com-

pressor at various speed of operations and developed the compressor characteristic

map base on experimental collected data.

Experimentally evaluate the performance of the compressor and characterize the ob-servations when it is operated into its instability region. Examine ways or parameters

that had to be observed so as to aid in the future development and implementation of

a surge controller.

Experimentally demonstrate and investigate that the axial movement of the impeller

tip clearance would influence the performance of the compressor and its surge points.

1.3 Thesis outline

The thesis is divided into 5 main chapters, followed by the Appendix section. Chapter

1 provided an overview on the background of the present study and centrifugal compressor

applications. It also outlined the reduction in compressor efficiency and operating envelop

due to the current industry practice that involves avoiding surge completely by maintaining

a safety margin. Chapter 2 is a literature review and gives an insight into the mechanism

that results in stall and surge instability. It further outlines the typical characteristics ex-

pected when these instabilities are encountered and also provides overview of surge related

research work. An overview of previous work on the compressor surge test rig of the

University of Virginia is summarized. Chapter 3 provides engineering details for all the

mechanical enhancement and upgrades that had to be performed on the existing test rig as

part of this study, such that it would ensure that the compressor test rig could be commis-

sioned and tested safely. Component run test results and also maximum speed test results

are provided. Chapter 4 focuses on the experimental test results and includes how surge is

identified on the test rig and also the influence of static impeller tip clearance adjustment


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on the instability points on a characteristic curve. Discussions of the experimental data

was collected and compared. Chapter 5 contains conclusions for this thesis and is further

broken into 3 parts; summary of conclusions, implications and recommendations. This

chapter not only aims to provide the conclusion drawn from the experimental results, but it

also provide recommendations that forms the fundamentals for future research work. The

Appendix section presents all the test procedures developed so as to ensure smooth tran-

sition of mechanical knowledge to future researchers working on the test rig. Mechanical

specifications of the main accessory components in the test rig are also included for future

reference.


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Chapter 2

Literature review

2.1 Introduction

This literature review aims to provide the background regarding centrifugal compressor

characteristics, the different types of instability and an overview about research work on

surge control. This background knowledge serves as reference when conducting experi-

mental testing on the compressor and also helps in drawing conclusions from the experi-

mental observations. Velocity triangles at the impeller is used to describe the mechanism

behind the drooping shape of the characteristic curve for the compressor. Surge and stall

phenomenon are explained and its differences highlighted as these two phenomenon are

commonly mixed up. Differences between deep surge and mild surge is also compared and

summarized. Finally an highlight of surge research work performed by other researchers is

summarized.

2.2 Compressor characteristic curve

The performance of a centrifugal compressor is best illustrated by what is called the

compressor characteristic curve, or commonly known as the compressor performance map.

It is most usual to see these maps plotted with compressor pressure ratios on the vertical

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CHAPTER 2. LITERATURE REVIEW 20

axis versus the inlet volumetric flow on the horizontal axis. These performance maps can

be used to explain many aspects of the performance of the compressor [14]. A basic un-

derstanding of how the slope of these curve is generated is a good aid for visualizing the

physics behind compressor performance and especially surge. Figure 1.3 shows a typi-

cal centrifugal compressor performance map operating at some fixed speed. Two extreme

points on the compressor curve, the high flow point and the low flow point, would be ex-

plained in details using velocity triangles to understand how the compressor pressure ratio

would increase as flow is reduced. Figure 2.1 shows the two velocity triangles at the exit

of the impeller at low flow and high flow.

To understand the slope of the centrifugal compressor characteristic curve, it is neces-

sary to first understand the characteristics of the flow process at the impeller blade exit in

terms of the velocity vector triangle shown in Figure 2.1(A) and Figure 2.1 (B) [1]. Refer-

ring to the velocity triangles for the high flow, W2 is the gas velocity relative to the blade exit

angle, U2 is the tip speed of the blade. The designation 2 is used in all this nomenclature

to indicate the velocity triangles are done at the exit of the impeller, which is commonly

depicted as subscript 2. The addition of these 2 velocity vectors gives the absolute velocity

of the gas particle,V2.

U2+W2 =V2 (2.1)

Having obtained V2, it can be then resolved into its own radial and tangential components,

V2r and V2t respectively as shown in the velocity triangles. The head of a centrifugal com-

pressor can be approximated as the product of the U2 and V2t as stated by the aerodynamics

head equation [38]:

HU2.V2t (2.2)

For a constant speed compressor, the vector is constant since it is derived from the ac-

tual speed of the impeller tip and the impeller rotates at constant speed. Therefore, the


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Figure 2.1: (A) Velocity vector triangle at high flow; (B) Velocity vector triangle at low

flow


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aerodynamic head of the centrifugal compressor is proportional to its absolute velocity tan-

gential component Vt. When the compressor operates in the high flow as system resistance

decreases, its vector W2 will be large in magnitude and this results in a small magnitude

Vt. A small magnitude will result in a small head to be produced by the compressor. This

would explain why a compressor would develop small heads during high flow operating

conditions. In comparison, if one is to analyze the velocity vector triangles when the com-

pressor operates in a low flow conditions (which could be obtained by closing a throttle

valve downstream of the compressor and increasing the system resistance of the compres-

sion system), it can be noticed that the operating point on the compressor characteristic

curve would move up the compressor curve and end up operating very close to the surge

limit. This low flow operating condition results in a smaller magnitude vector W

2 to be

produced. This would produce a larger magnitude vector ofVt as seen in Figure 2.1 and

results in a higher head produced by the compressor during a low flow condition. Hence

understanding how the velocity vectors change in a centrifugal compressor during different

flow applications allows one to understand why the compressor slope could be a negative

sloped curve on a compressor performance map.

