Transcript
Page 1: Experimental investigation of tribological …...EXPERIMENTAL INVESTIGATION OF TRIBOLOGICAL CHARACTERISTICS OF WATER-LUBRICATED BEARINGS MATERIALS ON A PIN-ON-DISK TEST RIG Yuriy Solomonov

EXPERIMENTAL INVESTIGATION OF TRIBOLOGICAL

CHARACTERISTICS OF WATER-LUBRICATED

BEARINGS MATERIALS ON A PIN-ON-DISK TEST RIG

Yuriy Solomonov

The University of Adelaide

School of Mechanical Engineering

Master of Philosophy Thesis

April 2014

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Abstract

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ABSTRACT

Friction is the most fundamental phenomenon accompanying the sliding motion of solid

bodies. Friction, vibration, and wear under conditions of contaminated water lubrication

are extremely important in many engineering applications such as water-lubricated

bearings, water pumps, and braking systems.

The aim of this project is to investigate the factors that could lead to an improvement in the

performance of water-lubricated bearings materials.

Previous studies have revealed the main factors contributing to power loss are friction-

induced vibrations, and wear in water-lubricated bearings. Those factors are the result of

contamination of the lubricant (sea water); bearing alignment (parallelism of the shaft and

shell); material characteristics; and condition of the contact (sliding) surfaces. The contact

mechanics of the water-lubricated bearings as well as the performance characteristics of

the bearings components on which friction is exerted also have a substantial influence on

the tribological characteristics of water-lubricated bearings materials.

Thus, the focus of the present study is on the effect of water contamination on the friction

coefficient, vibration, wear and the vibration–wear relationship under varying operational

conditions.

An experimental program was conducted to develop new methods and investigate the

effect of water contamination on the tribological characteristics of pairs of materials under

different operational conditions for water-lubricated bearings.

A Pin-on-Disk test rig was designed and built to adopt the operational environment of a

real water-lubricated bearing. This test rig was used to obtain experimental data regarding

the effect of water contamination on the long-term behaviour of the bearing systems, and

to investigate the friction, vibration, wear, and vibration-wear characteristics of the

materials. The effect of various parameters, such as the friction conditions, damping, and

operational environment on the behaviour of the bearing materials was also investigated.

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Abstract

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The experiments demonstrated that all three factors, namely contamination, material

properties and surface conditions, have a significant influence on the tribological

characteristics of water-lubricated bearings. It was also demonstrated that when the

operation of the water-lubricated bearing takes place in boundary and mixed regimes, the

adhesive and abrasive mechanisms of friction are significant and contribute to the

generation of excessive wear and vibration. This is contrary to what is claimed by many

manufacturers. It was observed that the wear mechanism in the water-lubricated bearing

materials was associated with low-frequency vibrations and severe contamination of the

lubricant. Also, as expected, the vibration–wear relationship of the water-lubricated

bearing materials was significantly affected by the contamination of the lubricant and can

be changed by magnetic field damping.

The present study identified the primary mechanism responsible for the high friction

coefficient, vibration, and wear to be a three-body mechanism caused by the abrasive

nature of the water contaminant. It was found that there was a significant increase in the

friction coefficient, vibration, and specific wear rate at the slowest sliding speed of 0.393

m/s. This is due to the boundary regime of lubrication, the adhesive-abrasive wear

mechanism, and specific material properties of NF22 (Railko) material. It was also

explored and reported that for a specific applied load of 8 N, at low and high sliding

speeds, and water contamination levels, damping has a strong effect on the vibration–wear

relationship which is also dependent on sliding speed and, as a result, on the lubrication

regime.

The significance of this experimental study is to improve the selection of water-lubricated

bearings materials and as a result, improve their performances. The outcomes of this

research project are:

Analysis of the existing types of materials, and experimental models and techniques

for modelling and simulating the operational conditions of water-lubricated

bearings

Identification of the existing problems associated with the contemporary

technology of water-lubricated bearings materials

Development of an experimental methodology and technique for the application of

a Pin-on-Disk test rig and determination of the main contributing factors

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Abstract

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Identification and analysis of various lubrication and operational conditions for

water-lubricated bearing materials and systems and development of further

recommendations for future work.

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Declaration

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DECLARATION

I, Yuriy Solomonov, certify that this work contains no material which has been accepted

for the award of any other degree or diploma in any university or other tertiary institution

and, to the best of my knowledge and belief, contains no material previously published or

written by another person, except where due reference has been made in the text.

I give consent to this copy of my thesis, when deposited in the University Library, being

made available for loan and photocopying, subject to the provisions of the Copyright Act

1968.

I also give permission for the digital version of my thesis to be made available on the web,

via the University’s digital research repository, the Library catalogue and also through web

search engines, unless permission has been granted by the University to restrict access for

a period of time.

Yuriy Solomonov

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Acknowledgements

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ACKNOWLEDGEMENTS

The following work presented in this thesis would have not been possible without the help

and invaluable support of my family, university staff, and my friends and colleagues from

ASC.

I am truly thankful to my Principal Supervisor, Mr Ian Brown, for his support, assistance

and effort in the planning, experimentation and guidance throughout this entire project. Mr

Brown, despite a heavy workload and looming deadlines, was most helpful and would

never hesitate to take the time to answer questions and responsively sort through various

ideas and problems.

Tatiana, my wife, has been extremely supportive in many ways and deserves a special

mention. Thank you for being patient with me when it counted most and encouraging me

to complete my thesis, especially during the most difficult period of my candidature.

The effort of my co-supervisor A/Prof. Reza Ghomashchi, throughout the final stage of my

project is also greatly appreciated.

The support of many people within the School of Mechanical Engineering, including

academic staff, mechanical and electronic workshop staff, postgraduate students, and

technical and administrative support staff, is also acknowledged.

I would also like to thank my friends and colleagues from ASC who have contributed in

one way or another throughout my candidature.

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Table of contents

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TABLE OF CONTENTS

ABSTRACT ......................................................................................................................... II

DECLARATION ................................................................................................................ V

ACKNOWLEDGEMENTS ............................................................................................. VI

TABLE OF CONTENTS ................................................................................................ VII

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

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

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

1.1 Overview .......................................................................................................................... 1

1.2 Scope and objectives ........................................................................................................ 5

1.3 Thesis outline ................................................................................................................... 7

CHAPTER 2 BACKGROUND AND LITERATURE REVIEW .............................. 9

2.1 Historical overview of friction ......................................................................................... 9

2.2 Fundamentals of friction ................................................................................................ 14

2.2.1 Dry regime of friction ......................................................................................... 14

2.2.2 Hydrodynamic regime ........................................................................................ 17

2.2.3 Boundary and mixed regimes ............................................................................. 19

2.3 Friction in water-lubricated bearings ............................................................................. 20

2.3.1 Lubrication regimes ............................................................................................ 21

2.3.2 Friction-induced vibrations ................................................................................. 24

2.4 Numerical models of friction in water-lubricated bearings ........................................... 25

2.4.1 Linear models ..................................................................................................... 27

2.4.2 Non-linear models............................................................................................... 29

2.5 Vibration–wear relationship in dynamic systems .......................................................... 30

2.5.1 Problem of vibration–wear dependency ............................................................. 31

2.5.2 Effect of damping on vibration and wear ........................................................... 32

2.6 Concepts of experimental apparatus and experimental technique ................................. 34

2.6.1 Pin-on-Disk system ............................................................................................. 34

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2.6.2 Effect of lubricant contamination ....................................................................... 35

2.7 Summary and research gaps .......................................................................................... 36

CHAPTER 3 PREVIOUS EXPERIMENTAL STUDIES OF TRIBOLOGICAL

CHARACTERISTICS IN WATER-LUBRICATED BEARINGS ............................... 38

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

3.2 Review of wear study .................................................................................................... 38

3.3 Review of the aft bearing study ..................................................................................... 40

3.3.1 Experimental apparatus....................................................................................... 40

3.3.2 Experimental results of the friction coefficient measurements .......................... 42

3.3.3 Experimental results of the wear tests ................................................................ 44

3.3.4 Discussions and conclusions of the experimental tests ...................................... 44

3.4 Conclusions .................................................................................................................... 45

CHAPTER 4 EXPERIMENTAL APPARATUS ...................................................... 48

4.1 Introduction .................................................................................................................... 48

4.2 Design requirements ...................................................................................................... 48

4.3 Variable sliding speed and applied load ........................................................................ 51

4.3.1 Variable sliding speed ......................................................................................... 51

4.3.2 Variable applied load .......................................................................................... 53

4.4 Bending arm ................................................................................................................... 54

4.5 Load cell ........................................................................................................................ 58

4.6 Data acquisition system ................................................................................................. 60

4.7 Water supply system ...................................................................................................... 61

4.8 Experimental test rig ...................................................................................................... 62

4.9 Experimental methods ................................................................................................... 64

4.9.1 Experimental programs ....................................................................................... 64

4.9.2 Samples preparation ............................................................................................ 64

4.9.3 Validation study .................................................................................................. 65

4.9.4 Investigation of friction ...................................................................................... 66

4.9.5 Wear experiments ............................................................................................... 66

4.9.6 Vibration-wear relationship study ...................................................................... 67

4.10 Microscopy examination .............................................................................................. 68

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4.11 Conclusions .................................................................................................................. 68

CHAPTER 5 MATERIALS ........................................................................................ 69

5.1 Introduction .................................................................................................................... 69

5.2 Effect of water contamination on bearing materials ...................................................... 71

5.3 Materials for water-lubricated bearings ......................................................................... 72

5.3.1 Polymer-based thermoplastic materials .............................................................. 73

5.3.2 Carbon - fibre reinforced materials based on thermosetting materials ............... 75

5.3.3 Shaft materials .................................................................................................... 77

5.4 Conclusions .................................................................................................................... 79

CHAPTER 6 VALIDATION STUDY ........................................................................ 80

6.1 Introduction .................................................................................................................... 80

6.2 Test plan and procedure ................................................................................................. 80

6.3 Results and discussions .................................................................................................. 81

6.4 Conclusions .................................................................................................................... 83

CHAPTER 7 EXPERIMENTAL INVESTIGATION OF FRICTION

CHARACTERISTICS....................................................................................................... 85

7.1 Introduction .................................................................................................................... 85

7.2 Experimental study of the effect of contamination on friction: experimental plan and

procedure ............................................................................................................................. 85

7.3 Experimental study of the effect of contamination on friction: results and discussions 87

7.3.1 Investigation of friction under water lubrication ................................................ 87

7.3.2 Investigation of the friction coefficient and the effect of contamination ........... 92

7.4 Conclusions .................................................................................................................... 96

CHAPTER 8 EXPERIMENTAL INVESTIGATION OF WEAR .......................... 99

8.1 Introduction .................................................................................................................... 99

8.2 Experimental study of wear: experimental plan and procedure .................................... 99

8.3 Experimental study of wear: results and discussions ................................................... 101

8.3.1 Wear tests .......................................................................................................... 101

8.3.2 Specific wear rate calculations ......................................................................... 108

8.4 Conclusions .................................................................................................................. 112

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CHAPTER 9 EXPERIMENTAL INVESTIGATION OF VIBRATION–WEAR

RELATIONSHIP ............................................................................................................. 114

9.1 Introduction .................................................................................................................. 114

9.2 Operational conditions and experimental methods ...................................................... 115

9.3 Experimental study of effect of water contamination on vibration–wear relationship:

results and discussions ....................................................................................................... 119

9.3.1 Friction force–time analysis.............................................................................. 119

9.3.2 Power spectral density analysis ........................................................................ 125

9.3.3 Vibration–wear analysis ................................................................................... 132

9.4 Conclusions .................................................................................................................. 140

CHAPTER 10 CONCLUSIONS AND RECOMMENDATIONS ........................... 144

10.1 Summary .................................................................................................................... 144

10.2 Conclusions ................................................................................................................ 147

10.3 Recommendations for future work ............................................................................ 149

REFERENCES ................................................................................................................. 150

APPENDIX A: PIN-ON-DISK ASSEMBLY DRAWINGS ......................................... 158

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List of figures

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LIST OF FIGURES

Figure 1.1 Diagram showing a simple tribological system ................................................... 1

Figure 1.2 Diagram showing the complicity and complexity of tribological processes

(Materials Tribology laboratory, 2008) ................................................................................. 2

Figure 1.3 An example of an aft bearing system (Solomonov et al., 2010) .......................... 4

Figure 2.1 Friction experiments suggested by Leonardo da Vinci. (Krim, 2002) ............... 10

Figure 2.2 The original “Stribeck” curves obtained by Martens in 1888 (Martens, 1888-

1889) .................................................................................................................................... 13

Figure 2.3 Elastic deformation of crystal lattices during dry sliding (Holinski, 2001) ....... 15

Figure 2.4 Tribological changes during initial sliding of two solid bodies (Holinski, 2001)

............................................................................................................................................. 16

Figure 2.5 Hydrodynamic lubrication .................................................................................. 18

Figure 2.6 The three lubrication regimes in the “Stribeck” curve ....................................... 19

Figure 2.7 Lubrication regimes in water-lubricated bearings, reproduced from Kotousov

(2009), p. 7, Figure 2.2.1 ..................................................................................................... 22

Figure 2.8 Boundary and hydrodynamic regimes of lubrication, reproduced from Kotousov

(2009), p. 7, Figure 2.2.2 ..................................................................................................... 23

Figure 2.9 Typical baseline friction–speed curve (Pan et al., 1971) ................................... 28

Figure 2.10 Analytical two-degree model representing a submarine aft water-lubricated

bearing, as displayed in Simpson and Ibrahim (1996), p. 90, Figure 2 ............................... 30

Figure 3.1 Experimental wear rates for water-lubricated bearing materials (Biswell, 2007,

Cumberlidge, 2009, WÄRTSILÄ, 6/09/2007) .................................................................... 40

Figure 3.2 Scaled test rig and major components, as displayed in Kotousov (2009), p. 13,

Figure 3.2.1 .......................................................................................................................... 41

Figure 3.3 Friction curves vs. sliding speed, m/s (rotation speed, rpm), reproduced from

Kotousov (2009, p.18) ......................................................................................................... 43

Figure 4.1 Schematic diagram of a Pin-on-Disk experimental apparatus ........................... 50

Figure 4.2 Variable drive system: a) 3-phase BALDOR: MM3550C-57 motor with b)

GENESIS: NEMA-4X/IP-65 adjustable frequency drive ................................................... 52

Figure 4.3 Calibration data for disk rpm versus sliding speed for the POD test rig ............ 53

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Figure 4.4 Load cell with one additional weight block (applied force 17.5 N) ................... 54

Figure 4.5 Bending arm with attached strain gauges ........................................................... 57

Figure 4.6 Calibration data friction force versus voltage output for bending arm 1 ........... 58

Figure 4.7 Load cell with pin and displacement sensor ....................................................... 59

Figure 4.8 Calibration data friction force vs. voltage output for load cell .......................... 59

Figure 4.9 Schematic diagram of the data acquisition system used for the measurements of

pin displacement and arm forces on the POD test rig .......................................................... 60

Figure 4.10 Schematic diagram of the water supply system used on the POD test rig ....... 61

Figure 4.11 Design sketch of the test rig identifying the major components ...................... 62

Figure 4.12 Fully-equipped Pin-on-Disk test rig ................................................................. 63

Figure 5.1 Typical baseline of viscosity of water vs. temperature T, 0C, reproduced from

Ginzburg et al. (2006, p.696), Figure 2 ............................................................................... 71

Figure 5.2 Water-lubricated bearing damage (subjected to long-lasting operation which

resulted in significant wastage and associated ovalisation of the bush), where Do=initial

dimension, Dp=actual dimension and Dw=wear due to water contamination as displayed in

Litwin (2009, p.44, Figure 6 and p.48, Figure 21) .............................................................. 72

Figure 5.3 PTFE test sample used during the validation study ........................................... 74

Figure 5.4 NF22 (Railko) sample used in the experimental study ...................................... 77

Figure 5.5 AISI 440C stainless steel test disk fitted on the POD test rig ............................ 78

Figure 6.1 Coefficient of friction of the PTFE pin against a stainless steel disk for a sliding

speed of 0.32 m/s ................................................................................................................. 82

Figure 7.1 Coefficient of friction of NF22 (Railko) material against stainless steel versus

normal applied load under dry conditions ........................................................................... 88

Figure 7.2 Coefficient of friction of NF22 (Railko) material against stainless steel versus

sliding speed under dry conditions ...................................................................................... 89

Figure 7.3 Coefficient of friction of NF22 (Railko) material against stainless steel versus

normal load under clean water-lubricated conditions .......................................................... 90

Figure 7.4 Coefficient of friction of NF22 (Railko) material against stainless steel versus

sliding speed under clean water-lubricated conditions ........................................................ 91

Figure 7.5 Coefficient of friction of NF22 (Railko) against stainless steel for 1%

contaminated water lubrication ............................................................................................ 93