On the compressor performance map, the useful operating region of a particular cen-

trifugal compressor is typically bounded by the compressor surge limit and choke lines.

The surge limit line is shown in Figure 1.3 . The surge limit line separates the regions of

stable and unstable compressor operation [25]. The stable region of operation is on the

right of the surge limit line while the region on the left is known as the unstable operation

region or the surge region. In modern day compressors, a surge margin is typically imposed

on the compressor operating region. The surge control line is typically placed on the neg-

ative slope of the compressor performance curve at a location of 10% [38]away from the

surge limit line on the compressor performance map. Providing this safety margin allows

the compressor to operate continuously and avoid operating near the surge region of the

compressor.


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Having understood how the velocity triangles affects the shape of the compressor curve,

the surge phenomena which typically occurs at low flow and peak head conditions, can be

further explained using velocity triangles at the impeller exit. As flow is reduced at constant

speed as shown in Figure 2.1, the system resistance would decrease in the compression sys-

tem and the magnitude of the velocity vector W2 would decrease proportionally. This would

result in a corresponding reduction of the flow angle at the impeller exit and an increase

in the incident angle i at inlet of the compressor blades[14]. The reduction of the flow angle

at the exit results in a longer spiral path of the gas particles across the diffusion section of

the compressor. And when gas particle flow path on the diffuser becomes long enough and

the flow angle i small enough, as compressor flow is reduced, the flow momentum of the

gas particle would reach a point where it is totally dissipated by the diffuser walls due to

friction [38]. This results in a situation whereby the frictional force within the compressor

begins to increase faster than the head that could be produced from the compressor as the

flow is reduced. The stable forward flow of the compressor reverses at this point and results

in surge [17]. Figure 2.2 illustrates and compares the difference in gas particle paths across

the diffuser as the flow is reduced. In Figure 2.2 L and H stand for the flow angle at low

flow and high flow respectively.

2.3 Compressor instability - surge and stall characteristic

The two types of instability commonly encountered in compressors are stall and surge.

Stall is a known as a local instability within the compressor itself and it could occur in both

a centrifugal or axial flow compressor. It is further classified into stationary stall and rotat-

ing stall phenomena [46]. Stationary stall typically occurs in the stationary components of

the compressor (e.g, the diffuser vanes) and is caused by a flow separation occurrence. Its

characteristic is that it typically results in a vibration frequency that is higher than running

speed frequency. Rotating stall, on the other hand, is a non-uniform circumferential pres-


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Figure 2.2: Illustration of gas particle path across diffuser [1]


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sure field which rotates at a different speed than the compressor operating speed, causing

unbalance in the rotor. It can occur in the impeller, IGV and impeller recess and its vibra-

tion frequency is typically subsynchronous. The most common stall to occur in a vaneless

centrifugal compressor (such as this test rig) would typically be rotating stall and the sub-

synchronous frequency would change with running speed change. Rotating stall is a special

class of stall phenomena of stall [46] and is characterized by circumferential non-uniform

mass deficits that propagate around the compressor annulus at a fraction of the compressor

operating speed. And once the stall cells are formed, it generally require to operate the

compressor away from the instability point by opening up the throttle valve to flush out the

stall cells. Stall is a commonly encountered in axial flow compressors.

Surge, on the other hand, is known as a phenomenon of the entire compression sys-

tem involving the compressor itself and its discharge piping or discharge plenums [19, 16].

The key difference between rotating stall and surge are that the average flow in pure rotat-

ing stall is steady over time, but the flow has a circumferential nonuniform mass deficit,

while in pure surge the flow is unsteady but circumferentially uniform [20]. Surge is to be

avoided in centrifugal compressors due to the potential of catastrophic damage associated

with its occurrence, especially in process manufacturing plants, where for example, the

main process gas compressors could be a large turbo machinery ranging between 20,000

- 50,000 hp. Any catastrophic damage to these large plant critical turbo machinery from

surge would result in extended downtime that would result in significant monetary losses

and also possibly compromise the safety of the plant personnel. Figure 2.3 shows an exam-

ple of an impeller rub that occurred due to surge. Surge cycles are typically known to be

short but could occur several times within a second [21]. This flow reversal during surge

not only results in sudden reverse bending loads on nearly all compressor mechanical com-

ponents [1], but the gas temperature will also increase significantly due to the continuous

energy added to the recycled gas from compression. This can lead to serious temperature

overloads in the machine and result in temperature related failures of machine components


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Figure 2.3: Photos to illustrate damage on impeller from compressor surge [1]

as well. Hence the combination of high sudden load reversal and high gas temperatures

puts surge as the unstable operating region to be avoided at all costs during the operation

of centrifugal compressors.

Surge may be broadly classified into two distinct categories - mild surge and deep surge

[30]. The occurrence of each of these two categories is dependent on the compression sys-

tem and the operating conditions of the compressor. Mild surge is closely associated with

the Helmholtz frequency, i.e. the resonance of the compressor duct and the volume con-

nected to the compressor [20]. Mild surge identification may be performed by identifying

the Helmholtz frequency derived from the compression system model where the compres-

sor and throttle valve are both modelled as throttle disk. The Helmholtz frequency remains

constant even when the compressor rotational speed is changed as it is considered a reso-

nance frequency and it generally occurs near or at the start of compressor instability limit.

The pressure profile observed in mild surge is typically linear and its pressure profile is

sinusoidal. Mild surge has been recognized as a low frequency trigger but more damaging

deep surge inception [30].

Deep surge, on the other hand, is known to be a non-sinusoidal or non-linear behavior

[28, 43], as outlined by many researchers who have tested compressors into deep surge.

The deep surge cycle typically depends on speed and the amount of throttling of the con-


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trol valve. Its occurrence is also usually accompanied by a deep audible sound and the

compressor rotational speed is often observed to fluctuate or even drop significantly [44].

The frequency of deep surge is governed by the filling and emptying of the discharge vol-

ume plenum connected to the compressor and this frequency would vary with changes in

the piping configurations of the compressor system (e.g, different piping orientation and

throttle valve location) [6]. Frequency of deep surge is known to be well below Helmholtz

frequency and it usually exhibits large pressure amplitude fluctuations and lower oscillation

frequency as compared to mild surge. Deep surge generally only occurs in system config-

urations that have significant amounts of energy (e.g, pressure) stored downstream of the

compressor. In an industrial setup, this would normally occur when there is significant

discharge pipework connected to the compressor which acts like a large plenum reservoir

[45]. Flow reversal is very possible during deep surge condition and this reversal would

bring about a change in bending load and thrust load on the compressors rotor system.

Usually in experimental work to investigate instability or surge in centrifugal compressors,

investigations and data recordings are carried out during surge inception, detecting mild

surge. With mild surge detected in the compression system, the observations of deep surge

is typically limited to a few consecutive cycle to prevent any damage to test compressors

or on site compressors [30].

Between surge and stall instability phenomena, the occurrence of stall instability in

centrifugal compressor is still a subject of discussion in the compressor community [3, 20].

It appears to be more readily observed in low pressure centrifugal compressor systems or

high pressure centrifugal compressor systems operating at partial speed of the compressor

design . For high speed operational compressors, the compressor generally goes directly

from stable operation to flow separation on all blades and immediately into flow reversal

when the flow reduces beyond the surge limit line[1]. Rotating stall appears to have little

effect on the pressure rise in centrifugal compressors and hence it will not have significant

contributions to surge [20].


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compressor discharge. In the event that the surge control line is reached during operation

for this process compressor, its recycle valve will be regulated and opened to allow more

discharge flow back to the suction of the compressor. The valve will open until the operat-

ing point of the compressor moves to the right of the surge control line again. Both these

methods are capable of avoiding surge completely, but they result in loss of efficiency of

the compressors from the venting or the recycling. The useful stable operating region of

the compressor is also greatly reduced as a result of introducing the surge control line and

defining a typical safe margin of 10%.

2.4.2 Surge suppression

In recent years, there have been increasing innovative attempts to suppress surge be-

yond the surge limits [20, 22, 6]. These studies are mainly aimed at increasing the stable

flow range of the centrifugal compressors. These research works can be further sub cate-

gorized into 2 approaches. The first approach focuses on attempts to improve the interior

of the compressor and is known as the design method, while the second approach aims to

include additional external devices on the compressor to expand the useful operating range

of the compressor. These are typically known as the operational methods to achieve surge

suppression [22].

Rodgers [24] investigated on the effect on operating range extension for a centrifugal

compressor by installing variable angle inlet guide vanes to the suction of the centrifugal

compressor. He operated the inlet guide vanes at varying angles, called preswirl angles

on a high speed overhung design single stage open impeller compressor designed to run at

64643 rpm. He was able to show that by allowing the inlet gas of the suction side of the

impeller to be preswirled, the surge line on the compressor map would move significantly

to the left, as compared to the original surge line that did not have the assistance of the inlet

guide vanes. Figure 2.4 shows the results plot and also an an overview of the test rig. The

test results showed that the surge margin could be considerably extended by the regulations


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of the inlet guide vanes due to increased stability on the impeller.

Gary [23] investigated how air injection into the compressors diffuser region would

help in stabilizing the flow range of the centrifugal compressor. The experimental investi-

gation was based on three-dimensional time accurate simulations of high speed compressor

impeller at surge condition. The simulation studies showed that flow reversal occurred on

the leading edge of the impeller blades at reduced mass flow conditions, such as near the

surge point. The study also showed that flow separation was the cause of the flow rever-

sal and that introducing additional air injection into the compressor diffuser region could

eliminate the local separation on the impeller blades. The air injection would be able to

suppress and prevent the flow separation from occurring and its introduction would be able

to improve the stability of the impeller. He designed a system of injectors that could provide

variations in position, direction and flow rate to affect as many regions of the diffuser as

possible in a high speed compressor with a design speed of 21789 rpm. The injected flows

directions could be either forward tangent injection or reverse tangent injections. Figure

2.5 shows a cross sectional schematic of where the air is injected into the compressor and

also the results from his testing. As can be observed from the results in Figure 2.5, injection

flows that were recycled from the compressor would help in stabilizing surge and therefore

increasing the surge margin of the compressor, as compared to the baseline data where no

flow injections was introduced.

2.4.3 Active surge control

Active surge control is fundamentally different from surge avoidance or surge sup-

pression. In an active surge control setup or scheme, the open loop unstable region of the

compressor map is sought to be stabilized through the use of feedback rather than avoided

by placing a surge margin line [48]. Spakovszky [29]developed and designed a high speed

axial compressor that could be used to investigate the effectiveness of using magnetic bear-

ings as servo-actuators to stabilize the rotating stall in axial compressors. The basis of his


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Figure 2.4: Illustration of inlet guide vanes to suppress surge [24]


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Figure 2.5: Illustration of air injection technique for compressor stabilization [23]


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Figure 2.6: Illustration of experimental setup for Spakovszky test rig [29]

investigation was that the blade tip clearance in axial flow compressors has a strong im-

pact on compressor stability and it also plays a major role in the interaction between the

rotordynamic shaft deflections and the aerodynamic behavior of the compressor. The ob-

jective was to use the magnetic bearing servo-actuator to actively whirl the shaft to induce

unsteady variations of the rotor blade tip clearance which would suppress prestall dynam-

ics. This would help in suppressing stall and improved the stability operating region of the

axial flow compressors. The test compressor facility consisted of a single stage axial flow

compressor that was capable of a pressure ratio of 2.05 and design speed of 17000 rpm.