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Figure 7.6 Coefficient of friction of NF22 (Railko) against stainless steel for 2%

contaminated water lubrication ............................................................................................ 93

Figure 7.7 Coefficient of friction of NF22 (Railko) against stainless steel for 4%

contaminated water lubrication ............................................................................................ 94

Figure 7.8 Coefficient of friction of NF22 (Railko) against stainless steel for 6%

contaminated water lubrication ............................................................................................ 94

Figure 7.9 Coefficient of friction vs. water contamination of NF22 (Railko) material

(sliding speed=0.393 m/s) .................................................................................................... 95

Figure 8.1 Mass loss for NF22 (Railko) material at different load and speed values under

1% water contamination .................................................................................................... 102

Figure 8.2 Mass loss for NF22 (Railko) material at different load and speed values under

2% water contamination .................................................................................................... 102

Figure 8.3 Mass loss for NF22 (Railko) material at different load and speed values under

4% water contamination .................................................................................................... 103

Figure 8.4 Mass loss for NF22 (Railko) material at different load and speed values under

6% water contamination .................................................................................................... 103

Figure 8.5 Mass loss versus degree of water contamination for NF22 (Railko) material at a

sliding speed of 0.393 m/s ................................................................................................. 104

Figure 8.6 Micrograph of pin’s worn surface before/after a full cycle of experiments, at a

magnification of X500 ....................................................................................................... 105

Figure 8.7 Micrograph of pin’s worn surface after a full cycle of experiments, at a

magnification of X100 ....................................................................................................... 106

Figure 8.8 Micrograph of pin’s worn surface after a full cycle of experiments ................ 107

Figure 8.9 Micrograph of stainless steel disk’s worn surface after a full cycle of

experiments, at a magnification of X100 ........................................................................... 107

Figure 8.10 Specific wear rate for NF22 (Railko) material at different load and speed

values with 1% water contamination ................................................................................. 109

Figure 8.11 Specific wear rate for NF22 (Railko) material at different load and speed

values with 2% water contamination ................................................................................. 109

Figure 8.12 Specific wear rate for NF22 (Railko) material at different load and speed

values with 4% water contamination ................................................................................. 110

Figure 8.13 Specific wear rate for NF22 (Railko) material at different load and speed

values with 6% water contamination ................................................................................. 110

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Figure 8.14 Specific wear rate versus degree of water contamination of NF22 (Railko)

material for a sliding speed of 0.393 m/s ........................................................................... 111

Figure 9.1 POD test rig equipped with a magnet ............................................................... 116

Figure 9.2 Calculated friction force (N) versus time for undamped and damped conditions

under load 8 N and sliding speed 0.393 m/s ...................................................................... 121

Figure 9.3 Calculated friction force (N) versus time for undamped and damped conditions

under load 8 N and sliding speed 1.557 m/s ...................................................................... 124

Figure 9.4 Welch power spectral densities for damped and undamped conditions under

load 8 N and 0.393 m/s, sliding speed ............................................................................... 128

Figure 9.5 Welch power spectral densities for undamped and damped conditions under

load 8 N and 1.557 m/s, sliding speed ............................................................................... 131

Figure 9.6 Calculated RMS values for NF22 (Railko) material at normal and damped load

(8 N) ................................................................................................................................... 134

Figure 9.7 Specific wear rate for NF22 (Railko) material for undamped and damped

conditions ........................................................................................................................... 137

Figure 9.8 Microscopy of the pin’s worn surface before and after a full cycle of

experiments at a high sliding speed of 1.557 m/s, at a magnification of X30 ................... 139

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List of tables

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LIST OF TABLES

Table 3.1 Basic parameters for the Wärtsilä wear tests (Biswell, 2007) ............................. 39

Table 3.2 Basic parameters of the bearing system (Kotousov, 2009, Solomonov et al.,

2010) .................................................................................................................................... 42

Table 4.1 Design requirements for the Pin-on-Disk test rig ................................................ 51

Table 4.2 Design requirements for the bending arms used on the POD test rig .................. 55

Table 4.3 Technical characteristics of bending arms ........................................................... 57

Table 5.1 Physical properties of PTFE material .................................................................. 74

Table 5.2 Physical properties of NF22 (Railko) material (WÄRTSILÄ, 6/09/2007) ......... 76

Table 6.1 Technical parameters adopted for the validation study ....................................... 81

Table 7.1 Operational parameters adopted for experimental study ..................................... 86

Table 8.1 Parameters for wear test ..................................................................................... 100

Table 9.1 Technical parameters adopted for the vibration–wear experiments .................. 117

Table 9.2 Calculated RMS acceleration values at different sliding speeds, lubrication, and

contamination conditions, under an 8 N load .................................................................... 133

Table 9.3 Average mass loss (g) at different sliding speeds, lubrication, and contamination

conditions, under an 8 N load ............................................................................................ 136

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CHAPTER 1 Introduction

1

CHAPTER 1 INTRODUCTION

1.1 Overview

Since prehistoric times, human activities have involved friction. It is due to friction that we

are able to stand, run, start a fire, or even swim. Yet while friction is useful and necessary

for many human activities, it can also create many technological problems.

Friction takes place in all mechanical applications. These can include bearings, braking

systems and transmissions that involve two interacting surfaces, resulting in energy loss

and wear. Mankind has therefore made every possible effort to defeat the negative effects

of friction during the whole of human history.

It is known that friction can be reduced by lubrication. Lubrication involves placing

another material such as water or oil between the two surfaces (refer to Figure 1.1) in order

to extend the life cycle of a machine or a mechanism. In the 1960s, the science and

technology concerning friction, wear, and lubrication were unified under the discipline

named “tribology.” (Hori, 2006).

Figure 1.1 Diagram showing a simple tribological system

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CHAPTER 1 Introduction

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The modelling of tribological behaviour is a complex and difficult task. It is extremely

important for all stages of the life cycle of a machine and its components (as shown in

Figure 1.2). During machine design, accurate tribological simulation allows material

performance to be predicted and the mechanical design of materials and lubricants to be

engineered in a such way that the life cycle of the machine elements can be optimised

(Nikolakopoulos and Papadopoulos, 2008).

Figure 1.2 Diagram showing the complicity and complexity of tribological processes

(Materials Tribology laboratory, 2008)

The problem of friction, vibration and wear in rotating machinery is significant in many

engineering applications, particularly in water-lubricated bearings where it is undesirable

to experience friction and vibration between moving parts (Hori, 2006). Therefore,

increasing interest has been expressed in rotor dynamics, particularly the stability problems

encountered in high speed rotating machinery supported by water-lubricated bearings.

Three types of lubrication regimes are considered in water-lubricated journal bearings:

boundary, mixed and hydrodynamic. In hydrodynamic lubrication regime, lubricant is

squeezed between a rotating shaft and its bearing. If the load is relatively low and the

speed of the rotating shaft is high enough, a complete film of lubricant is formed (UoS,

2008). In real engineering applications using this type of lubrication, such as in propulsion

shaft systems, hydroelectric turbines and water pumps, the frictional energy loss is due to

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CHAPTER 1 Introduction

3

the viscous forces in the lubricant and any water contamination (UoS, 2008, Kotousov,

2009, Maru et al., 2005, Ibrahim, 1994a, Maru et al., 2007b, Meuter, 2006, Hori, 2006,

Nikolakopoulos and Papadopoulos, 2008). Viscosity cannot be completely eliminated,

however, because the separation of the surfaces is dependent upon it; as the viscosity

decreases so does the separation until the surfaces come into contact (UoS, 2008). When

the rotational speed of the shaft in water-lubricated bearing decreases, contact begins to

occur as a continuous film of fluid is being broken. At this stage, the situation is a mixture

of hydrodynamic and boundary lubrication, which is called, mixed lubrication. If the shaft

rotation is decreased further, the film of lubricant is reduced to localised patches which are

only a few molecules thick. This type of lubrication is known as boundary lubrication. In

boundary lubrication, the friction coefficient does not depend on the viscosity of the

lubricant but rather on its material properties (UoS, 2008).

“A good boundary lubricant is one that will attach itself firmly to the clean surfaces formed

as the cold-welded junctions are sheared. A layer is then formed that acts as a lubricating

film. If that layer can be easily sheared, then the friction is low. Typically, when the

coefficient of friction is of the order of 0.1, the wear is insignificant” (UoS, 2008).

Water-lubricated bearings are one of the most important and promising directions for

further development in the ship-building industry. Propeller shafts of ships and submarines

are supported in stern-tube bearings. These bearings are water-lubricated, but the rotational

speed of the shaft may not be high enough to achieve a hydrodynamic regime. In this case,

unlubricated frictional mechanisms would prevail at the region of the surface area in direct

contact. The overall friction torque would be considerably higher than in the hydrodynamic

regime (full separation of the sliding surfaces) (Simpson and Ibrahim, 1996). Therefore,

contact mechanics and friction phenomena in water-lubricated bearings are of interest to

contemporary researchers and engineers, especially in the field of tribology. During the last

two decades, many theoretical analyses and experimental studies have been undertaken to

investigate the operational characteristics and to identify the best materials for water-

lubricated bearings (Harrison, 2008).

For marine engineering applications, two types of aft bearing systems exist. One is a water-

lubricated bearing system that is mainly applied to small vessels. This is because of the low

allowable bearing pressure (Yamajo and Kikkawa, 2004). The other type of aft bearing

system is an oil-lubricated system for large vessel applications (Read and Flack, 1987).

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This is because, at low speeds, oil-lubricated bearings provide better boundary lubrication

and therefore can withstand higher loads compared to water-lubricated bearings.

A water-lubricated journal bearing (a stern-tube bearing) is a simple bearing in which a

shaft ("journal") rotates in the bearing with sea water as the liquid lubricant. Journal

bearings are simple, compact, and inexpensive to manufacture and maintain. This type of

bearing is also characterised by lightweight, high load-carrying capacity and good damping

characteristics.

According to Ryadchenko (2003), the mechanics of contact in journal bearings can be

described as follows. When a journal bearing begins rotating at low speed, little lubricant

exists between the shaft and the bearing at the contact area. This is a condition when

lubrication is provided by boundary and mixed regimes of lubrication, and high friction is

generated. When the shaft rotation approaches sufficient speed, the lubricant wedges into

the contact point and hydrodynamic regime is attained. Various approaches have been

developed for keeping the hydrodynamic regime in a journal bearing. However, with the

continuous development of journal bearing applications in rotating machinery, the serious

wear, power loss, and friction-induced vibrations still exist and need to be addressed

(Ryadchenko, 2003).

As an aft bearing system normally utilises a water-lubricated bearing, the water-lubricated

bearing plays a key role in marine engineering (see Figure 1.3).

Figure 1.3 An example of an aft bearing system (Solomonov et al., 2010)

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Aft shafts used in submarines, boats, and ships are supported by water-lubricated bearings

that are lubricated and cooled by sea water. One of the problems associated with this is that

“during start-up and shutdown, the rotation speed of the shaft may not be high enough to

achieve full separation of the surfaces by water. In this case, unlubricated frictional

mechanisms prevail at that portion of the surface area in direct contact. In addition, the

overall friction torque is considerably higher than the normal running value” (Simpson and

Ibrahim, 1996).

Another example of friction-related problems is the appearance of power loss, vibration

and excessive wear. This is due to two variables: the viscous forces in the lubricant; and

lubricant contamination by sand. Other researchers (Ledocq, 1973, Mokhiamer et al., 1999,

Tworzydlo et al., 1994, Qiao and Ibrahim, 1999) have theoretically and experimentally

investigated these negative influences of friction.

1.2 Scope and objectives

The problem of water contamination in water-lubricated bearings and its effect on bearings

materials have received limited attention from researchers (Younes, 1993, Mosleh et al.,

2002, Solomonov et al., 2010, Maru et al., 2007a, Maru et al., 2007b). According to

Simpson and Ibrahim (1996), “those studies that have been conducted were dominated by

experimental tests of section scaled models that emulated the actual bearing dynamics”.

However, the materials tribological behaviour was not considered. Simple analytical

models such as one- and two-degree-of-freedom models were put forward to predict and

simulate water-lubricated bearings (Kotousov, 2009, Tworzydlo et al., 1994, Younes,

1993, Mosleh et al., 2002). However, the current theoretical and experimental models are

too simplistic and inadequate (Ibrahim, 1994a, Hori, 2006, Kingsbury, 1997, Platt et al.,

1994, Younes, 1993, Hirani et al., 1997, Simpson and Ibrahim, 1996, Kotousov, 2009,

Tworzydlo et al., 1994, WÄRTSILÄ, 6/09/2007). Not considered within these

investigations is the role of nonlinearity, which is due to the friction–speed curve, and the

effects of lubricant contamination and damping on water lubricated bearing materials

(Simpson and Ibrahim, 1996, Kotousov, 2009, Tworzydlo et al., 1994, Younes, 1993,

Mosleh et al., 2002).

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With the growing interest in ship and submarine construction technologies in South

Australia, further experimental investigation and theoretical development of nonlinear

models that emulate the tribology of water-lubricated bearings and their materials are of

great interest to researchers and designers in marine engineering.

The significance of this research project is to extend the previous work of Kotousov (2009)

and to develop new methods for experimental investigations (Solomonov et al., 2010) of

tribological characteristics of water-lubricated bearings materials using a Pin-on Disk test

rig. The goal is to achieve a better understanding and modelling of tribological processes in

marine water-lubricated bearings. This enhanced understanding could lead to further

development of new materials and design solutions with a focus on efficiency and

noiselessness.

The expected outcomes from this research project are:

Identification of the existing problems in the contemporary technology of water-

lubricated bearing materials and systems

Analysis of the existing types of materials, and analysis of experimental and

theoretical models of water-lubricated bearings

Further development and extension of the experimental methodology of

investigation of friction, vibration, and wear of water-lubricated bearings materials

using a Pin-on-Disk test rig

Identification and further development of methods to analyse friction, vibration,

wear, the vibration-wear relationship, and the effect of water contamination on the

tribological behaviour of water-lubricated bearing materials

Identification and further development of methods to analyse the effect of

contamination and damping on the vibration-wear relationship of water-lubricated

bearings materials

Development of practical recommendations to improve the materials design,

performance, selection, and overall modelling and performance of water-lubricated

frictional systems.

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1.3 Thesis outline

In Chapter 1, the background, scope and objectives of this project are introduced. A brief

summary of this thesis is then presented.

Chapter 2 presents existing and relevant literature on water-lubricated bearings and

friction, vibration, and wear problems associated with them. The commonly used

theoretical analysis and experimental techniques regarding water-lubricated bearings are

summarised. The brief overview of theoretical and experimental methods to analyse and

model the water-lubricated bearings is described. Proposed Pin-on-Disk experimental

methods used to analyse friction, are also introduced. Literature review is summarised and

research gaps are then identified.

Chapter 3 is the review of the results from a prior theoretical analysis and experimental

studies on the effect of water contamination on a water-lubricated bearing system which

were undertaken by other researchers in the School of Mechanical Engineering, University

of Adelaide and at Wärtsilä Pty Ltd, a UK-based supplier of water-lubricated bearing

materials. These previous studies became a starting point for the experimental

investigations undertaken herein.

Chapter 4 deals with the experimental Pin-on Disk test rig which is used for the

experimental study of friction, vibration and wear. The main elements of the test rig and

data acquisition system and applied methods are discussed.

Chapter 5 presents a review of contemporary polymer-based materials that are used for

water-lubricated bearings and then describes the process used in the selection of the

available materials for further experimental investigation.

Chapter 6 includes the results of a validation study program to verify the capability of the

newly-designed and fabricated test rig. It also includes a discussion on the results obtained

from the experimental program which was conducted to examine the effect of varying

levels of contaminated water lubrication on the friction coefficient of materials for water-

lubricated bearings.

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Chapter 7 discusses the experimental results which were obtained to analyse the effect of

contaminated water lubrication regimes on the mass loss and the specific wear rate of

materials used for water-lubricated bearings. This study investigated the sliding response

of NF22 (Railko) material against stainless steel.

Chapter 8 presents an experimental investigation of the effect of water contamination on

the vibration–wear relationship for water-lubricated bearings. The contacting

characteristics of sliding bodies were considered in order to ascertain the existence of

correlations between vibration phenomena and tribological aspects.

Chapter 9 presents the conclusions and recommendations for future work on theoretical

and experimental investigation of friction phenomena in water-lubricated bearing systems.

The results from this experimental investigation are also discussed and summarised.

Appendix A presents assembly drawings detailing the experimental Pin-on-Disk test rig.

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CHAPTER 2 BACKGROUND AND LITERATURE REVIEW

2.1 Historical overview of friction

Friction, the most fundamental phenomenon accompanying sliding movement of solid

bodies, is defined as the force resisting the relative lateral (tangential) motion of solid

surfaces, fluid layers, or material elements that are in contact. Friction can be subdivided

into several types (Ruina and Pratap, 2002):

Dry friction resists the motion of two solid bodies in contact. Dry friction can be

divided into two types: static and dynamic friction (also known as kinetic or sliding

friction).