Figure 2.6 shows a schematic of the test compressor. The driver of the train was a 3000

HP motor. The bearing nearest to the impeller was the magnetic bearing actuator, while the

other supporting bearing was a fluid film bearing.

2.5 Compressor surge test rig overview

Sanadgol [3]proposed a first of a kind innovative active surge control method by using

a magnetic thrust bearing for a high speed centrifugal compressor. The theoretical back-


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ground to her surge control method is based on the fact that compressor performance and

efficiency are highly dependent on the clearance between the blade tip and adjacent sta-

tionary shroud or also known as the scroll [12]. This relationship is especially important

for high pressure ratio compressors, because the specific volume of the gas is reduced sig-

nificantly at the blade exit, and hence this results in a short height design of the exit blades.

This makes these compressors much more sensitive to tip clearance adjustment because

the ratio of its tip clearance to its exit blade height is significantly larger than similar com-

pressors that would be operated at lower pressure ratios. Therefore, any adjustment to the

impeller tip clearance would affect the efficiency of the compressor [26]. This results in

less energy transfer from the impeller to the fluid and would result in a loss of pressure

ratio and efficiency of the compressor.

Sanadgol developed a mathematical model of the compressor to include the tip clear-

ance and compressor efficiency relationships from Senoo and Ishida [26] into a one di-

mensional incompressible compressor model developed by Greitzer [27]. The improved

model was capable of showing the sensitivity of the centrifugal compressor characteristics

curve parameters to impeller blade tip clearance. Her simulation results based on the model

showed that if the tip of the impeller can be axially modulated by moving the position of

the rotor shaft using a magnetic thrust bearing with sufficient speed and control authority,

the induced pressure modulation makes the control of surge promising. Figure 2.7 shows

the simulated result without the surge controller. The two plots shows pressure ratio p

plotted over time and delta change in impeller tip clearance cl plotted over time. As can

be observed, when surge occurs, pressure ratio pwould start and continue to oscillate.

There is no controller and hence cl remains at zero. Figure 2.8 shows the simulated re-sulted with surge controller in place. The plot shows that with a surge controller in place to

actuate the impeller tip clearance as can be seen in the change ofcl over time, there would

be almost no pressure ratio p oscillations over time. This simulation results show that the

surge controller could stabilize the compression system. Sanadgol chose to use magnetic


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bearings as the control actuators for this form of control method as it would fully make use

of the active nature of the magnetic bearing systems and make real time static and dynamic

positioning of the rotor and the modulation of the impeller tip clearance possible. Coupled

with the increase in the use of magnetic bearings for turbomachinery in recent years, an

innovative control philosophy as such is very attractive.

An industrial size compressor test rig was planned to be built to validate the model and

its results in ROMAC Laboratory at University of Virginia. The experimental results from

the compressor test rig would be able to provide actual results that could be used to refine or

enhance the mathematical model as required. The final validated model of the compression

system would be used for designing an active surge control systems through impeller tip

clearance modulations. The controller would be implemented into the compressor test rig

and used to assess the amount of compressor stable flow range that could be improved with

this form of active control surge suppression method. The test rig had been mechanically

designed and its components built over the years. However, it is still not yet operational

and therefore no test data had been obtained to support the simulation results. Further detail

engineering work was still required to be done on the test rig, before it could be assembled,

commissioned and operated into high speed for the experimental surge testing and data

collection.

2.6 Summary of literature review

This literature review highlights surge control research work performed to stabilize

surge and also increase the stable flow range of a centrifugal compressor. The combination

of surge avoidance (e.g, safety margin) and surge suppression (e.g, Inlet guide vanes) is

still widely applied methods in the industry as compared to active surge control. How-

ever, these two methods are already mature technology field and could only offer possible

incremental improvements[20]. Active surge control, on the other hand, is still a devel-


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Figure 2.7: Simulated surge occurrence without any surge controller [3]

Figure 2.8: Simulated surge control result with impeller tip clearance actuation [3]


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oping field that has the potential to product substantial improvement in surge control and

extending the stable operating range of the compressor. In addition, active surge control

using magnetic bearings offers even better potential since there are already a substantial

amount of industrial compressors suspended by magnetic bearings. This would serve as a

platform to further developed magnetic bearings as surge control actuators or mechanisms

in compressors. Therefore, the control theory put forth by Sanadgol has a good potential

as a practical means to actively control surge in high speed centrifugal compressors.


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Chapter 3

Engineering analysis and experimental

setup

3.1 Introduction

The compressor surge test rig mainly consists of a few sections. These are the driver,

the driven equipment and the supporting system such as the pipings and chilling unit. The

driver is a high speed induction motor that is controlled by a variable frequency drive. The

compressor was direct coupled to the motor. The compressor is the single stage type with

an overhung semi-open impeller and is fully supported on active magnetic bearings both

radially and axially. The compressor service is atmospheric air. Due to the unique test

location of the test rig, it required inlet piping with an inlet filter to take in clean air from

the atmosphere. And in order to install an orifice flow meter and also varying locations for

the throttle valve, the piping was assembled using vitaulic couplings to provide flexibility

of installation. The test rig had been developed and designed over the years to provide a

hardware to allow actual compressor surge testing. Major components of this rig had been

sponsored by Rotating Machinery and Controls Laboratory members, such as the single

stage overhung centrifugal compressor by Kobe Steel, the magnetic bearings by Revolve

38


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CHAPTER 3. ENGINEERING ANALYSIS AND EXPERIMENTAL SETUP 39

and the motor rotor by SKF. Though the hardware had been designed and fabricated over

the years, the test rig had not been assembled and tested before. Engineering review and

improvement work had to be performed in order to install and commission the test rig

safely and reliably. The following sections provide an overview of the test rig, the bearings

selection for the motor, the modal analysis performed on the piping supports using Ansys, a

rotordynamics analysis performed when the compressor rotor assembly had to be modified

with the addition of a spacer ring and the accessory components selection such as the chiller

and the orifice flow meter.

3.2 Test Rig Overview

The motor is an induction motor rated for 125 KW at 30000rpm. It was provided by

SKF and was driven by a variable frequency drive by Alcomel. The motor was designed

with more than sufficient required power to drive the test rig over its entire operating range

and up to a maximum design speed of 23000rpm. As per the compressor supplier Kobe, the

compressor only requires 52KW of power at its maximum speed. The coupling selected

was a Thomas flexible disk-pack coupling. This coupling allows both axial and radial mis-

alignment between the motor and the test sections. The coupling was balanced and was

rated for the maximum design speed of 23000rpm [4]. The test section consisted of two

radial magnetic bearings and one thrust bearing built by Revolve. Figure 3.1 shows the

cross sectional drawing of the compressor test rig. Position measurements are very critical

in these active magnetic bearings as they provide the position feedback to the bearing con-

trollers to ensure proper control of these bearings to its optimum reference positions. The

two radial bearings use reluctance sensors to provide accurate position measurements of

the compressor rotor for the controller. The axial position is measured with two SKF 5mm

button eddy current probes. One probe is placed on either side of the thrust bearing and

targets a shoulder on the compressor rotor to measure the axial position of the rotor for the


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Figure 3.1: Compressor test rig cross sectional drawing

controller.

The compressor is overhung in design and only has a single stage. Its impeller is un-

shrouded and it can be used with either a vane or vaneless diffuser. At the current test

facility, it has been setup as a vaneless diffuser. The compressor is rated to provide a max-

imum flow rate of 2500m3/hr and develops a pressure ratio of 1.7. The required gas power

at this point would be 52KW and the operating speed would be 23000 rpm. In order to

investigate surge and stall, the compressor housing is drilled and tapped with holes to allow

the connection of twenty Kulite silicon on silicon pressure transducers. This arrangement

of sensors enables the capture of stall cells and visualization of surge or stall effects in the

compressor. The major data acquisition is performed by Lab-view. High speed data acqui-

sition cards (DAQ), PXI-6052 and PXI-6071, are used together with National Instruments

(NI) signal processing cards [4].

A thorough review of the design and mechanical components of the test rig was con-

ducted as a preparation for the next phase - which was to assemble, install and commission

the test rig. Hence this mechanical design audit is an important step to ensure the compo-


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Figure 3.2: Exploded view of compressor test rig

nents are finally designed to meet the required design speed and performance as required

in the compressor characteristic performance. A few critical mechanical items were iden-

tified that required furthering engineering analysis and improvement. Figure 3.2 shows the

exploded view of the compressor and the test section, where all the magnetic bearings are

housed.

3.3 Motor bearings improvement

The motor was designed to be supported on ball bearings that had a 47mm outer

diameter and an inner diameter of 25mm. For anti-friction bearings or commonly also

called rolling element bearings, a speed factor evaluation is commonly used as a evaluation

criteria to determine if the bearings selected for an equipment is the appropriate bearing

type and would provide a suitable running life for the bearings. The speed evaluation is

also used as a decision indicator to identify if a machine is suitable to be supported on ball

bearings or should it be upgraded to a fluid film bearings, which has always been preferred

as the supports for rotating shafts due to its better damping properties. This speed factor

is commonly known as the ndm factor [8]. The guidelines for the speed limit ndm factor is

commonly found in industrial references such as API 610 for centrifugal pumps and also


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in various bearing manufacturer literature such as Timken bearings, who have published

recommended speed limit for particular bearing design. ndm is actually the multiplication

of the:

n= rotating speed in rpm

dm= average bearing diameter = (Outer diameter of bearing + Inner diameter of bearing)/ 2

The use of speed as a parameter in the ndm factor calculation allows this simple eval-

uation to determine the suitability of the bearing design. This is because speed directly

relates to many things such as friction, centrifugal forces and stresses encountered on the

bearing cages and balls. Coupled with the nominal size of the bearing, this factor readily

indicates if a bearing selection would likely work. The many limit references quoted in typ-

ical bearing standards are cumulative data compiled over years to arrive at a recommended

ndm value to aid in evaluation. It is commonly known in the industry that exceeding the

ndm numbers would compromise the reliability of a bearing significantly and increase the

chance of premature failure of the equipment. The motor is controlled on a variable fre-

quency drive (VFD) and there is a speed range where the motor can potentially run duringthe surge evaluation. Hence Table 3.1 summarizes the ndm for speeds between 16000 rpm

- 23000 rpm, which is the intended compressor test speed range. The dm is fixed for all the

cases since the housing bore of 47mm and shaft diameter of 25 mm already set the size of

the ball bearings. The motor dm is hence:

Motor dm = (47+25)/2 = 36

From the comparison illustrated in table 3.1, it is clear that the ndm factor for the test

rig exceeds the recommended limits stated in API 610 and Timken bearing manufacturer

literature. Even at the lower intended operation speed of 17000 rpm, the ndm factor already

exceeds the limitations listed in API 610 by 22%. In order to fulfill the ndm factor of 500

000, the compressor would have to run at a maximum speed of only 13800 rpm. This speed


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Speed (rpm) ndm factor API 610 [8] Timken [49] % Exceed API 610

23000 828 000 500 000 300 000 65 %

21000 756 000 500 000 300 000 51 %

19000 684 000 500 000 300 000 36 %

17000 612 000 500 000 300 000 22 %

Table 3.1: ndm - Speed limit calculation for compressor test rig at possible running speeds

is considered low for the compressor design and it will not develop significant pressure rise

at such a low speed and hence may not be suitable for experimental surge investigation.