Internal friction: is when the resisting force exists due to movement between the

components of a solid body while it is being deformed.

Fluid or Lubricated friction: is when two sliding solid bodies are separated by

lubricant.

Skin friction: is when a solid is dragged through a lubricant.

Several famous researchers have contributed to the understanding of friction in the past,

such as Guillaume Amontons, Leonardo da Vinci, Lohn Theophilus desaguliers, Leonard

Euler, and Charles-Augustin de Coulomb (Engineering-ABC.com, Courtel and Tichvinsky,

1964).

Leonardo Da Vinci (1452-1519) was one of the first scientists to investigate friction

(Courtel and Tichvinsky, 1964, Krim, 2002). He focused on many types of friction. He

found differences between rolling and sliding friction. Figure 2.1 presents sketches from

(Krim, 2002) showing Da Vinci “experiments to investigate: a) the force of friction

between horizontal and inclined planes; b) the influence of the apparent contact area upon

the force of friction; c) the force of friction on a horizontal plane by means of a pulley and

d) the friction torque on a roller and half bearing” (Krim, 2002).

Da Vinci was the first to proclaim two laws of friction. He claimed that “the frictional

resistance was the same for two different objects of the same weight but making contacts

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over different widths and lengths”. He also found that “the force needed to overcome

friction is doubled when the weight is doubled” (Krim, 2002).

Figure 2.1 Friction experiments suggested by Leonardo da Vinci. (Krim, 2002)

Da Vinci stated these two original laws of friction 200 years before Sir Isaac Newton (He,

2010, Zhuravlev, 2010). Da Vinci made the observation that different materials move with

different degrees of ease. He concluded that this was a result of the roughness of the

material; “smoother materials will have smaller frictions” (Krim, 2002). Da Vinci did not

publish his theories, so he never received recognition for his ideas. The only evidence of

their existence is in his voluminous journals.

Guillaume Amontons (1663-1705) rediscovered the two original laws of friction that had

been discovered by Da Vinci. Amontons concluded that “friction was predominately a

result of the work done to lift one surface over the roughness of the other, or from the

deforming or the wearing of the other surface” (Werktuigbouw&Tribologie, 2010,

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Zhuravlev, 2010). “For several centuries after Amontons' work, scientists believed that

friction was due to the roughness of the surfaces” (Werktuigbouw&Tribologie, 2010).

Leonard Euler (1707-1783), a famous mathematician, was also “much concerned with

friction problems. With the use of classical dynamics, he expressed the value of the

coefficient of friction by parameters which could be easily measured. Two of his famous

works were published under the auspices of the Royal Berlin Academy, one called

“Friction of Solid Bodies” and the other called “Decrease of Friction Resistance”. He

pointed out that the friction force is always tangential to the sliding velocity, and he

indicated the conditions of constant-acceleration motion with friction for plane and

inclined surfaces” (Courtel and Tichvinsky, 1964).

Charles Augustin de Coulomb (1736-1806) added to the second law of friction (Liang,

2005, Zhuravlev, 2010). He proclaimed that “strength due to friction is proportional to

compressive force” although “for large bodies friction does not exactly follow this law”

and, as a consequence, the friction force and the area of contact are not dependable (Liang,

2005, Zhuravlev, 2010). Coulomb published his study, and referred it to Amontons’

previous work. The second law of friction is known as the "Amontons-Coulomb Law"

referring to work done by the two scientists in 1699 and 1785, respectively (Werktuigbouw

& Tribologie, 2010, Zhuravlev, 2010). The Amontons-Coulomb law of friction has been

true “for many combinations of materials and their geometries but, unlike Newton’s first

law, nothing fundamental can be derived from it” (Werktuigbouw & Tribologie, 2010).

Philip Bowden and David Tabor (1950) provided a physical explanation for the laws of

friction (Werktuigbouw&Tribologie, 2010, Courtel and Tichvinsky, 1964). They found

that the area of actual contact is a very small percentage of the total area of contact. The

actual area of contact is formed by the asperities and depends on the applied load

(Werktuigbouw&Tribologie, 2010). As the load increases, more asperities contact occur

and the actual area of each asperity’s contact increases (Kaarstad, 2009).

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Overall, the above findings were formulated into three friction laws (Zhuravlev, 2010):

1. The friction force is directly proportional to the applied load. (the First Law of

Amontons)

2. The friction force is independent of the total contact area. (the Second Law of

Amontons)

3. Dynamic friction is independent of the sliding speed. (Coulomb's Friction Law).

A method to reduce the negative influence of friction by placing a lubricant, such as

grease, water or oil, between contacting surfaces has been known for centuries. If the

lubricant is placed between the two interacting surfaces, it allows significant reduction of

the friction coefficient and, as a consequence, the friction force (Teidelt, 2012).

Further research on the relationship between friction and lubrication was undertaken and

published in the late 1870s by Dr Robert H. Thurston at the Stevens Institute of

Technology (USA) (Thurston, 1879, Thurston, 1894), and in 1885 by Professor Adolf

Martens (1850-1914). The research was later finalised by Professor Richard Stribeck

(1861-1950) at the Royal Prussian Technical Testing Institute in Germany (Stribeck,

1902a).

Based on the results obtained in the form of so-called “Stribeck” curves by Professor

Stribeck in the early 1920s, the friction regimes for sliding lubricated bodies were

identified as static friction, boundary friction, mixed friction and hydrodynamic friction.

The “Stribeck” curve is a tribological characteristic which is utilised to characterise the

dependence of friction force on sliding speed between two liquid lubricated bodies

(Harrison, 2008, Hersey., 1934).

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Figure 2.2 The original “Stribeck” curves obtained by Martens in 1888 (Martens,

1888-1889)

The science of friction and lubrication was named “tribology” and introduced in 1966.

Further study of tribology became possible with the invention of the atomic force

microscope (AFM) in 1986 and was concentrated on the micro-scale mechanics of contact

established asperities with adhesion and deformation. This approach enables researchers to

investigate friction on the atomic scale (Palaci, 2007). Thus, researchers are able to

determine how the mechanisms of friction change at macroscopic scale under different

operational conditions.

Micro- and nano-tribology have been introduced in recent years (Palaci, 2007). Frictional

interactions in microscopically small components are crucial to the development of new

products in contemporary technology such as electronics, sciences, chemistry, and

microelectronics.

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2.2 Fundamentals of friction

Most textbooks on engineering mechanics recognise two types of friction opposing the

relative motion of bodies in mechanical contact: dry or Coulomb friction and fluid or

viscous friction.

2.2.1 Dry regime of friction

Dry friction may affect machine behaviour in a way that reduces or even disables machine

operation capacity. Dry friction-induced vibrations (often referred to as stick-slip) lead to

noise problems such as bearing squeal or chatter.

Dry friction between two solid bodies in mechanical contact is caused by the dry friction

force of two components: the kinetic and static friction forces. These forces develop in a

perpendicular direction to the contact plane and oppose the relative motion of the contact

surfaces. A dry sliding friction between two surfaces can be modelled as elastic and plastic

deformation forces of microscopic asperities in contact (Olsson et al., 1997).

Consequently, it is possible to manipulate friction characteristics by employing surface

films of suitable materials between the bodies in contact. These surface films can also be

the result of contamination or oxidation of the bulk material or material displacement

(Nouira, 2008).

Holinski (2001) made a statement as to why dry friction occurs in composite materials.

According to his work, “the reason for friction is lattice vibration. Lattice vibration occurs

during the sliding of solid surfaces when atoms of one surface make the atoms of the other

surface vibrate. Part of the mechanical energy which is required to move both surfaces

over each other is transformed into sound waves and heat” (Holinski, 2001). The

tribological interactions of a solid surface’s exposed face with interfacing materials and

environment may result in a loss of material from the surface.

According to Holinski (2001), on the atomic scale, the crystal lattices of both solid

materials in contact are in a state of equilibrium. When a shear stress is applied to one

component, both lattices deform elastically. If the shear is further increased and instability

is reached, the atoms move to a new position of equilibrium. The crystal lattice vibrates

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until all strain energy has been dissipated as seen in Figure 2.3, and all elastic energy has

been converted to heat.

Figure 2.3 Elastic deformation of crystal lattices during dry sliding (Holinski, 2001)

Holinski (2001) stated “It has been found that during the initial sliding of the surfaces of

two solid bodies over each other, friction force is increased, as is frictional temperature. In

this phase, no material transfer from one surface to the other occurs. Only at the frictional

maximum are particles transferred”, as shown in Figure 2.4 (Holinski, 2001).

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Figure 2.4 Tribological changes during initial sliding of two solid bodies (Holinski,

2001)

When the transfer is initiated, the friction force decreases until a level of stability is

achieved. A transfer film depends on material properties and grows to a certain thickness.

For example, in a graphite-based composite, a homogeneous layer is not formed. However,

small graphite islands are built. The friction force decreases after formation of islands. It

has been found that during the sliding of solid bodies of different composite materials

against each other, a friction layer is formed on the surface of the harder material. Surface

materials properties are changed by heat and friction. This friction layer is responsible for

the tribological characteristics such as friction coefficient, vibration, and wear rate

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(Holinski, 2001). Further investigations by Raman spectroscopy (named after Sir C. V.

Raman) and microprobe analysis show that the newly - formed transfer film has chemical

differences compared to the matrix of material. A surface layer is located on the harder

sliding composite material. Often, the total composition from the softer material is not

transferred but only some of the chemical components. Both layers are governing the

tribological characteristics of sliding systems. In some cases, chemical components are

added to one sliding material to form a particular friction layer to achieve the desired

tribological characteristics (Holinski, 2001).

In many cases, dry friction process leads not to transfer but to loss of material. This process

is known as "wear" (PML, 2009). Major types of wear include abrasion, adhesion

(friction), erosion, and corrosion. Wear can be minimised by modifying the material

properties or the surface properties of solids by “surface engineering” processes (also

called surface finishing) or by the use of lubricants (for frictional or adhesive wear) (PML,

2009).

According to Holinski (2001), depending on operational parameters such as pressure,

sliding speed and type of material, different frictional temperatures occur which give rise

to various power losses. Dry friction always generates heat and consequently diffusion

processes. Thus, dry sliding friction creates surface deformation which results in crystal

lattice deformation and higher concentration of dislocations. Also, chemical reactions such

as tribological oxidation and decompositions occur. All these factors finally lead to cracks,

fractures and wear.

2.2.2 Hydrodynamic regime

Rac and Vencl (2005) stated “Complete separation of sliding surfaces with lubricant can be

achieved by hydrodynamic lubrication. The selection of the design and the tribological

parameters in the region of the hydrodynamic lubrication should ensure adequate thickness

of the lubricant film and temperature of the bearing” (Rac and Vencl, 2005).

According to MarineDiesels (2011), “Hydrodynamic lubrication was first researched by

Osborne Reynolds (1842-1912) in an experimental test rig for modelling a liquid-

lubricated bearing. When a lubricant was applied between the shaft and bearing, Reynolds

found that the rotating shaft pulled a converging wedge of lubricant between the shaft and

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the bearing. He also noted that as the shaft gained velocity, the liquid flowed between the

two surfaces at a greater rate. The viscous lubricant produced a liquid pressure in the

lubricant wedge that was sufficient to keep the two surfaces separated. Under ideal

conditions, Reynolds showed that this liquid pressure was great enough to keep the two

bodies from having any contact and that the only friction in the system was the viscous

resistance or viscosity of the lubricant.”(MarineDiesels, 2011)

Figure 2.5 Hydrodynamic lubrication

Viscosity of the lubricant is an important parameter: the higher the viscosity of the

lubricant, the higher the friction between lubricant and shaft but the thicker the

hydrodynamic film. However, friction generates heat. Heat will reduce the viscosity and,

therefore, the thickness of the film which may result in shaft-bearing contact. Reduced film

thickness occurs when a lubricant with low initial viscosity is used (MarineDiesels, 2011).

Care needs to be taken to ensure that the distance between the two surfaces is greater than

the largest surface defect. The distance between the two hydrodynamicaly sliding surfaces

decreases with higher loads on the bearing, lower speeds, and less viscous fluids. A

hydrodynamic regime is an excellent method of lubrication. This is because it is possible to

achieve coefficients of friction as low as 0.001 with no wear between the moving parts.

However, because the lubricant is heated by the frictional force and since viscosity is

temperature-dependent, special additives can be used to decrease the viscosity temperature

dependence (MarineDiesels, 2011).

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2.2.3 Boundary and mixed regimes

Friction with boundary lubrication modes or in mixed (transition) operating conditions can

prevail in several practical cases.

Contact between the surfaces at a few high surface points (micro asperities) occurs during

boundary lubrication and then mixed lubrication as shown in Figure 2.6.

Mixed lubrication is the intermediate regime between boundary lubrication (where friction

is mostly due to asperity contact), and hydrodynamic lubrication (where the hydrodynamic

separation is achieved) (Qiang, 2009).

Figure 2.6 The three lubrication regimes in the “Stribeck” curve

For some materials, when a mixed regime occurs, the interacting surfaces contact at their

asperities developing heat from the localised loads introducing a co-called stick-slip

motion where some asperities can be broken off. When the high loads and temperature

occur, chemically active lubricant components interact with the contact surface. They form

a highly-resistant film, on the moving solid surfaces. This layer is capable of withstanding

the high loading pressure, and excessive wear or damage are eliminated (Mokhtar et al.,

1984, Nikolakopoulos and Papadopoulos, 2008).

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Friction-induced vibrations in boundary and mixed regimes are of particular concern to the

designers of mechanical applications involving sliding surfaces at comparatively low

speeds (up to 5 m/s) (Simpson and Ibrahim, 1996, Tworzydlo et al., 1994). These

applications include liquid-lubricated bearings, wheel-rail systems, disk brake systems, and

machine tool-work piece systems. Numerical analyses to predict vibrations and noise such

as chatter and squeal in a mixed regime have not been developed, and no appropriate

theory exists that can be generalised to analyse these vibrations. Most research activities

are based on experimental tests of physical models. In general, friction-induced vibrations

on boundary and mixed regimes vary and depend on temperature, normal load, speed, and

other conditions (Simpson and Ibrahim, 1996).

2.3 Friction in water-lubricated bearings

The plain bearing was initially invented to reduce friction by placing animal fat between a

sliding shaft and a bore. Plain bearings are the simplest and the least expensive type of

bearings. They are also, lightweight, compact and are capable of withstanding high loads.

Historically, the plain bearings originated approximately 4,500 years ago from Phoenician

seafarers. Their plain bearings were built from lignum vitae, a type of wood with an

appropriate combination of density, strength, and toughness (Weichsel, 1994).

According to Weichsel (1994), modern plain bearings can be divided into three basic

categories based on the lubricant system required for successful operation:

Bearings that require oil, grease or some other lubricant to operate. They receive

this liquid or semi-solid lubricant from an outside source.

Bearings that contain the necessary lubricant within their walls: i.e., a plastic

bearing such as polyacetal, which utilises a silicone lubricant.

Bearings that are in and of themselves the lubricant, such as metallised carbon

graphite or have a running surface that contains Teflon.

Contemporary aft water-lubricated bearings are plain bearings that require lubrication by

sea water. The problems of power loss, vibration, and wear associated with water-

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lubricated bearings are of great interest to researchers (Rac and Vencl, 2005, Hori, 2006,

Kingsbury, 1997).

2.3.1 Lubrication regimes

The theory on which water-lubricated bearings’ designs and calculations is based is

complex. It presupposes that the bearings’ operating conditions such as the load, size,

clearance, and properties of the materials and lubricant are known. Various

recommendations concerning the selection of those values, such as those proposed by Rac

and Vencl (2005), appear in the literature. For example, for a known load and sliding

speed, the preliminary sizes of the hydrodynamic sliding bearings can be selected.

Alternatively, for a selected size of bearing, the load capacity for a given speed can be

determined from the recommendations. These values are related to the steady-load

condition. If the load varies in magnitude and/or direction, the procedure for the selection

of bearing parameters is more complex and requires additional design considerations (Rac

and Vencl, 2005).

Friction in this type of water-lubricated bearing can be described by the “Stribeck” model

(Stribeck, 1902b). According to this model, four different friction–speed regions exist (see

Figure 2.7):

1. Static friction, where two static surfaces are mostly in contact with each other.

2. Boundary lubrication, where the fluid films are negligible and sufficient asperity

contact exists.

3. Partial fluid lubrication (transition or mixed) where two surfaces are partly

separated, partly in contact and the relative motion of the surfaces is insufficient to

generate the hydrodynamic action required to separate them completely (Ivanov

and Ivanov, 2012, Barwell, 1984).

4. Hydrodynamic fluid lubrication, which is sufficiently developed to completely

separate the interacting surfaces (Barwell, 1984).