However to completely redesign the motor bearings at this stage of the project would in-

volve significant cost and time. After much considerations, it was decided to still proceed

with the use of ball bearing. As the test rig is not intended to run continuously like typical

machinery in E.g. manufacturing plants and hence , it was decided to proceed with the ball

bearing design because it is expected the compressor will only run intermittently during

testing.

The SKF 7005CD/P5A angular contact ball bearings were originally selected as the

bearings for the motor rotor as these bearings had the required speed ratings to allow the

motor to run safely at its design speed of 23000rpm [4]. However, in order for the bearing

to run reliably up to 23000rpm, it was identified that meeting the speed limit alone is not

sufficient. Maintaining the lubrication grease within the bearings is equally important, if

not more important. At such a high design speed of 23000rpm, ball bearings will generate

significant amount of heat, and hence the viscosity of the lubricant would reduce as the

motor is spun up to the desired speed of 23000rpm. The reduction in the viscosity of the

grease at these high speed would result in the originally filled grease in the bearings to

flow easily out of the bearings and result in poor lubrication on these high speed bearings.

This leakage would occur because the SKF 7005 CD/P5A bearing selected was an open

design bearing and does not come with end seals, and hence it would not able to retain

the grease within the bearings. In order to apply this bearing correctly in an application,

modifications would have to be made in the bearing housing to have a grease path to the


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bearings and provide periodic greasing to these bearings in order to ensure there is sufficient

lubricants in the bearings. Seals would have to be designed in the bearing housings to keep

the grease inside the bearings. This approach would require significant modifications in

the existing motor housing and was not the optimum option for the test rig from a cost and

time considerations. Hence it was decided to upgrade the motor bearings to a sealed for life

bearings that would also have the adequate speed and load ratings for this service. Sealed

for life bearings are bearings that have grease prelubricated inside the bearings from the

manufacturers. The seals in these bearings would keep the grease inside the bearings and

allow the bearings to operate smoothly and without the need to provide periodic greasing

to these bearings.

Hybrid bearings typically refers to a class of bearings that has material of construc-

tion besides regular bearing steel. Commonly used construction material for ball bearings

are stainless steel or chrome steel. Hybrid bearing incorporates the use of ceramic material

for its balls and the rest of the construction material for the bearing are retained as steel

material. The unique combination of ceramic material with steel material offers significant

amount of advantages and therefore hybrid bearings have been also termed as performance

bearings since its introduction decades ago [47]. Bearing manufacturer FAG began the

study of the use of ceramic materials in rolling bearings some 30 years ago. And out of

the many different ceramic material studied, silicon nitride (Si3N4) is found most suitable

to be used in bearings applications. Silicon nitride is a hard solid substance. Its main

component in silicon nitride ceramics and offers very superior shock resistance, thermal

and mechanical properties over many other ceramic materials. The silicon nitride balls are

much harder than metal and hence it reduces contact with the bearing inner race and outer

race tracks. This reduction in friction results in less wasted energy and allowed the hybrid

bearings to achieve higher speed ratings as compared to conventional steel bearings of the

same size. Silicon nitride also has a much lower density than bearing steel. Its density is

only about 40% of the density of bearing steel [47] This results the balls or rollers made


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from silicon nitride to weigh much less and have lower inertia. This means less damage to

the cage during rapid start up and stops and also much less frictional forces in the bearings.

The hybrid bearings can hence runs much cooler, last longer and also has higher speed

rating.

Besides the above mentioned mechanical properties given by the hybrid bearings, it

also provides a key property that is also valuable when the author decides to select hybrid

bearings for the motor. The hybrid bearings have very good electrical insulation property.

This makes them especially useful for induction motor or synchronous motor applications,

because the material insulation offers protection from any form of electric arc damage

which is a result of stray current in electrical machines like the motors. Silicon nitride

material in hybrid bearings provides insulation from electric currents in both AC and DC

motors [47]. Table 3.2 compares the main electrical and mechanical properties between

regular bearing steel and silicon nitride used in hybrid bearings. It is clearly obvious that

silicon nitride is superior to steel in many aspects and hence make a more suitable material

for ball bearing balls or rollers.

Material Property Bearing Steel Silicon Nitride

Density [g/cm] 7.9 3.2Hardness, HV10 [kg/mm] 700 1600

Modulus of elasticity [GPa] 210 310

Thermal expansion [x10/k] 12 3

Electrical resistivity [m] ~0.4e-6 ~10e12

Dielectric strength [kV/mm] - ~15

Relative dielectric constant - ~8

Table 3.2: Material construction for ball bearings [47]

These hybrid bearings are typically designed with a seal designed on both sides of the

bearings. Unlike the open bearing design of the SKF 7005 CD, the sealed designs main

advantage is that lubricant are kept away from contaminants as they are sealed in within the

bearing during manufacturing. Which means the grease is expected to be in the highest state

of cleanliness and hence minimum impurities. And these sealed bearings, when operated


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Figure 3.3: Comparison between SKF 7005 CD bearing versus SKF hybrid bearing

on site, it requires no further re-greasing. The seals on both side of the bearings also act

as protectors that keep out any form of contaminants that may commonly infiltrate in open

design. Figure 3.3 shows a comparison between the two bearings.