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Figure 2.7 Lubrication regimes in water-lubricated bearings, reproduced from

Kotousov (2009), p. 7, Figure 2.2.1

When a water-lubricated bearing starts up (Region 2), the combination of a low sliding

speed, low viscosity, and high load will produce boundary lubrication. Boundary

lubrication is characterised by the small amount of fluid in the interface and the large

surface contact. This results in extremely high friction (STLE, 2008, Solomonov et al.,

2010).

As the sliding speed and fluid viscosity increase, or the load decreases, the surfaces begin

to separate (Region 3), and a fluid film begins to form. The film is still insufficient but acts

to support more of the load. Mixed lubrication is the result of a fast drop in the friction

coefficient (STLE, 2008). This results in less surface contact and more fluid separation

(Neveu et al., 2012). Any of the following can prevent the build-up of a film thick enough

for hydrodynamic lubrication: insufficient surface area, a drop in the sliding speed of the

moving surfaces, reduced quantity of lubricant delivered to a bearing, an increase in the

bearing load; an increase in lubricant temperature resulting in a decrease in viscosity, or

contamination of the lubricant. If the film thickness is still insufficient, the highest

asperities may not be fully separated by lubricant (STLE, 2008, Solomonov et al., 2010).

The surfaces will continue to separate as the sliding speed or viscosity increase until full

fluid film separation has been achieved and no surface contact exists (Region 4). The

friction coefficient will reach its minimum and a transition to hydrodynamic lubrication

occurs. At this point, the load on the interface is entirely supported by the fluid film. There

Friction coefficient

Region 2

Region 3 Region 4

Region 1

Velocity

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is low friction and no wear in hydrodynamic lubrication since there is full fluid film

separation and no contact between asperities (STLE, 2008, Solomonov et al., 2010).

With an increase in the sliding speed, the friction increases further in the hydrodynamic

region due to fluid drag (friction produced by the fluid): higher speed may result in a

thicker fluid film, but it also increases the fluid drag on the moving surfaces. Thus, a

higher viscosity will increase the fluid film thickness but it will also increase the drag

(STLE, 2008, Solomonov et al., 2010).

Water-lubricated bearings experience boundary lubrication and mixed lubrication at start-

up and shutdown (low speeds and thin film) before the transition to hydrodynamic

lubrication at normal operating conditions (high speeds and thick film) (STLE, 2008). For

example, during start-up, as the sliding speed between the shaft and bearing increases, the

change from boundary lubrication to mixed and then to the hydrodynamic regime is not a

sudden or abrupt process (Dulias, 2002). A boundary-type and then mixed lubrication

occur first, and then as the surfaces slide faster, the hydrodynamic-type lubrication

becomes predominant with the change of the position of the journal shaft as shown in

Figure 2.8 (Kotousov, 2009, STLE, 2008, Solomonov et al., 2010).

Figure 2.8 Boundary and hydrodynamic regimes of lubrication, reproduced from

Kotousov (2009), p. 7, Figure 2.2.2

hmin

P

R

o P

R

f

Hydrodynamic Regime

(Region 4) Boundary Regime

(Region 2)

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2.3.2 Friction-induced vibrations

When the friction coefficient depends on the sliding speed and has a negative slope with

respect to the velocity (as shown in Figure 2.7), the friction gives rise to different types of

friction-induced vibrations (EOV, 2004). These vibrations can include stick-slip

(Schallamach, 1963, Schallamach, 1971, Younes, 1993, Bhushan, 1980), quasi-harmonic

oscillation (Kotousov, 2009, Ibrahim, 1994a, Ibrahim, 1994b, Qiao and Ibrahim, 1999) and

surface-induced vibration (Platt et al., 1994, Ibrahim, 1994a, Ibrahim, 1994b).

A comprehensive literature review of friction-induced vibrations and related problems in

water-lubricated journal bearings was recently documented by Kotousov (2009). Several

distinct mechanisms can contribute to these different types of oscillations:

Self-excited vibrations of the stick-slip type occur at very low sliding speeds

(typically 0-0.3 m/s) that can be associated with start-up or shutdown. They are

attributed to the difference between the static and kinematic coefficient of friction

(Dautzenberg, 1986). At low sliding speeds, friction between the interacting

surfaces is a result of the bearing surface characteristics and properties of the

lubricant traces other than viscosity, such as metal–liquid adhesion energy. The

adhesion energy is expressed as surface wettability, the actual process in which a

liquid spreads on a solid substrate or material (Younes, 1993).

Quasi-harmonic oscillations are believed to be connected with specific

characteristics of the friction versus sliding speed curve, in particular, with the

negative slope of this curve, which can spread to the sliding speed of 2-3 m/s

depending on the materials, surface, and lubrication conditions. This type of

vibration is significantly affected by surface roughness; however, it can also occur

under any surface roughness conditions including “perfectly” smooth surfaces

(Kotousov 2009).

Roughness-induced vibrations are associated with surface roughness and asperities

on the contact surfaces (Kotousov 2009).

Many attempts have been made to investigate and analyse friction-induced vibrations in

water-lubricated bearings. For example, Simpson and Ibrahim (1996) conducted a series of

experiments to examine the mechanism for generating vibrations in water-lubricated

bearing systems. Their study suggested that the presence of water-lubricated bearing

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vibration “is associated with both the contact mechanics of the bearing and the dynamic

characteristics of the structural components of the bearing system” (Simpson and Ibrahim,

1996).

Simpson and Ibrahim (1996) considered several mechanisms which can give rise to

friction-induced vibrations in water-lubricated bearings. They stated that, “when one of the

sliding surfaces is characterised by a certain degree of elastic freedom, the motion may not

be continuous, but may be intermittent and proceed as a stick-slip process. During stick-

slip motion, two different deformation mechanisms take place. The first is elastic

deformation, where the two contact surfaces stick and the asperities deform elastically. The

second is plastic deformation, where sliding takes place and the asperities deform

plastically” (Simpson and Ibrahim, 1996). The authors concluded that “the occurrence of

stick-slip is unpredictable, mainly because the slope of the friction–speed curve is not

constant but varies randomly with contamination, surface finish, misalignment of sliding

surfaces and other factors” (Simpson and Ibrahim, 1996). They experimentally simulated

bearing dynamics and derived a linear analytical model for water-lubricated bearings.

Different dynamic characteristics were identified from the numerical simulation of the

equations of motion. However, the role of nonlinearity, due to the friction–speed curve,

was not considered (Simpson and Ibrahim, 1996).

Aronov et al. (1983) presented some experimental results of an investigation of the

interaction between friction, vibration, wear, and system rigidity. The results were obtained

from experimental investigation of a metal pin sliding on a steel disk under the action of

clean water as a lubricant. In this investigation, the load normal to the surface of the disk

was varied and the sliding speed was kept constant at 0.73 m/s. They found that severe

friction and wear are independent of system rigidity but dependent on the normal load.

However, it was also shown that mild wear rate increases with normal load and also with

system rigidity (Aronov et al., 1983).

2.4 Numerical models of friction in water-lubricated bearings

Frictional forces between interacting surfaces exist due to different and complex

mechanisms. Frictional forces can be responsible for undesirable dynamic characteristics.

Unfortunately, there is no appropriate method for determining or measuring friction,

vibration, and wear between sliding bodies under water lubrication. The modelling of the

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friction in mechanical systems depends on several factors that were classified by Ibrahim

(1994). These factors include the material properties, characteristics of the sliding surfaces

geometry, surface roughness and structure, normal load, sliding velocity, and temperature.

Usually, a water-lubricated bearing is designed to satisfy technical requirements such as

the power and/or the flow rate to be delivered, the torque required, and other performance

factors which are independent of friction, vibration and wear (Read and Flack, 1987).

During the design process, the tribological characteristics of the bearing system are either

considered last or not considered at all. The simplest form of analytical design is based on

short- and long- bearing approximations that are inaccurate in most practical design ranges

(Hirani et al., 1997). If any problems appear, either in the design stages or testing stages,

redesigning the entire bearing system is complex, highly difficult, and sometimes

impossible. The designers and users of water-lubricated bearings require reliable tools for

assessment and adjustment for the tribological characteristics of different bearing systems

(Read and Flack, 1987). Thus, numerical modelling of friction in water-lubricated bearings

would be a useful tool.

The numerical modelling and evaluation of the coefficient of friction, vibration, and wear

as functions of sliding speed have been the subject of numerous studies which considered

the influence of such factors as material properties, surface roughness, load, and type of

lubrication (Younes, 1993). Younes undertook investigation to “determine the dynamic

behaviour of a rigid balanced rotor with its journals rotating semi-dry in their bearings.

Their equation of motion, with the exponential coefficient of friction model, has been

solved numerically to determine the vibration responses”. Conditions for stability, the

increase of the system vibrations, and the planetary motion of the journal along the bearing

surface were reported for different design parameters (Younes, 1993).

A fast analytical method for evaluating design parameters such as load capacity, maximum

pressure, flow rate, power loss, and maximum temperature in the lubricant film of liquid-

lubricated bearings was developed by Hirani et al. (1997). According to Hirani (1997),

these parameters “are either too involved because of the mathematical complexities of

hydrodynamics or are based on simple methods and design charts. To predict the

tribological characteristics of bearings, an analytical model was proposed which provides

results comparable with time-consuming techniques such as thermo-hydrodynamic,

adiabatic, and isothermal analysis”. Finally, the design methodology was arranged in

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analytical expressions and tabular form for adequate and easy prediction of load capacity,

loss of power, flow rate, and maximum temperature. These approaches allowed to avoid

majority of numerical and mathematical complexities (Hirani et al., 1997). However, due

to the continuous development of rotating machinery, this approach is rarely adequate for

conditions involving contaminated lubricants because of the complexity of the friction

mechanism. Further development is required to enable adequate predictions and to make it

easy to use for all practical applications.

2.4.1 Linear models

Existing linear analytical models of water-lubricated bearings were reviewed by Simpson

and Ibrahim (1996), page 90. They stated that “different dynamic characteristics are

predicted from the numerical simulation of the equations of motion. Several mechanisms

can give rise to friction-induced vibration in sliding systems”. For example, in one-degree-

of-freedom systems, where contacting surfaces interact against each other in the one

direction, a necessary condition that contributes to friction-induced vibration is the

negative slope of the friction–speed curve (as shown in Figure 2.6) (Simpson and Ibrahim,

1996).

As was shown by Kotousov (2009) and Tworzydlo et al. (1994), in water-lubricated

bearings, the total friction at contact consists of two components: one is due to the asperity

contact friction, Fas, and the second is due to hydrodynamic viscous shear friction, Fv. The

relative importance of the two parts depends on the relative magnitude of the film

thickness between sliding surfaces and the surface roughness. According to Kotousov

(2009, p.8), the friction due to asperity contact is calculated as follows:

PA

AfF

t

aasas (2.1)

where:

fas is the asperity sliding friction coefficient;

Aa is the actual area of contact;

At is the total area of contact when the film thickness is zero;

P is the bearing pressure.

The viscous shear friction force is:

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0 1 2 3 4

10-3

10-2

10-1

1

Sliding speed, m/s

Friction coefficient

Negative slope

Positive slope

t

at

min

ovA2

AA2c2

h

VF

(2.2)

where:

o is the viscosity of the lubricant;

minh is the minimum film thickness;

c2 is the contact dimension;

V is the sliding speed.

Dividing by the bearing pressure P and summing the Equations 2.1 and 2.2, the total

friction coefficient is given by:

ft

aas

A

Af +

t

at

min

oA2

AA2c2

Ph

V . (2.3)

According to Pan et al. (1971), viscosity and speed always appear as a product in the

elastohydrodynamic theory of lubrication, and the lift-off sliding speed is roughly

proportional to the viscosity. The speed, at which full film separation between shaft and

bearing takes place, decreases with decreasing temperature.

Based on the elastohydrodynamic theory of lubrication, Pan et al. (1971) established

representative baseline calculations, showing the dependence of the friction coefficient on

the shaft sliding speed as shown in Figure 2.9. The friction coefficient increases with the

effective surface roughness (or asperity size).

Figure 2.9 Typical baseline friction–speed curve (Pan et al., 1971)

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According to Figure 2.9, the boundary and mixed regimes of lubrication for typical water-

lubricated bearings are within the sliding speed range up to ~2 m/s (which is in agreement

with the “Stribeck” model) (Pan et al., 1971).

2.4.2 Non-linear models

Friction forces between sliding surfaces arise due to a complex mechanism, and lead to

numerical models which are strongly non-linear, discontinuous and non-smooth (Qiao and

Ibrahim, 1999). The inclusion of non-linearity in the equations of motion of a dynamic

system leads to differential inclusions in the mathematical model, adding further difficulty

to the problem (Qiao and Ibrahim, 1999, EOV, 2004). Many researchers have conducted

numerical studies of friction, vibration, and wear in water-lubricated bearings (Simpson and

Ibrahim, 1996, Qiao and Ibrahim, 1999, Mottershead et al., 1997). Their general

conclusions are:

1. Most of the friction-induced vibrations takes place at low sliding speeds (typically

up to 2.0 m/s) where the slope of the friction coefficient with respect to the sliding

speed is negative (boundary-mixed regimes).

2. The variation of the friction force, vibrations, and wear decrease as the speed of the

shaft increases to a critical speed above which the friction coefficient begins to

increase due to viscous shear (hydrodynamic regime).

The system model of a stern-tube bearing was considered by Simpson at al. (1996) who

developed an analytical nonlinear two-degree-of-freedom model for the water-lubricated

journal bearing, as shown in Figure 2.10. The mass, stiffness, and damping coefficient of

the tangential shear of the water-lubricated bearing are indicated by 1m , 1k and 1c

respectively. The constants ( 2k and 2c ) represent the stiffness and damping of the flexible

shaft which drives the disk of mass ( 2m ) and of the mass moment of inertia ( I ) about the

shaft axis.

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Figure 2.10 Analytical two-degree model representing a submarine aft water-

lubricated bearing, as displayed in Simpson and Ibrahim (1996), p. 90, Figure 2

The non-linear response of this system which emulated the dynamics of a submarine

water-lubricated bearing shaft was investigated numerically. “The influence of a time

variation of the sliding speed results in a time variation of the friction force. This time

variation of the friction force is found to be responsible for the occurrence of vibration and

noise” (Simpson and Ibrahim, 1996).

2.5 Vibration–wear relationship in dynamic systems

In spite of the need to correct prediction and design of the different dynamic systems

(including water-lubricated bearings) and the increased interest of designers and engineers,

the problem of the vibration–wear relationship in different dynamic systems has not

received adequate attention from researchers. Vibration in different dynamic systems

(machines, bearings) is generally due to the dynamic friction forces as discussed by

Krishna Kumar et al. (1997), and the vibration–wear relationship depends on many factors.

These factors include contact materials; operational parameters (load, sliding speed, type

of lubrication and lubricant contamination); and the characteristics of the dynamic system

(natural frequencies and inertia of fixtures and components) (Krishna Kumar and

Swarnamani, 1997).

While the vibration–wear relationship in most dynamic systems is often unknown, the

assessment of wear of the contacting surfaces of many dynamic systems, such as bearings

and sliders, by vibration monitoring using various shock pulse measurements, methods and

tools, has been employed by many researchers and industries (Jonson, 2000).

V

Bearing Disk

m1 m2, I

k2 k1

c2 c1

Friction force

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A reliable and inexpensive methodology is required to simulate and assess vibration–wear

dependency for a physical water-lubricated bearing.

2.5.1 Problem of vibration–wear dependency

The influence of friction-induced vibrations on the wear that occurs in many mechanical

systems (brakes, bearings, wheel-rail contact, etc.) can lead to higher wear and the

formation of unwanted features on the contacting surfaces (corrugations, micro-cracks,

wash boarding) as well as noise and more vibration (Popov, 2010).

According to Chowdhury (2009), it was observed by several authors (Kato et al., 1982,

Maru et al., 2005, Maru et al., 2007b, Bryant and York, 2000) that the wear reduction

depends on interfacial conditions such as normal load, geometry, relative surface motion,

sliding speed, surface roughness of the contact surfaces, type of interacting materials,

system rigidity, temperature, humidity, type of lubrication, vibration and lubricant

contamination (Chowdhury and Helali, 2009).

Several studies have been conducted in which researchers have found that wear can be

reduced by vibration. Goto and Ashida (1984) found that ultrasonic vibration can

significantly reduce wear rates on Pin-on Disk (POD) experimental apparatus which uses a

steel pair of interacting materials. Bryant and York (2000) showed that micro-vibrations

(10-100 mm amplitude, 10-100 Hz) of a slider can reduce sliding wear by up to 50%,

particularly for rigid body rocking vibration. Moriwaki and Shamoto (1991) used

ultrasonic vibrations to reduce tool wear without damaging the surface finish. Kato et al.

(1982) also found that vibrations sometimes increased and sometimes decreased wear

rates, depending on the pairs of materials involved. In their experimental investigation,

Weber et al. (1984) used ultrasonic vibrations to extend carbide tool life in the machining

of glass (Chowdhury and Helali, 2009).

Chowdhury and Helali (2007) also considered the lack of correlation between wear rate

and other vibration-related operating parameters during an investigation of the wear

behaviour of mild steel under vertical vibration. Initially, the aim of their research was to

find a suitable correlation and a way of reducing wear rate, by applying a known,

controlled frequency and amplitude of vibration in a particular direction (Chowdhury and

Helali, 2009). A POD experimental test rig was used and results showed that the wear rate

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at a particular amplitude decreases with increasing frequency of vibration. The authors also

found a reduction of the friction coefficient as a function of different amplitude and

frequency of vibration on a pair of materials comprised of mild steel. However, the wear

behaviour of mild steel and their dimensional analysis in relation to both frequency and

amplitude under vertical vibration required further investigation. It is expected that the

application of these results will contribute to the improvement of the performance of

different sliding mechanical systems (Chowdhury and Helali, 2009).

2.5.2 Effect of damping on vibration and wear

During the design stage of rotating machines, wear and vibration in bearings are significant

problems for designers. Several investigators have been interested in the damping

phenomenon. In previous studies on wear initiation and development, theoretical analyses

predicted that contact damping is the most effective factor for preventing wear (Suda and

Komine, 1996). It is known that increased damping reduces contact vibration, resulting

from surface roughness, although the mechanism may be different (Suda and Komine,

1996).

A few methods exist for contact damping to improve bearing performance. These methods

include (Sutherland, 2002):

Structural damping (which refers to energy dissipation within the structure by add-

on damping devices such as an isolator, by structural joints and supports, or by

structural member's internal damping);

Self-balancing damping (when the vibrations of rotors can be reduced or eliminated

by adding self-balancing mechanical components);

Liquid damping (when the vibrations of rotors can be reduced or eliminated by

adding self-balancing liquid);

Speed control damping (when the vibrations of rotors can be reduced or eliminated

by a change of rotation speed);

Magnetic damping (when the vibrations of rotors can be reduced or eliminated by a

magnetic field).

Structural damping is not always sufficient to limit vibration and wear to within the desired

level. In these cases, magnetic damping may provide a solution.

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Padmanabhan (1992) emphasised the importance of damping at joints and experimentally

investigated the effect of damping on different tribological characteristics (such as friction

and wear) of interacting surfaces. A mathematical equation was formulated to predict the

effect of damping on lubricated and non-lubricated joints.

Verichev at al. (2010) outlined the principles for damping lateral vibrations of rotary

systems and undertook an experimental study to investigate the effect of speed control

damping on lateral vibrations in rotating machinery using motor speed modulation. This

method was based on the generation of a harmonic additive to the constant speed of

rotation that provided significant damping of lateral vibrations at critical rotation speeds.

The analytical solution and numerical calculations proved this concept and showed a

significant decrease in the amplitudes of lateral vibrations compared to those in a similar

undamped system (Verichev, 2010).

Kasadra et al. (2004) considered rotor instabilities in rotating machinery due to re-

excitation of the first rotor’s critical speed (when a rotor’s radian frequency is in a state of

resonance with a first natural frequency) resulting in lateral rotor vibrations at frequencies

below the rotor’s operating frequencies. They found that active magnetic dampers were

very promising for reducing or even eliminating rotor vibrations regardless of the

excitation source. Experimentally, it was shown that an active magnetic damper was used

to effectively add damping to reduce a rotor’s vibration response. The experimental results

also demonstrated that the active magnetic damper can significantly dissipate the

vibrational energy and prevent system dynamic characteristics such as natural frequency,

which may lead to an increase in rotor vibrations (Kasarda, 2004).

The analysis of previous experimental investigations showed that magnetic damping is a

promising technology for reducing vibrations in water-lubricated bearings. However, it

may or may not be the best choice because the effect of damping on the vibration–wear

relationship is relatively untested in accepting the overall tribological characteristics and

needs to be addressed to ensure the effective use of active magnetic damping.

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2.6 Concepts of experimental apparatus and experimental technique

The problem of predicting the performance of water-lubricated bearings has led to a need

for further development of a reliable and inexpensive experimental methodology to

simulate a real water-lubricated bearing in an actual operational environment

(WÄRTSILÄ, 6/09/2007).

2.6.1 Pin-on-Disk system

In order to conduct an experimental investigation of friction, vibration, and wear in water-

lubricated bearings, many experimental systems, including a Pin-on-Disk (POD)

experimental system, have been considered. The POD model has been successful for

similar experimental studies (Dweib and D'Souza, 1990, Aronov et al., 1983, Tworzydlo et

al., 1994, Ibrahim, 1994a, Ibrahim, 1994b, Mosleh et al., 2002, Qiao and Ibrahim, 1999,

Tworzydlo et al., 1999).

Experimental tests using a POD-type sliding system have indicated that the friction force,

vibration, and wear depend mostly on the normal load for a constant sliding speed (Qiao

and Ibrahim, 1999). According to Qiao (1999), depending on the value of the normal load,

four different friction regimes were observed:

1. The steady-state friction regime where the frictional force increases linearly with

the normal load.

2. The non-linear friction regime where the friction force increases non-linearly with

the normal load and the coefficient of friction is no longer constant but increases

with the normal load.

3. The transient regime, characterised by intermittent variation of the friction force.

When the friction force reaches a sufficiently high value, a temporary burst of self-

excited vibrations occur and the friction force falls to a lower value.

4. The self-excited vibration regime where the friction force drops to a low value and

is accompanied by high-amplitude periodic self-excited oscillations.

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Chowdhary and Helali (2009) experimentally investigated a vertical vibration–wear

relationship using a POD test rig. The pair of materials used was mild steel. They found

that the wear rate for mild steel decreased with an increase in the amplitude and frequency

of the vertical vibrations. Therefore, maintaining a certain frequency and amplitude of

vertical vibration may keep wear to some lower value, which would improve the

tribological characteristics of different mechanical systems. Moreover, it was considered

that a lack of knowledge and available information on the vibration–wear relationship in

different mechanical systems existed. A lack of a suitable correlation method for reducing

wear rate by applying, or damping, known frequencies and amplitudes of vibration in a

particular direction also existed (Chowdhury and Helali, 2009).

2.6.2 Effect of lubricant contamination

Another problem which must be addressed during bearing design and operation is lubricant

contamination (Solomonov et al., 2010).

According to Maru et al. (2007a), in water-lubricated bearings, contamination of the

lubricant by wear debris or solid particles is one of the main reasons for early bearing

failure. In order to deal with this problem, it is important not only to use reliable techniques

for detection and removal of solid contaminants, but also to investigate the effects of

certain contaminant characteristics on bearing performance (Maru et al., 2007a).

Maru et al. (2007a and 2007b) presented the results of an investigation into the effect of

lubricant contamination by solid particles on the dynamic behaviour of bearings to

determine the trends in the amount of vibration induced by contamination of the oil and by

the bearing wear itself. Experimental tests were performed with radial ball bearings

lubricated in an oil bath. Quartz powder at three concentration levels and in different

particle sizes was used to contaminate the oil. Vibration signals were analysed in terms of

the root mean square (RMS) values. The results showed that changes in the RMS values of

vibration in the high-frequency band, from 600 to 10,000 Hz, were associated with changes

in the oil lubrication of the bearing as a result of contamination and wear damage to the

bearing surfaces. It was shown that the effect of contaminant concentration on vibration

was distinct from that of particle size. The vibration level increased with the percentage of

concentration, tending to stabilise at a particular limit. On the other hand, as the particle

size increased, the vibration level first increased and then decreased. Vibration level

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increased during the test in contaminated oil only after 16 minutes of testing. Such an

increase in vibration was related to an effect produced by the wear of the bearing elements.

The bearing surfaces were reported to be severely damaged by a three-body abrasive

mechanism distributed along all the surfaces. Abrasion was also identified through

ferrography, a microscopic technique for analysing the particles present in fluids that

indicate mechanical wear. This indicated a severe wear regime, although measurements of

internal radial clearance of the bearings are reported to have shown an absence of

dimensional wear. The amount of vibration due to bearing wear was dependent on the

contamination level. A correlation was observed between the trends of the wear due to

bearing vibration and those of its overall surface damage. The vibration due to the presence

of particles was shown to be proportional to the vibration of the worn bearing as particle

concentration increased. On the other hand, when the contaminant particle size increased,

the dynamic action of the particles passing through the contact interface increased, but the

vibration level of the worn bearing was the same. The vibration due to the larger particles

is reported to be reduced due to the particle settling phenomenon (the decrease in vibration

with larger particles suggested that it was more difficult for larger particles to go into the

contact interface and, therefore, the vibration of the respective worn bearing was reduced)

(Maru et al., 2007a, Maru et al., 2005, Maru et al., 2007b).

2.7 Summary and research gaps

This literature review has presented that theoretical and experimental studies of friction,

vibration, wear, vibration-wear relationship, and bearings lubrication, which have been of

interest to researchers for various mechanical applications. However, existing theoretical

and experimental tribological models, presented in the literature, are mostly linear, simple

in nature, and do not consider damping, and contamination of the lubricant. Selection of

water-lubricated bearings materials generally depends on the designers’ previous

experience and is simplified. The relationship between various water-lubricated bearing

parameters, lubrication regimes (hydrodynamic, mixed film or boundary), lubricant

contamination in addition to the influence of water contamination and damping on friction,

friction-induced vibration, wear and the vibration–wear relationship on water-lubricated

bearings materials performance is very limited and requires further investigation.

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CHAPTER 2 Literature review

37

Severe contamination in water-lubricated bearings can cause excessive wear of

components, surface damage, fatigue failure, and noise. The effect of water contamination

on friction, vibration, wear and the vibration–wear relationship has not previously been

considered. It needs to be further investigated in order to predict the tribological behaviour

of water-lubricated bearings materials during operation. It remains a matter of serious

concern that no research-based methods and models have been designed for its evaluation.

A Pin-on-Disk (POD) experimental apparatus has been chosen to investigate the effect of

water contamination and damping on friction, vibration, wear behaviour and the vibration–

wear relationship in boundary and mixed regimes for water-lubricated bearings materials.

It is essential to qualify the effect of water contamination on the water-lubricated bearing

materials performance for developing design guidelines. At present, there are not design

guidelines available for water-lubricated bearing materials selection which experience

water contamination. It is therefore expected that this project will contribute to the

development of industry guidelines, and provide a better understanding and modelling

approach for the tribological process between materials used for water-lubricated systems.

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CHAPTER 3 Previous experimental studies of tribological characteristics in water-lubricated bearings

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CHAPTER 3 PREVIOUS EXPERIMENTAL STUDIES OF

TRIBOLOGICAL CHARACTERISTICS IN WATER-

LUBRICATED BEARINGS

3.1 Introduction

This chapter represents the experimental results and analysis of a wear study which was

conducted by Wärtsilä Pty Ltd, a UK-based supplier of water-lubricated bearing materials

(Biswell, 2007, Cumberlidge, 2009, WÄRTSILÄ, 6/09/2007). It is followed by an aft

bearing study on the effect of water contamination on a water-lubricated bearing system

that was previously undertaken by researchers in the School of Mechanical Engineering,

University of Adelaide (Kotousov, 2009). These previous studies became a starting point

for the experimental investigations undertaken and results presented in the following

chapters of this thesis.

The aft bearing of ships and submarines is most commonly a plain journal bearing which is

lubricated and cooled with sea water. The relationship between operational parameters and

the system, the bearing lubrication regimes (hydrodynamic, mixed [transition] or

boundary) and the dynamic motions associated with the aft bearing in operation are largely

unknown (Biswell, 2007, Cumberlidge, 2009, WÄRTSILÄ, 6/09/2007).

The main objective of the studies undertaken previously was to investigate the operational

characteristics of the aft bearing assembly as a part of the propulsion shaft system under

water contamination conditions. These studies provide an understanding of the current aft

bearing design requirements and provided a framework for future investigations on similar

systems (Solomonov, 2009, Solomonov et al., 2010).

3.2 Review of wear study

A comprehensive wear study was conducted by Wärtsilä Pty Ltd, a UK-based supplier of

water-lubricated bearing materials (WÄRTSILÄ, 6/09/2007, Biswell, 2007). The aim of

this study was “to analyse the performance of various bearing materials in highly-abrasive

conditions against stainless steel counterface material” (Biswell, 2007, Cumberlidge, 2009,

WÄRTSILÄ, 6/09/2007). According to Biswell (2007), “Substitute” sea water was used,

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CHAPTER 3 Previous experimental studies of tribological characteristics in water-lubricated bearings

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which had silica particles added. The grit used was equivalent in particle size and shape to

that found in the Portland area of the UK, at concentration accepted by the UK Ministry of

Defence as representing of aggressive British coastal water. To simulate the worst water

conditions, and to accelerate the test, the concentration of silica was increased by a factor

of 10. The grit was kept in suspension in the sea water by a stirrer agitating the mixed

solution in the contaminated water tank. A pump was used to deliver the contaminated sea

water to and from the bearing. The flow rate was set at 7.5 litres per minute” (Biswell,

2007, Cumberlidge, 2009, WÄRTSILÄ, 6/09/2007). Table 3.1 details the test parameters.

Table 3.1 Basic parameters for the Wärtsilä wear tests (Biswell, 2007)

“The initial testing comprised running each material under the adopted conditions for a

duration of 100 hours, where the wear rate was measured at 20-hour intervals” (Biswell,

2007, Cumberlidge, 2009, WÄRTSILÄ, 6/09/2007). To keep experimental results

consistent, all bearings were tested with the same geometry and configuration. This was

done to ensure that the performance and not the design of the bearing material was tested.

The results from the experimental wear tests for the different bearing materials are shown

in Figure 3.1.

Parameters Metric Units

Sleeve: Stainless steel

Sleeve diameter

ENISO 316

50.8 mm

Shaft rotation

Bearing load

Bearing pressure

8.8 m/min (55 rpm)

2500 N

4.8 kg/cm2

Pressure velocity (PV) rating

Lubricant flow rate

42 kg/cm2.m/min

7.5 litres/min

Lubricant tank capacity 88 litres, agitated

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CHAPTER 3 Previous experimental studies of tribological characteristics in water-lubricated bearings

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Figure 3.1 Experimental wear rates for water-lubricated bearing materials (Biswell,

2007, Cumberlidge, 2009, WÄRTSILÄ, 6/09/2007)

As can be seen from the Figure 3.1, although all the materials were tested under the same

conditions, a difference in wear performance noticeable. Most materials performed well

during this period of time, except the elastomeric material, which showed significant

bearing wear. Smearing of the bearing material on the shaft and scoring on the shaft liner

were noted (Biswell, 2007, Cumberlidge, 2009, WÄRTSILÄ, 6/09/2007).

3.3 Review of the aft bearing study

According to Kotousov (2009, p.12), the physical (scaled test rig) modeling approach was

used by researchers from the School of Mechanical Engineering, at the University of

Adelaide, to simulate a water-lubricated aft bearing system. The friction conditions of the

scaled physical model (test rig) were simulated as close as possible to the actual system.

This is to ensure the similarity of the dynamic behaviour between the scaled model and

actual aft bearing systems behaviour.

3.3.1 Experimental apparatus

The test rigs, presented in Figure 3.2, were built to investigate the effect of water

contamination by applying the following aft bearing system design control parameters

(Kotousov, 2009):

Lubrication contamination;

Radial clearance;

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CHAPTER 3 Previous experimental studies of tribological characteristics in water-lubricated bearings

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Sliding speed;

Bearing pressure;

Coolant flow.

A total of three test rigs were manufactured and two types of lubrication conditions were

used: clean tap water, and contaminated tap water with added sand particles to simulate

contamination. The water was maintained at room temperature.

According to Kotousov (2009), “the contaminated conditions were not specified but the

adding of sand particles seemed to be an appropriate way to create the contaminated

environment and investigate the qualitative response of the system to the contamination”.

The duration of the long-term testing was approximately four months. It allowed the

achievement of a similar wear pattern as in an operational aft bearing system. This time

was sufficient to achieve a steady-state wear rate and to enable theoretical methods to be

applied to extrapolate the wear process for longer time periods.

Figure 3.2 Scaled test rig and major components, as displayed in Kotousov (2009), p.

13, Figure 3.2.1

The first test rig simulated the long-term operation of the aft bearing system in clean water

lubrication conditions. The second test rig aimed to simulate the long-term operation of the

aft bearing system in a contaminated water regime. The third test rig was used to

investigate the effects of alignment, lubrication contamination, radial clearance, sliding

Shaft

Assembly

Sleeve Housing

Assembly

Load Cell

Arch Support

Mounting Base

Variable Speed Motor

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CHAPTER 3 Previous experimental studies of tribological characteristics in water-lubricated bearings

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speed, and roughness of the sliding surfaces on the tribological behaviour of the whole

system.

3.3.2 Experimental results of the friction coefficient measurements

The value of the friction coefficient in the system was investigated using a radial clearance

of 1 mm and high flow rates. Smooth, quiet and oscillation-free regimes were easily

achievable for large radial clearance.