Deep-groove ball bearings is one of the most common type of ball bearings in the world

of bearing applications. It main advantage is its capability to handle both radial and direc-

tional thrust load. Its only disadvantage is that it has a low tolerance for misalignment

and hence for installation that involves deep groove ball bearings, alignment accuracy is

very important. And for this particular high speed motor application, Browns [4] esti-

mates that the axial load on the motor would be relatively low. Hence this make deep

groove ball bearings suitable for this application. SKF 6005 series hybrid bearings fulfills

this requirement and typically had a speed limiting range for a ball bearing inner diame-

ter of 25mm to be 28000rpm, which is rated to run 20% above the desired design speed of

23000rpm. Manufacturers of ball bearings typically provide a reference speed and limiting

speed as specifications for a particular bearing. Limiting speed is preferred to be used as

the evaluation criteria over reference speed because they are typically determined based on

cumulative actual application experience from the industry, whereas the higher reference

speed stated in a bearing catalog is based on test lab conditions which has the lubrication,

cleanliness, alignment and installation of the test bearings onto the test rig set up in a very

controlled and near perfect condition. Hence the reference speed rating determined in such


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Dynamic C (kN) Static C0 (kN) Limiting rpm Bearing Type Lubrication Design

9.56 5.6 34000 Angular contact Non lubricated

7.02 4.3 18000 Explorer class Sealed Lubrication

11.9 6.55 16000 Regular deep groove Sealed Lubrication

11.9 6.55 28000 Hybrid Sealed Lubrication

Table 3.3: Deep groove versus angular ball bearings design properties for 25 mm inner

diameter bearings

testing condition are typically higher than the limiting speed. These test laboratory con-

ditions are very unlikely to be replicated easily in an actual application environment, e.g.

in the refinery or chemicals plant, where there are significant uncertainties such as dust,

bearing handling cleanliness, equipment alignment and the proper installation of the bear-

ings onto the equipment. Therefore, in the selection of bearings, the limiting speed limit is

preferred to be used over reference speed. Table 3.3 illustrates and compares the various

types of SKF bearings that could are dimensionally similar (25 mm inner diameter), but

has varying properties and lubrication design that would result in different bearing proper-

ties. As can be seen from the comparison, the hybrid bearings provided the best bearing

properties in terms of dynamic loading limit of the bearing, its static loading, the maximum

recommended speed limit and the type of lubrication provision. Because all these bearing

are of the same inner diameter size (25mm), their ndm factor would all still be high at the

designed running speed. The hybrid bearing is the best in class from such a comparison and

is therefore selected as the bearings for this high speed motor. This would give the motor

its best possible reliability and the self lubricated design of the bearing. It also means there

is no modifications required on the motor housing to provide for grease nipple and grease

path to the motor bearings. The SKF hybrid bearing model 6005-2RSLTN9/HCFC3WT

was selected. The details and explanation for its bearing designations are listed in table

3.4.


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Designations Descriptions

6005 Model of bearing and last 2 digit (x5) would be size of shaft

2RSL Low friction seal on both side of bearings to keep grease in

TN9 Injection molded snap type ring of glass reinforced fibre

HC5 Rolling elements are of Silicon Nitride material

WT Grease with polyurea. -40 Degree to +160 degree Celsius rangeC3 Radial clearance is greater than normal clearance for bearings

Table 3.4: Descriptions for hybrid bearing 6005-2RSLTN9/HCFC3WT

3.4 Piping system design

As discussed previously, the compressed medium for this test compressor is atmospheric

air. The suction of the compressor takes in air from the atmosphere, compresses it and then

discharges it into the atmosphere again. It is an open loop system and there is no recycling

of the discharge air or neither is there a blow off vent as is commonly noticed in the air

compressor used in the industry as a form of surge protection. Therefore the piping layout

and piping support is an important part of the test rig to ensure the air is brought smoothly

to the compressor and discharges the hot compressed air safely back to the atmospheric.

At the same time, instruments were also to be installed along the piping system to measure

the pressure, flow and temperature during the test run. These measurements were critical

to the accuracy of surge investigation and the development of the compressor characteristic

curve, and their location will be installed in accordance to Power Test Code ( PTC 10)

1997, which is an commonly applied industrial testing standards for compressors. Since

this compressor is a prototype and the nature of the research investigation is similar to

that of a new test compressor as if in a factory acceptance test, adopting PTC 10 would

be suitable. Pressure and temperature measurement tappings are drilled and spaced on the

inlet and outlet piping immediately before and after the compressor. The length of the

straight section piping immediately before and after the compressor were also designed

to meet the requirement of PTC 10 1997 [32]. The straight length of the piping at the

inlet and immediately at the discharge had to be maintained to ensure a smooth laminar


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flow of the air and hence ensure the recorded test data are stable. The suction piping is

of straight length 72 inches or 9 times of the pipe diameter. The pressure tappings were

placed upstream of the temperature probes as per PTC 10. This is mainly because the

temperature probes will be placed into the flow and hence as the air flow passes the probes,

turbulence will be created and hence if pressure is measured at locations after the pressure

probes, the pressure readings may fluctuate due to the turbulence. In a similar fashion,

the temperature and pressure tappings are placed on the discharge piping of 64 inches or

8 times the piping diameter. However, in this case, the pressure tappings will be closed to

the discharge on the compressor, followed by the temperature probes. The same reasoning

on the placement of the inlet pressure and discharge probes apply on the discharge piping

as well. Four measurement locations spaced at ninety degrees apart are installed for each

pressure and temperature measurement location on the inlet and discharge piping. This is

to allow average pressure and temperature measurements to be taken and ensure consistent

data comparison without bias to any particular side of the piping.