Measurements of friction were conducted using strain gauges mounted on the vertical arm

supporting the sleeve housing. Two sets of measurements were taken at different points on

the arm to reduce errors, and also to demonstrate the consistency of the measurement

technique.

Table 3.2 provides a summary of the parameters of the bearing system used to estimate the

thermal regime and power losses due to friction and vibrations.

Table 3.2 Basic parameters of the bearing system (Kotousov, 2009, Solomonov et al.,

2010)

Figure 3.3 presents the results of the experimental investigation of the coefficient of

friction against the sliding speed/shaft rotation speed under clean water lubrication.

Parameters Experimental Rig

Bearing pressure 0-0.6 MPa

Actual flow rate 0-1.0 l/s

Angular speed range 200-3000 rad/s

Shaft diameter 30 mm

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CHAPTER 3 Previous experimental studies of tribological characteristics in water-lubricated bearings

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Figure 3.3 Friction curves vs. sliding speed, m/s (rotation speed, rpm), reproduced

from Kotousov (2009, p.18)

These results indicated that the friction coefficient is significantly affected by the sliding

speed. It changes from 0.03 at low speeds to 0.002 at high speeds of rotation (high sliding

speeds). It is interesting to note that these tendencies are in agreement with the theoretical

predictions made by Pan et al. (1971), especially for high sliding speeds. This boosts

confidence in the experimental technique adopted in this study. These results provide

information for the determination of the operating and temperature conditions as well as

for estimating the power losses in the oscillation-free regime.

Kotousov (2009) reported that the level of water contamination in the aft bearing system

was unknown and the results presented were aimed at providing a qualitative assessment of

the effect of the contaminated lubricant on friction, vibration, and noise characteristics of

the water-lubricated bearing system. Moreover, the water contamination was not constant

with time and it was extremely difficult to control and characterise the level of

contamination.

0

0.01

0.02

0.03

0 500 1000 1500

Trend line

700

Shaft rotation frequency, rpm

0 0.79 1.6 2.4 1.1

Sliding velocity, m/s

Elastohydrodynamic theory of lubrication, Pan et

al. (1971)

Fri

ctio

n c

oef

fici

ent

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CHAPTER 3 Previous experimental studies of tribological characteristics in water-lubricated bearings

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According to Kotousov (2009), the difference in the friction coefficient between the rig

with contaminated lubricant and the rig with clean lubricant operating in the mixed

lubrication regime was reported to be an order of magnitude. At the same time, the

difference between friction coefficients of the two test rigs operating within the boundary

regime is small, reaching 30-40%. Therefore, for a bearing intended to work in mixed and

hydrodynamic regimes, contamination of the lubricant is one of the major issues. However,

if the bearing is intended to operate in the boundary regime, the level of water

contamination does not appear to exert significant influence on the friction properties of

the system.

3.3.3 Experimental results of the wear tests

The rig for long-run investigations was operated in contaminated and clean lubrication

conditions at a sliding speed V ~1.57 m/s and bearing pressure P = 0.47 MPa. Periodic

measurements of the wear of the bearing shell and shaft were recorded. These results

demonstrated that wear occurred in the bottom part of the shell (Kotousov, 2009).

The wear rate due to water contamination was found to be five times higher than that for

the clean water. The friction in the clean lubricant environment left a smooth shiny surface,

whereas in the contaminated environment, the shell showed signs of heavy damage such as

deep footprints of abrasive damage.

3.3.4 Discussions and conclusions of the experimental tests

The effect of water contamination on values of the friction coefficient and wear were

experimentally investigated and the following conclusions were drawn by Kotousov

(2009):

1. The characteristics of water-lubricated bearings depend on the operational

conditions of the aft bearing system as well as the material properties. For boundary

and mixed lubrication regimes, complex design considerations need to be

undertaken such as: careful selection of bearing materials, filtering the cooling

water, proper alignment and smooth contact surfaces. None of these steps on its

own is able to significantly improve the water-lubricated bearings’ performance. If

all steps are implemented, it seems that the rate of water cooling is not a critical

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CHAPTER 3 Previous experimental studies of tribological characteristics in water-lubricated bearings

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factor as, at very low friction conditions; the thermal regime does not require higher

flow rates than are currently used in the aft bearing system. However, this set of

improvements will work if one maintains the operational rotational speed of the aft

bearing above 40 rpm (sliding speed ~0.06 m/s). Below this critical speed of

rotation, the friction versus sliding speed curve for the tested materials has a large

negative slope which will lead and feed the friction-induced vibrations and

excessive wear. Consequently, additional measures are required if the bearing is

intended to operate below 40 rpm.

2. For the boundary regime of operation, structural modifications to reduce the effect

of water contamination on the bearing assembly are required. In this case, the flow

rates should be much higher than those currently used to avoid overheating and

reduce the wear rates. The water-lubricated bearing system would also benefit from

filtering of the lubricant. This will significantly reduce the wear rates and keep the

surface in good condition.

3. Other recommendations can be directed to the redesign of water-lubricated bearing

systems and the use of new solutions, such as:

a separate cooling system to keep the temperature of the lubricant low (low

temperature of the lubricant increases its viscosity),

new composite materials with low friction properties, specifically at the

required low operating speeds,

design of the additional lift pressure system to create a film layer between the

shaft and shell, etc.

3.4 Conclusions

The aim of the experimental studies (Biswell, 2007, Cumberlidge, 2009, Kotousov, 2009,

WÄRTSILÄ, 6/09/2007) was to provide a series of scaled model tests demonstrating the

long-term running and wear characteristics and the tribological behaviour of water-

lubricated bearings under clean and contaminated water-lubricated regimes. The test rigs

simulated the lateral loading associated with the aft bearing system of a drive shaft used in

large boats and ships.

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CHAPTER 3 Previous experimental studies of tribological characteristics in water-lubricated bearings

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Their primary goal was to increase the efficiency and maintenance intervals of water-

lubricated bearing systems.

The experimental programs (Biswell, 2007, Cumberlidge, 2009, Kotousov, 2009,

WÄRTSILÄ, 6/09/2007) were undertaken to identify and provide understanding of the

general problems of the aft bearing system currently used in ship and submarine

applications. However, the fundamental sources of those problems, such as the effects of

water contamination and damping on the tribological behaviour of water lubricated

bearings materials, were not investigated due to the limitations of the scaled model of

water lubricated bearing test rig capability.

Further extensive experimental and theoretical studies are therefore required to properly

investigate and understand the tribological behaviour and the physical consequences of a

variation in the material properties of composite materials used for water-lubricated

bearings.

The control of friction, wear, and vibrations in moving machine parts is a major problem

for many manufacturers and designers. It has become crucial to have quantitative data

obtained at varying operational conditions and in the presence of lubrication and

contamination (AIphaISS, 2009). POD experimental tribometers have proven their

reliability in many laboratories worldwide (AIphaISS, 2009, Dweib and D'Souza, 1990,

Aronov et al., 1983, Tworzydlo et al., 1994, Ibrahim, 1994a, Ibrahim, 1994b, Mosleh et al.,

2002, Qiao and Ibrahim, 1999, Tworzydlo et al., 1999, Chowdhury and Helali, 2007) for

studying new materials (ceramics, metals, polymers, composites); lubricants and

contaminant additives; self-lubricating systems; and quality assurance.

Compared to the scaled model of water-lubricated bearing test rig, a Pin-on-Disk

experimental apparatus can provide the following features (AIphaISS, 2009):

Precisely calibrated friction, vibration, and wear measurements

A pin is loaded onto the test disk with a more precisely known force

Stable contact point and no parasitic friction

Variable sample materials, sizes, and shapes

Tests conducted in dry, liquid-lubricated, and controlled contamination

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PC-controlled data acquisition and instrument control, featuring real-time graphical

data display, friction coefficients, wear and vibration data, sliding lifetime, etc.

Based on the analysis of the previous experimental approaches and, together with the

conclusions outlined in within this chapter, the further POD experimental program was

developed and undertaken to investigate the effect of water contamination on the friction

coefficient, vibration, wear and the vibration-wear relationship under varying operational

conditions (Solomonov, 2009, Solomonov et al., 2010). The results and the discussion of

the POD experimental study are presented in the following chapters of this thesis.

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CHAPTER 4 Experimental apparatus

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CHAPTER 4 EXPERIMENTAL APPARATUS

4.1 Introduction

A thorough understanding of the tribological behaviour of water-lubricated contact

surfaces during the design and development of new engineering applications is vital.

According to the literature review, the boundary and mixed regimes cause many

tribological problems. These problems can include vibration, power loss and excessive

wear. The simulation of friction in boundary and mixed regimes is difficult. To obtain a

good prediction of the overall performance of water-lubricated bearings, the precise

measurements of friction, vibration, and wear in the boundary and mixed lubrication

regimes are critical. The friction coefficient, vibration, and specific wear rate for a range of

sliding speeds, applied normal loads, and contaminated lubricants need to be collected and

analysed. In order to use this data in simulations and obtain a better understanding of the

tribological behaviour of sliding bodies, it is preferable to conduct these measurements on

small test samples where different operational conditions can be applied.

This chapter presents the test rig used for the experimental study of friction. Only the main

elements of the test rig and data acquisition methods are discussed.

4.2 Design requirements

A measurement method based on the use of a Pin-on-Disk (POD) technique can fulfil the

previously-mentioned demands such as the precise investigation of the dependence of

friction, vibration, and wear on sliding speeds, applied loads, water contamination, etc.

This method can yield a better understanding of tribological behaviour under water-

lubricated conditions (Tworzydlo et al., 1999, Godfrey, 1995, Marklund and Larsson,

2008).

POD tribometer is able to precisely measure the magnitude of friction, vibration, and wear

between two interacting surfaces (Dweib and D'Souza, 1990, Aronov et al., 1983,

Tworzydlo et al., 1994, Ibrahim, 1994a, Ibrahim, 1994b, Mosleh et al., 2002, Qiao and

Ibrahim, 1999, Tworzydlo et al., 1999, Chowdhury and Helali, 2007, CSM-Instriments,

2013). In one measurement, a flat or a spherical sample (pin) is placed on the rotating test

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CHAPTER 4 Experimental apparatus

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disk and loaded with a precisely known weight. The resulting frictional forces and

vibrations between the pin and the disk are measured and recorded. Additionally, the wear

of the sample is measured and calculated from the volume of the material lost during the

test (CSM-Instruments, 2013).

General requirements for a POD test rig are as following (CSM-Instriments, 2013):

High resolution obtained with unique frictional force measurement system design

Easy calibration processes and procedures

High-precision feedback controlled motor

Easy sample replacement and adjustment

Environmental (operational) configuration (dry, lubrication etc.)

Tests in dry and lubricated regimes, controlled contamination, and damping

Tribological investigation in accordance with the ASTM G 99-04 Standard test

Method for wear testing with a pin-on-disk apparatus was used to collect wear data

(ASTM, G 99 – 04).

A special POD test rig was designed and built to enable the precise measurements of

friction, vibration and wear for water-lubricated bearings materials. This POD test rig has

the capacity to be used for the study of friction, vibration and wear behaviour of any solid

combination of materials, for varying times, applied loads, sliding speeds, damping,

lubricants, contaminants, and fluids.

In POD tests, a load cell with a pin is held by an arm and loaded axially so that the pin is in

contact with a rotating disk, as shown in the schematic diagram in Figure 4.1.

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CHAPTER 4 Experimental apparatus

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1. Arm

2. Disk

3. Load cell

4. Pin

5. Strain gauge sensor

6. Displacement gauge

sensor

Figure 4.1 Schematic diagram of a Pin-on-Disk experimental apparatus

The initial stages of this project involved the design, building, and commissioning of a

POD test rig to study the effect of contaminated lubrication on friction, vibration, and wear

with variable sliding speeds and loads. The test rig needed to be robust and reliable with all

major control parameters such as sliding speed, applied normal load, water lubrication and

contamination, and friction force monitored and recorded using a computer-based data

acquisition device. Appropriate samples of available materials, operational conditions

including a range of sliding speeds, applied loads and levels of water contamination needed

to be chosen before commencement of the design process.

In accordance with the deficiencies discussed in Chapter 3, design requirements were

selected. The major design requirements are shown in Table 4.1.

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CHAPTER 4 Experimental apparatus

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Table 4.1 Design requirements for the Pin-on-Disk test rig

Requirements Range

Load force, N 8-50

Coefficient of friction, µ 0.1-1

Sliding speed, m/s 0.3-4.5

Lubrication regimes: Dry

Clean water

Contaminated water

(with 1%, 2%, 4% and 6% levels of

contamination)

Pin diameter, m 0.01

Disk diameter, m 0.3

Variable speed motor available BALDOR: MM3550C-57

Adjustable frequency drive GENESIS: NEMA-4X/IP-65

4.3 Variable sliding speed and applied load

4.3.1 Variable sliding speed

Employing a range of sliding speeds during the experimental work was essential for

obtaining friction–speed, vibration–speed and specific wear rate–speed curves.

Adjustments were made using a variable-speed drive system. The main purpose of the

variable-speed drive system was to enable the rotating disk to rotate at an appropriate

speed during the experiments. The variable drive system consisted of a three-phase AC

motor (BALDOR: MM3550C-57) with an adjustable frequency drive (GENESIS: NEMA-

4X/IP-65). The variable drive system mounted on the test rig is shown in Figure 4.2.

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CHAPTER 4 Experimental apparatus

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a) Motor b) Frequency drive

Figure 4.2 Variable drive system: a) 3-phase BALDOR: MM3550C-57 motor with b)

GENESIS: NEMA-4X/IP-65 adjustable frequency drive

The AC motor had the following major parameters:

230/415 volt/50 Hz/3 phase

Fully reversible

Rolled steel stator frame

Met the requirements of Australian Standard AS1359, including ‘MEPS Minimum

Efficiency’.

The AC motor was equipped with the programmable, adjustable frequency drive

GENESIS: NEMA-4X/IP-65 to set the appropriate rpm speed.

The adjustable frequency drive (NEMA-4X/IP-65) had the following major parameters:

Multifunction keypad with 4-digit LED display

Simplified group programming

8 LED status indicator

Rated for 208-230/400-460 volt/50 and 60 Hz.

To obtain an appropriate sliding speed on the POD test rig, the AC motor was mounted on

the same base as the test rig and connected to the disk assembly using a pulley system. The

disk to motor pulley ratio was 13:1. This enabled a sliding speed range of 0.3 - 3.69 m/s,

which was required for the current experimental work. A digital tachometer model DT-

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CHAPTER 4 Experimental apparatus

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1236L was used for the calibration of the rotational speed. The calibration data of rpm

(LED display) versus actual sliding speed (at the point of pin-disk contact measured using

the DT-1236L) are shown in Figure 4.3. The calibration was conducted in accordance with

ASTM E74-06 Standard Practice of Calibration of Force-Measuring Instruments for

Verifying the Force Indication of Testing Machines (ASTM, 2006).

Figure 4.3 Calibration data for disk rpm versus sliding speed for the POD test rig

4.3.2 Variable applied load

To investigate friction and wear within the required range of applied loads, it was

necessary to make the applied load variable. The applied load was achieved by sliding

along the central axis of the pin, which was perpendicular to the rotating disk surface.

The load is applied using a series of dead weights together with a load cell assembly that

can slide vertically on the arm and produce a normal force of 8 N. To increase the value of

applied normal load, blocks each with weights of 0.97 kg (applied normal force 9.5 N)

were bolted to the top of the load cell an M10 screw. An example of a normal load of 17.5

N is shown in Figure 4.4.

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CHAPTER 4 Experimental apparatus

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Figure 4.4 Load cell with one additional weight block (applied force 17.5 N)

4.4 Bending arm

The bending arm has two main functions: to hold the load cell in an appropriate position of

contact, and to collect data on the friction force as voltage outputs using its “effective

length”. “Effective length” of the bending arm a length of the thin and relatively flexible

part of the bending arm with strain gauges attached, as shown in Figure 4.5.

ASTM G 115-04 Standard Guide for Measuring and Reporting friction Coefficients

(ASTM, 2004b) was used for the bending arms’ design. According to this guide, the

measurement of dynamic friction force using elastic beams (bending arms) is one of the

recommended methods. To meet this standard guidance, the bending arm as a major

component of the friction force measuring system of the POD test rig weas designed and

fabricated to be elastic enough to measure the dynamic frictional force and stiff enough to

avoid the influence of negative bending arms dynamic motions, which could lead to stick -

slip behaviour or resonance. The major design requirements for the bending arms are

shown in Table 4.2.

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Table 4.2 Design requirements for the bending arms used on the POD test rig

Requirements Range

Load force, N 8-50

Coefficient of friction, µ 0.1-1

Safety factor, F.S.

where:

F.S. = (σy/ σn)

σn – allowable normal stress, N/m2

σy - yield stress, N/m2

5

Allowable strain range, strain

(1 10)×10-3

Cross section factor, a=h/b,

where:

h – hight of cross section area, m

b – wigth of cross section area, m

4

Resonance condition

(This condition is to avoid an effect of

resonance on vibration measurement).

ωn >> ωmax

where:

ωn – natural frequency

ωmax – maximum frequency of disk

rotation

Stiffness condition

(This condition is to separate the

supporting part and the effective part of an

arm and avoid any effect of supporting

part on stiffness of effective part and

vibration measurement).

KS >> Ke

where:

KS – stiffness of supporting part of

bending arm

Ke – stiffness of effective part of bending

arm and n is a number of bending arm (n-

1,2,3 or 4)

To meet these requirements, a maximum allowable normal stress, a natural frequency, and

stiffness for each arm needed to be calculated. The one degree of freedom cantilever beam

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model with one fixed and one free end was used for these calculations (Sundararajan,

2009, Bhattacharjee, 2013, Gere, 2002, Feodosiev, 1974).

The allowable normal stress is calculated using:

σn= σy/ F.S (N/m2), (4.1)

where:

σy is the yield stress (N/m2);

F.S. is the safety factor.

The cross sectional moment of inertia for a rectangular cross section area is given by:

I=(b×h3)/12 (m

4), (4.2)

where:

b is the width of cross section area (m);

h is the height of cross section area (m).

The stiffness is:

Kn=3×E×I/L3 (N/m), (4.3)

where:

E is the modulus of elasticity (N/m2);

L is the effective length of arm (m).

The first natural frequency is:

(Hz), (4.4)

where:

ρ is the density of the arm (kg/m3);

A is the cross section area of the arm (m2).

The arm can be clamped at any one of nine positions. This allows the changing of the

contact point of the pin on the disk without changing the effective arm length as shown in

Figure 4.5b. The natural frequencies in the frictional direction corresponding to the four

values of stiffness, K1, K2, K3, and K4, were calculated. Table 4.2 presents the main

technical parameters of each bending arm.

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Table 4.3 Technical characteristics of bending arms

Arm 1 Arm 2 Arm 3 Arm 4

Material Stainless steel Stainless steel Aluminium Aluminium

Cross-section area,

h, mm x b, mm 18 x 4.5 12 x 3 18 x 4.5 12 x 3

Natural frequency,

Hz 431.26 234.14 255.14 138.88

Stiffness, kN/m 117.1 23.14 40.3 7.96

Effective length, m 0.225 0.225 0.225 0.225

The arms were manufactured from stainless steel or aluminium bar with a rectangular

cross-sectional area. To cover all ranges of applied forces, four bending arms were

fabricated.

To collect friction force data, four strain gauges were mounted to each arm. One of the

bending arms equipped with strain gauges is shown in Figure 4.5.

a) b)

Figure 4.5 Bending arm with attached strain gauges

To measure the actual friction force, all arms were calibrated, friction force versus voltage

output, using a Digital Force Gauge, model 475040. For further experimental work,

bending arm 1 was chosen. The calibration data for bending arm 1 is shown in Figure 4.6.

The calibration was conducted in accordance with ASTM E74-06 Standard Practice of

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Calibration of Force-Measuring Instruments for Verifying the Force Indication of Testing

Machines (ASTM, 2006) and Digital Force Gauge, model 475040 operation technical

manual.

Figure 4.6 Calibration data friction force versus voltage output for bending arm 1

4.5 Load cell

According to the test rig design requirements, the load should be applied along the central

axis of the pin which is perpendicular to the rotating disk surface. For this purpose, a load

cell was designed and built.

The load cell can slide vertically on the bending arm. The arm can only bend horizontally,

providing a measurement of the actual friction force. At the same time, the load cell allows

the pin holder with the pin to slide within the load cell in a horizontal direction if

necessary.

Friction force can be measured by a displacement accelerometer which is mounted on the

load cell in a position perpendicular to the pin and parallel to the disk surface. The

accelerometer is held in position by two springs placed inside as shown in Figure 4.7. The

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springs are flexible in the direction tangential to the disk assembly. This allows measuring

of a sliding displacement of the pin within the load cell. The load cell is equipped with a

displacement sensor as shown in Figure 4.7.

a) b)

Figure 4.7 Load cell with pin and displacement sensor

For precise measurement of the low range of friction forces (0-10 N), where the bending

arm is not sensitive to low friction force, the load cell displacement sensor was calibrated

using a Digital Force Gauge, model 475040. Calibration data for the load cell are shown in

Figure 4.8. The calibration was conducted in accordance with ASTM E74-06 Standard

Practice of Calibration of Force-Measuring Instruments for Verifying the Force Indication

of Testing Machines (ASTM, 2006) and Digital Force Gauge, model 475040 operation

technical manual.

Figure 4.8 Calibration data friction force vs. voltage output for load cell

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4.6 Data acquisition system

In order to monitor experimental work and collect experimental data for further analysis

and calculations, a data acquisition (DAQ) system was designed and built. The initial

requirement for the DAQ system was to record the voltage outputs due to friction forces

from the displacement sensor and the bending arm separately during friction tests.

A computer fitted with a DAQ card was used for recording the voltage outputs from the

displacement sensor and strain gauges. The USB-1408FS DAQ card was manufactured by

Measurement Computing Corporation. To acquire and log data from the DAQ card, the

applications InstaCal and TracerDAQ strip chart/data logger were installed on the desktop

computer. This allowed to record and analyse data received from the displacement sensor

and strain gauges separately or simultaneously. A schematic DAQ system diagram is

shown in Figure 4.9.

Figure 4.9 Schematic diagram of the data acquisition system used for the

measurements of pin displacement and arm forces on the POD test rig

The two amplifiers were used to take the input charge from the load cell accelerometer and

arm strain gauge to provide an output in the form of acceleration. The DAQ device

incorporates built-in integrators which provide displacement outputs.

The minimum computer system requirements were:

Windows 2000 or XP

Microsoft. NET framework v1.1 or 2.0

Adobe Reader

Pentium, 90 MHz

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CHAPTER 4 Experimental apparatus

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RAM 96 MB

Microsoft mouse or compatible pointing device.

4.7 Water supply system

In order to investigate the different lubrication regimes, a simple water supply system was

designed and built. The water supply system contains the following components:

Water tank with tap, capacity of 3 litres

Laboratory magnetic variable stirrer

PVC pipe, diameter 5 mm and length 1.5 m

Water nozzle attached to the load cell

Water collector for water removal

PVC pipe of diameter 10 mm and length 1.5 m which is connected to the test rig

water collector to drain contaminated water.

To simulate the worst of the water-lubricated conditions, and to accelerate the comparative

friction and wear tests, the maximum percentage of contamination was increased by a

factor of 10 with respect to the normal concentration of sand in sea water (Polan et al.,

1981). The grit was kept in suspension in the water using a magnetic stirrer agitating the

water and silica in the supply tank. The schematic diagram of the water supply system is

shown in Figure 4.10.

Figure 4.10 Schematic diagram of the water supply system used on the POD test rig

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The lubrication supply was uniform and thick enough to cover the Pin-Disk contact area.

Ordinary tap water was used as a lubricant because of its low viscosity. A rate of one drop

per second was applied to the disk track at a distance, approximately one cm ahead of the

pin-disk contact point. Visual control during experimental study confirmed that this water

rate was uniform and thick enough to create full water coverage in the Pin-Disk contact

area. The silica sand was introduced as a contaminant. The particle size range of the sand

was measured to be 53-106 µm. The concentration range of the contamination was set at

1%-6%. Visual control (followed by microscopy analysis) confirmed that the sand particles

were distributed evenly under the pin in the Pin-Disk contact area.

4.8 Experimental test rig

A design sketch of the test apparatus is shown in Figure 4.11.

Figure 4.11 Design sketch of the test rig identifying the major components

The test rig compromises a horizontally-situated massive base with a vertically-mounted

variable speed AC motor. The motor is connected to a disk assembly through a pulley

system which has a ratio of 13:1. To significantly reduce to acceptable level any influence

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from the motor vibration to the DAQ system, the motor and the arm are isolated with

rubber pads.

The disk assembly consists a base disk with a test disk that is clamped on top. The test disk

can be changed using a clamping system above the base disk.

The experimental POD test rig, shown in Figure 4.12, was designed, built and

instrumented to investigate the friction between a pair of composite material-steel disk

under dry and water-lubricated conditions.

Figure 4.12 Fully-equipped Pin-on-Disk test rig

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4.9 Experimental methods

In accordance with the design requirements, the following variables were measured:

Friction force voltage outputs

Mass loss

Speed of rotation

Applied load

Water contamination.

4.9.1 Experimental programs

To investigate the tribological behaviour of water-lubricated bearings materials using the

POD test rig within the required range of applied loads and contamination, the

experimental program was developed. According to the research objectives, this

experimental program consisted of the following experimental investigations:

Validation study

Friction study

Wear study

Vibration-wear study.

Material samples were tested in pairs under dry and water-lubricated conditions. According

to the design requirements and scope of the project, the samples were pins fabricated from

selected material, which is commonly used in water-lubricated bearings and disk fabricated

from selected shaft material. The pin is pressed against the disk at a specified load using

dead weights.

4.9.2 Samples preparation

The POD test required the selection of materials for the pin and the disk. According to the

design requirements and scope of the project, the pin was fabricated from selected water-

lubricated bearings material. The pin was fabricated with a flatted tip, and positioned

perpendicular to a flat circular disk. The disk was manufactured from a selected shaft

material.

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The pins used for experimental runs were cylindrical and had a diameter of 10 mm and a

length of 15-20 mm. The test disk was 300 mm in diameter and had a thickness of 6 mm.

The surface roughness of the pin and the disk were measured to be 0.8 µm Ra.

Before each experimental run, the following surfaces preparation procedures were used:

Both, the pin and disk surfaces were cleaned using ethanol and then gradually

polished using progressively finer wet and dry abrasive papers up to extra-fine 1200

grit

The pin was cleaned ultrasonically in an ethanol bath for one minute

The pin after removal from the ultrasonic cleaner ethanol bath and dried using high

pressure air, was weighed to the nearest 0.0001 g, and then carefully placed into the

pin holder of the POD test rig

The disk surface was also cleaned using ethanol and then dried using high pressure

air.

4.9.3 Validation study

The surfaces of the Polytetrafluorethylene (PTFE) pin and the stainless steel disk were

prepared using the same preparation methods discussed in section 4.9.2 before each run.

After sample preparation was completed, the pin was brought into contact with the rotating

disk under a predetermined light normal load. The disk rpm was set corresponding to a

sliding speed of 0.32 m/s. This state was continued for an hour to ensure that there was full

contact between the pin and the disk surfaces. The full experimental load was then applied.

For each load, the friction force (Fx), normal force (Fy) and sliding speed (Vx) were

recorded. A trace of the friction force variation was also measured and recorded. The

friction force readings were taken at a rate of 100 samples per second for the one hour test

period.

The friction force was measured using the voltage outputs from the strain gauges on the

arm, and from these the friction coefficients were calculated. This procedure was repeated

three times. Friction force calculations were verified using a digital forcemeter in

accordance with ASTM E4 - 03 Standard Practise for Force Verification of Testing

Machines. The limit error was calculated to be +1.0 %, which was acceptable in

accordance with ASTM E4 – 03 (ASTM, 2003).

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4.9.4 Investigation of friction

The pin (Railko NF22) and stainless steel disk surfaces were prepared, using the methods

described in section 4.9.2, before each run. After sample preparation was completed and

the lubricant introduced, the pin was brought into contact with the rotating disk under a

predetermined light normal load. The disk rpm was set corresponding to a sliding speed

range of 0.3-4.5 m/s. If contamination was to be used, then the range was set to be between

0-6 %. This state was continued for an hour to ensure that there was full contact between

the pin and the disk. The full experimental load in the range of 8-46 N was then applied.

For each load, sliding speed and percentage of contamination, the friction force (Fx),

normal force (Fy), contamination (%) and sliding speed (Vx) were recorded. A trace of the

friction force variation was also obtained and recorded. During the test, friction force

voltage outputs were measured by strain gauges mounted on the loading arm and by the

displacement accelerometer mounted in the load cell. From these the friction force was

calculated. The friction force voltage outputs readings were taken at 100 samples per

second for the period of 60 minutes. For this purpose, a USB-1408FS microprocessor-

controlled data acquisition system was used. Each experimental run was repeated three

times and then the average of the measurements was recorded.

4.9.5 Wear experiments

The wear experiments were conducted in accordance with ASTM G99-04 Standard Test

Method for Wear Testing with Pin-on-Disk Apparatus for pair of Railko NF 22 - Stainless

steel materials (ASTM, 2004a). The pair of materials tested were Railko NF 22 as the pin

and stainless steel as the disk.

At the beginning of each experiment, the pin and the stainless steel disk were polished and

cleaned using the specimen preparation procedure outlined previously. The lubricant was

introduced and the pin and the disk were brought into contact for sufficient time, usually

one hour, to ensure full contact between the pin and the disk surface. The pin was then

removed, the mass measured accurately (up to four decimal places of a gram) and the pin

was placed in the pin holder at its original position. After wear-testing the pin over a

sliding distance of 15,000-16,000 m, the mass of the pin was remeasured using digital

scales and the difference and the mass loss determined. From this the specific wear rate

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was calculated. Wear tests were repeated three times for each selected sliding distance,

sliding speed, applied load, and contamination level.

Wear tests results were recorded as mass loss in grams and specific wear rate in square

meters per Newton for the pin and plotted as mass loss/specific wear rate versus sliding

speed, applied load, and percentage of contamination.

4.9.6 Vibration-wear relationship study

Vibration–wear tests were conducted under undamped and damped conditions.

The pin (Railko NF22) and the stainless steel disk surfaces were prepared according to

section 4.9.2 before each run. After the lubricant was introduced and the pin was brought

into contact with the rotating disk under a light predetermined normal load. The disk rpm

was set corresponding to a minimum (0.393 m/s) and maximum (1.557 m/s) sliding speeds

and contamination range of 0 - 2 %. After sufficient time for the pin and the disk to

achieve full contact, usually one hour, the pin was removed, the mass was measured

accurately (up to four decimal places of a gram) and then the pin was placed into the pin

holder at its original position. An8 N normal load was then applied for the tests.

For each condition (damped and undamped), sliding speed and contamination, the friction

force (Fx), normal force (Fy), contamination (%), sliding speed (Vx) and mass loss were

recorded. A trace of the friction force variation was also obtained and recorded. During the

test, the friction force voltage outputs were measured by strain gauges mounted on the

loading arm and by a displacement accelerometer mounted in the load cell. The friction

force voltage outputs readings were taken at 100 samples per second for the period of 60

minutes. For this purpose, the USB-1408FS microprocessor-controlled data acquisition

system was used. After testing, the pin mass was measured by the digital scales and the

difference was calculated as mass loss and specific wear rate calculated. Vibration-wear

tests were repeated three times for selected sliding distance, sliding speed, condition

(damped/undamped) and lubrication and the average of mass loss measurements was

recorded.

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4.10 Microscopy examination

The specimen surfaces were examined using a scanning electron microscope (SEM),

QUANTA 450, and a light microscope, Olympus BX60m, with an Olympus DP21 digital

camera and an Olympus SZH stereomicroscope. The surfaces were photographed and

analysed.

The specimens were cleaned prior to examination with alcohol in an ultrasonic bath for one

minute followed by drying in compressed air.

Using these methods of examination allowed the investigation of the effect of water

contamination on the pin and disk materials’ microstructures before and after wear tests to

reveal any possible surface changes.

4.11 Conclusions

The POD test rig was designed and built in accordance with the initial design requirements

with some minor adjustments. An experimental methodology was developed.

The purpose of the POD test rig and the developed experimental technique was to analyse

the performance of various bearing materials in dry, clean water-lubricated and highly-

contaminated water conditions against a stainless steel counterface material. To try to

simulate the worst of the water-lubricated conditions, and to accelerate the comparative

friction and wear tests, the maximum percentage of contamination was increased by a

factor of 10 with respect to the normal concentration of silica in sea water. A PVC pipe

was used to deliver the gritted solution to the pin-disk contact area.

Using these experimental methods, the following experiments were conducted and the

results are presented in this thesis:

Validation study (Chapter 6)

Experimental investigation of friction characteristics (Chapter 7)

Experimental investigation of wear (Chapter 8)

Experimental investigation of vibration-wear relationship (Chapter 9).

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CHAPTER 5 MATERIALS

5.1 Introduction

The metallic antifriction materials (such as gun metal, steel and aluminium alloys) that are

able to work with oil lubrication are conventionally used in liquid-lubricated bearings

designs. In recent years, increasingly close attention has been given to environmentally-

friendly water-lubricated bearings as a cheap and reliable alternative. According to Litwin

(2009), bearings made from polymer-based composites are often used when aggressive

environments prevent the application of other materials (Litwin, 2009).

Another substantial difference between the design of the water-lubricated bearings and

many other bearing systems is the selection process of the materials. The problems with

high performance materials for water-lubricated bearings are particularly relevant to ship

and submarine engineering applications.