The piping system was also designed to include the flexibility to adjust the outlet

plenum. The outlet plenum is defined as the volume between the compressor discharge

to the discharge throttle valve. Mizuki and Tamaki et al [39, 42] investigated and found

that the discharge plenum has an effect on the surge points of the compressor. Since the

main objective of this test rig is to investigate active surge control, the author decided to

include the additional flexibility into the piping design by providing 3 throttle valve loca-

tion, such that the discharge plenum volume could be adjusted as required. Three throttle

valve locations were included in the overall piping length, with position 1 being closest

to the compressor discharge, position 2 at about mid span along the discharge piping and

position 3 as the furthest location away from the compressor discharge, and hence provided

the largest plenum volume. The initial valve position is recommended to be at position 1

so that it limits the amount of plenum volume for the compressor when it is brought into

surge and also serves as the position for initial testing and design of first surge controller.


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Figure 3.4: Piping layout to show the three discharge throttle valve positions

Figure 3.4 shows the three locations of the discharge throttle valve in the piping system.

3.4.1 Piping supports modal analysis

One of the main concern with piping support design was that the surge frequency

encountered in the compression system would occur close to the resonance frequency of

the piping supports. This concern was evaluated by performing a modal analysis on the

designed piping supports using Ansys finite element software to determine the natural fre-

quency of the supports. These would then be compared to estimated mild surge frequency

of the compression system, using the widely used Helmholtz resonator frequency formula

utilized by many researchers [3, 27, 6, 20, 42] as the mild surge frequency. Based on the lit-

erature review established in Chapter 2, once mild surge frequency (Helmholtz frequency)


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Figure 3.5: Frequencies of several types of aerodynamics flow instabilities by Willems [40]

is estimated, it is generally observed that deep surge frequency would be be lower than

mild surge frequency. The reason for this is that deep surge frequency is generated by the

plenum emptying and filling times, and is hence normally well below Helmholtz frequency

[44]. Willems [40] developed a summary plot to show the frequencies range of occurrence

for several types of aerodynamics flow instabilities. Figure 3.5 shows the plot referenced

to Willems research work and it summarized the various frequencies on the occurrence of

deep surge, classic surge, mild surge and also rotating stall.

One main observation from the literature review about using Helmholtz resonant fre-

quency estimation by modeling the compressor system as a lumped mass parameter model

developed by Greitzer [27, 28] is that there are varying approaches used by different re-

searchers in the estimation. An example would be Mizuki and Willems [42, 40] used the

exit area of the impeller as the equivalent area for the compression model. In Mizukis

experiment, he obtained an experimental surge frequency of 27 Hz versus his Helmholtz

estimate of 32 Hz. Meulesmann [6] uses the eye of the impeller for the equivalent area Ac

instead for his Helmholtz estimation. Maulesmann [6] concluded that using the Greitzer

lumped mass parameter compression model to estimate the Helmholtz frequency is not a

straightforward approach and explained it was because the equivalent compressor length,

areas and volumes are not always directly obtainable from physical dimensions of the com-

plex compression system. Willems [40] explained that Lc is the most difficult to determine

because it is difficult do determine the transient mass flow rate and at times, it is chosen


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to match to the experimental studies, instead of first being determined from the simula-

tion. Tamaki [39] also concluded from his experiment with different plenum volumes that

the lumped mass parameter compression model using the Helmholtz frequency estimate

would typically have an error of 10% as compared to the actual experimental frequencies

obtained.

Helmholtz frequency can be determined by [39]:

H= a

Ac/(VpLc) (3.1)

where Vp is the volume between the compressor discharge to throttle valve position or

commonly known as plenum volume, Lc is the compressor equivalent pipe length, a is the

speed of sound, and Ac is the inlet flow through area. For this compressor, the inlet flow

through area Ac is the area of air flow at the eye of the impeller [6] or it could also be the

area at exit of impeller, which is what some researchers used as a parameter. Figure 3.6

shows the schematic equivalent compression system model to determine the Helmholtz and

Figure 3.7 shows how Ac was obtained based on the impeller hub diameter and impeller tip

diameter at the inlet of compressor [6, 15, 39]. The dimensions of the hub diameter and

impeller tip diameter were obtained from Kobe Steels dimensional drawing of the impeller

and compressor assembly.

The equivalent pipe length Lc could be determined from the following formula [28, 44]:

Lc

Ac

model

=

dx

Ax

Actualpiping

(3.2)

where the integral would be over the entire actual length of the compressor test rig[39]

and Ax would be the actual piping area. Table 3.5 summarizes the estimated Helmholtz

frequency based on the resonator frequency equation using different equivalent areas.

This compressor test rig would have three Helmholtz frequency, since it has three dif-

ferent throttle valve locations which results in three different discharge plenum volumes


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Figure 3.6: Schematic of the equivalent model to represent compressor system

Figure 3.7: Flow area to be considered when calculating the Helmholtz frequency

Descriptions Using inlet area, short

pipe length [39]

Using inlet eye area,

full pipe length [39]

Using impeller outlet

area [40]

Vp(m3) 0.053 0.053 0.053Lc(m) 1.233 4.1 4.1

Ac(m2) 0.008 0.008 0.006

a (m/s) 341 341 341

fH (Hz) 19.1 10.54 9.13

Table 3.5: Helmholtz frequency estimated comparison table


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(The volume between the compressor discharge to the discharge throttle valve). The piping

supports natura

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KT Thesis July2009 Final Print