According to Rac and Vencl (2005, p.15), apart from adequate strength, the materials for

water-lubricated bearings must also have other crucial characteristics. This is to satisfy the

requirements for reliable and long-lasting operation. These characteristics include low

friction, vibration, good thermal resistance and wear resistance. Rac and Vencl (2005,

p.15) also reported that the mechanical loading is a function of strength of the bearing

materials, while the limits of the thermal loading are determined by the thermal stability of

the selected material (Rac and Vencl, 2005).

Rac and Vencl (2005, p.15) discussed that, with water-lubricated bearings, there is no

direct relationship between the physical and mechanical properties of the materials and the

water-lubricated bearing performance. Parameters, such as the thickness of the lubricant

film, the applied load, and the temperature of the water-lubricated bearing do not depend

on the type of materials used. However, they do have an influence on the material’s

behaviour and, as a consequence, on the selection of the suitable materials (Rac and Vencl,

2005).

Rac and Vencl (2005, p.15) also suggested that water-lubricated bearings materials must

also possess a series of other characteristics that are related to the wear resistance and the

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surface layer properties. Those characteristics are identified as tribological and include

conformability, embeddability, compatibility, deformation, wear, corrosion, and fatigue

resistance (Rac and Vencl, 2005).

Rac and Vencl (2005, p.15) stated that “the tribological behaviour is not just a function of

the bearing material, but also of the surface finish, the lubricant, the design, and the

conditions of the environment in which the bearing operates. The complexity of the

tribological properties of bearing materials and their strong system-dependent properties is

thus explained” (Rac and Vencl, 2005).

It was decided that the proposed experimental work to investigate friction, vibration, and

wear using a POD test rig would include wear modes including fracture, tribochemical

effects, and material loss. According to Unal et al. (2004), transitions between regions

dominated by each of these damage modes can give rise to changes in the friction

coefficient, amplitude of vibration, and wear rate with load, sliding speed, and

contamination (Unal et al., 2004). Therefore, the proper selection of materials is essential.

This chapter presents an analysis of contemporary polymer-based materials for water-

lubricated bearings and of the selection of the materials available for further experimental

work.

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5.2 Effect of water contamination on bearing materials

The viscosity of water is more than 100 times lower than that of oil (Ginzburg et al., 2006).

Figure 5.1 presents the dependence of water viscosity on temperature.

Figure 5.1 Typical baseline of viscosity of water vs. temperature T, 0C, reproduced

from Ginzburg et al. (2006, p.696), Figure 2

The low viscosity of water means that the water-lubricated bearings operate in boundary

and mixed regimes during initial start-up and shutdown. The effects of boundary and

mixed lubrication are determined by the properties of the contacting materials and their

surfaces.

In the case of water contamination, preliminary filters or other devices are commonly used

to remove the contaminants. However, these measures frequently fail to prevent ingress of

soil, sand, clay, or other impurities into the gap between the contacting surfaces. Figure 5.2

provides an example of the severe effect of abrasive wear due to water contamination in

water-lubricated bearings, as reported by Litwin (2009).

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a) b)

Figure 5.2 Water-lubricated bearing damage (subjected to long-lasting operation

which resulted in significant wastage and associated ovalisation of the bush), where

Do=initial dimension, Dp=actual dimension and Dw=wear due to water contamination

as displayed in Litwin (2009, p.44, Figure 6 and p.48, Figure 21)

Polan et al (1981) stated that the abrasive nature of wear due to water contamination needs

to be taken into account when choosing the materials for experimental work. The extent of

this wear depends on the operation regime in the “Stribeck” curve (Polan et al., 1981).

As contaminants find their own way into the clearance, the nature and extent of subsequent

damage will be determined by the properties of the materials. If the contaminants are softer

than these materials, the damage will be insignificant. If, however, the contaminant

particles are harder than the contacting materials, the softer surface will suffer scuffing. As

was shown in the literature review in Chapter 2, little or no information is available on

abrasive wear resistance when operating with contaminated water.

5.3 Materials for water-lubricated bearings

The choice of materials to be lubricated with water cannot be based on the tribological

behaviour under dry sliding friction. This selection would ignore the chemical aspects of

the interaction between water, the contaminant, and the contacting bodies.

Many friction applications operating in the boundary and mixed regimes of lubrication are

characterised by a low range of sliding speeds, frequent surface interruptions, a wide range

of loads, and a comparatively short sliding distance during the lifecycle.

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The main materials used for water-lubricated applications are rubber, ceramics, carbon,

polymer-based materials, and composites of these materials.

5.3.1 Polymer-based thermoplastic materials

Many commercial materials are used for water-lubricated applications. They can be

classified into the following (Ginzburg et al., 2006):

polytetrafluoroethylene (PTFE) and its copolymer materials (Teflon, Polyflon, etc.)

polyamides (PA) (Kapron, Nylon, Rilsan, etc.)

polyformaldehydes (Delrin, Hostaform, etc.)

polycarbonates (Diflon, Lexan, etc.)

polyphenylene oxides (Ariloks, Noryl, etc.)

polyurethanes (Thordon SXL, etc.).

According to Thordon Bearings Inc., comparison tests of different classes of polymer

materials demonstrate that the presence of water as a lubricant in tribocontact reduces the

friction coefficient when compared with the dry friction regime. However, the wear rate

does not directly depend on the friction coefficient, but rather depends on the properties of

the material. The polymer-based thermoplastic materials available for experimental

friction, vibration, and wear tests were PTFE and polyurethane (Thordon SXL).

For the proposed validation study, PTFE was chosen because many previous experimental

studies have been undertaken that are considered PTFE material and its composites for

sliding applications, and its tribological behaviour is well known (Bayer, 2002, Ginzburg

et al., 2006, Ledocq, 1973, Yamajo and Kikkawa, 2004). Typical physical properties of

PTFE are shown in Table 5.1.

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Table 5.1 Physical properties of PTFE material

Properties Value

Density, g/cm3 2.2

Melting point, °C 327

Young's modulus, MPa 500

Yield strength, MPa 23

Coefficient of friction 0.05 - 0.1

Hardness, Dur meter ’D’ 50

Water absorption, % 0.001

A typical PTFE sample used for the validation study presented in this thesis is shown in

Figure 5.3. In accordance with the initial requirements and test rig design specifications,

this sample was fabricated to have a diameter of 10 mm and a length of 20 mm. It also

contains an M8 threaded section for securing the pin into the pin holder during the tests.

Figure 5.3 PTFE test sample used during the validation study

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5.3.2 Carbon - fibre reinforced materials based on thermosetting materials

The strength and rigidity of polymeric materials are higher if they are thermosetting and

reinforced by fibres or textolites (Latin textus – a cloth and Greek lithos – stone), materials

which consist of several layers of fabric (filler) connected by synthetic resin.

Phenol-formaldehyde resins have been used for quite some time as a matrix with cotton

cloth or synthetic fibres as the reinforcing materials. A group of materials suitable for

reliable water-lubricated applications is carbon-fibre reinforced composite with

thermosetting matrices.

One of these materials (NF22 [Railko]) was available for this current experimental work.

Railko NF22 material was developed in the early 1980s for water-lubricated and dry

bearing systems and according to Cumberlidge (2009), this bearing material has been

especially designed to cope with extreme operational conditions such as loads, speeds,

temperature fluctuations, water contamination, etc. It is reported by Cumberlidge (2009)

that this material is used by more than 30 Navies around the world for water - lubricated

bearings on ships, ferries and submarines (Cumberlidge, 2009).

Depending on the application and grade, NF22 (Railko) can operate dry, partially

lubricated, or fully lubricated in oil or sea water (Cumberlidge, 2009). Typical physical

properties for the NF22 (Railko) material are shown in Table 5.2.

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Table 5.2 Physical properties of NF22 (Railko) material (WÄRTSILÄ, 6/09/2007)

Properties Value

Density, g/cm3

1.64

Ultimate compressive strength, kg/cm2

normal to laminate (radial)

parallel to laminate (axial)

1800

1000

Max. working compressive strength, kg/cm2

radial

axial

450

250

Compressive modulus (radial), kg/cm2

41000

Ultimate tensile strength, kg/cm2

310

Tensile modulus (axial), kg/cm2

310

Normal working pressure, MPa 55

Max. operating temperature, oC

continuous

occasional

100

120

Impact strength, KJ/m²

Charpy

Izod unnotched

35

75

Shear strength, MPa 41

Brinell hardness, HB 29

% swell in water, at 20°C <1

Coefficient of friction (dry) 0.25-0.4

Coefficient of thermal expansion, 10-6

/°C (normal)

radial

axial

60

40

Hardness, KJ/m² 85

Thermal conductivity at 50°C, W/m.k 0.74

Fire rating per NF F-16-101 F1.I3

NF22 (Railko) material was chosen as the primary material for the experimental

investigation of friction, vibration, and wear because, according to the Wärtsilä manual

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(WÄRTSILÄ, 6/09/2007), the NF22 (Railko) grade of material comprises an anti-scuffing

thermosetting resin reinforced by organic fibres and enhanced with dry lubricants.

An NF22 (Railko) sample used for the experimental study is shown in Figure 5.4. In

accordance with the initial requirements and test rig design specifications, this sample was

fabricated to have a diameter of 10 mm and a length of 15 mm. It also features an M8

threaded section to secure it in the pin holder during tests.

Figure 5.4 NF22 (Railko) sample used in the experimental study

5.3.3 Shaft materials

The shafts of water-lubricated bearings for the marine environment are manufactured from

gun metal (e.g. BS 1400 LG4 or similar), stainless steel (e.g. AISI 440C, AISI 316 or

similar) or a combination of both materials (WÄRTSILÄ, 6/09/2007, Yamajo and

Kikkawa, 2004).

To simulate water-lubricated bearing shafts, an AISI 440C stainless steel material was

chosen as the disk material for the experimental studies presented in this thesis. The typical

chemical composition in weight percent (%) is:

Carbon - 1.08;

Silicon - Max 1.0;

Manganese - Max 1.0;

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Chromium - 17.0

Molybdenum - 0.75

Phosphorus – 0.04

Sulphur – 0.03

Iron - Balance.

The test disk was fabricated with the following specifications:

Disk diameter, 300 mm

Disk thickness, 6 mm

Surface roughness of the disk, 0.8 µm Ra.

A fabricated stainless steel test disk, fitted on the POD test rig for experimental study, is

shown in Figure 5.5. It has eight clamp holes to secure it to a base disk during the

experimental work.

Figure 5.5 AISI 440C stainless steel test disk fitted on the POD test rig

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5.4 Conclusions

In accordance with the materials requirements, the water–lubricated bearings materials

were selected so as to investigate the process of friction, vibration, and wear with the

different load (0-50 N) and sliding speed (0.3-4.5 m/s) ranges in dry, boundary, and mixed

lubrication regimes.

The test disks and sample pins were fabricated from the same materials as used in

commercial water-lubricated bearings, e.g.:

For the validation study: PTFE (pin), stainless steel (disk)

For the experimental study: NF22 (Railko) (pin), stainless steel (disk).

The cylindrical pin specimens, 10 mm in diameter and 15-20 mm in length, were tested

against an AISI 440C stainless steel disk. The surface roughness, prior to the disk testing,

was measured to be 0.8 µm Ra.

The aim of this chapter was to analyse the performance of various water-lubricated

bearings materials in dry, water-lubricated and highly abrasive conditions, against a

stainless steel counterface material.

This chapter presented a review of contemporary water-lubricated bearings materials, their

analysis and selection of available materials for the following validation study and

experimental investigations, which are presented in the next chapters of this thesis

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CHAPTER 6 Validation study

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CHAPTER 6 VALIDATION STUDY

6.1 Introduction

A POD test rig validation study program was set up to verify the capability of the newly-

designed and fabricated test rig. These validation friction tests were conducted under dry

lubrication at a sliding speed of 0.32 m/s and a normal applied load range of 8 - 36.5 N.

The results were compared with the published results (Godfrey, 1995, Unal et al., 2004,

Yamajo and Kikkawa, 2004).

6.2 Test plan and procedure

A validation study was conducted to assess the new experimental POD test rig, described

in Chapter 4, against the design requirements and to assess whether a POD test rig was

appropriate for this use. The validation study was conducted in accordance with ASTM

E4-03 Standard Practise for Force Verification of Testing Machines (ASTM, 2003) and

ASTM G115-04 Standard Guide for Measuring and Reporting Friction Coefficients

(ASTM, 2004b)”.

The purposes of the validation study were:

To quantifiably characterise the POD test rig performance

To assess the potential for errors

To identify method-to-method differences

To determine if the experimental test rig would meet existing regulations (ASTM

E4-03 Standard Practise for Force Verification of Testing Machines and ASTM

G115-04 Standard Guide for Measuring and Reporting Friction Coefficients).

The technical parameters, including the specific test conditions and experimental samples

adopted for the validation study are provided in Table 6.1.

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Table 6.1 Technical parameters adopted for the validation study

A number of POD experimental studies have been conducted to examine the influence of

test speed and load values on the friction and wear behaviour of pure PTFE, glass fibre

reinforced PTFE, bronze and carbon filled PTFE polymers (Godfrey, 1995, Unal et al.,

2004, Yamajo and Kikkawa, 2004).

Unal et al (2004) carried out the friction and wear tests versus AISI 440C stainless steel

disk were conducted on a Pin-on-Disk test rig at a dry condition. Tribological tests were at

room temperature, using 5 N, 10 N, 20 N, and 30 N loads and at 0.32 m/s, 0.64 m/s, 0.96

m/s and 1.28 m/s speeds. These experimental results were used to validate a newly built

POD test rig (Unal et al., 2004).

The current validation study, presented in this Chapter, was conducted at room

temperature. To measure the friction force under dry regime, a constant sliding speed of

0.32 m/s was selected. The load was incrementally increased from 8 N to 36.5 N.

6.3 Results and discussions

Figure 6.1 presents variations of the friction coefficient values (between each run) for the

PTFE pin against the stainless steel disk tested at room temperature under dry wear

conditions at 0.32 m/s sliding speed and at 8 N, 17.5 N, 27 N and 36.5 N normal loads.

Parameters Experimental Rig Applied normal force 8 N

17.5 N 27 N 36.5 N

Lubrication Dry

Sliding speed 0.32 m/s

Materials: Disk Pin

Stainless steel (AISI 440C) Polytetrafluorethylene (PTFE)

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CHAPTER 6 Validation study

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Each run was repeated three times. The error bars represent the variation of the friction

coefficients calculated from measured friction forces.

Figure 6.1 Coefficient of friction of the PTFE pin against a stainless steel disk for a

sliding speed of 0.32 m/s

The experimental measurements of the friction forces and calculated friction coefficients

were in good agreement with experimental results reported in the literature (Unal et al.,

2004, Godfrey, 1995, Yamajo and Kikkawa, 2004).

The experimental results of the validation study show that for the PTFE material, the

coefficient of friction decreases with an increase in applied load. This effect is known for

polymers that are visco-elastic materials and their deformation under applied load is visco-

elastic. Thus, the variation of the friction coefficient with the load for visco-elastic

materials is given by the following equation (Unal et al, 2004):

µ=kN(n-1)

(6.1)

where:

µ is the coefficient of friction;

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CHAPTER 6 Validation study

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N is the normal load;

k is a constant;

n is a constant, its value 2/3 <n <1.

Unal et al (2004) stated that according to this equation, the friction coefficient decreases

with an increasing normal load due to the visco-elastic deformations. In this case, when the

load increases to the load limit values (just before yield) of the polymer, the friction and

wear will decrease due to the critical surface energy of the polymer. Unal et al (2004)

reported that this is due to frictional heat increasing the temperature of the sliding surfaces

which leads to the relaxation of the polymer chains. Therefore, molecules at the polymer

surfaces are compressed, drawn and sheared. Highly active radicals then react with

unbroken chains and give rise to a series of new chains (Unal et al., 2004).

6.4 Conclusions

A newly built POD experimental test rig has been used for a validation study of the effect

of an applied normal load on the friction coefficient under dry friction conditions.

The experimental results of the validation study show that for the PTFE - stainless steel

materials the coefficient of friction decreases with an increase in normal load due to visco-

elastic deformations.

It was investigated and confirmed that the obtained results of validation study are in a good

agreement with published experimental results.

Wear and vibration-wear validation is not required for the POD test rig in accordance with

ASTM E4-03 Standard Practise for Force Verification of Testing Machines where “A

testing machine shall be verified as a system with the force sensing and indicating devices

in place and operating as in actual use” (ASTM, 2003). Digital scales were used to

measure the mass loss of the pin used in these wear and vibration-wear tests. These digital

scales are not a part of POD arrangement; they were verified apart from this project as

laboratory equipment.

Using this information, the following chapter presents experimental investigation of

friction characteristics of NF22 (Railko)-AISI 440C materials pair. This is used to

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experimentally examine the effect of water contamination on the friction coefficient of

water-lubricated bearing materials.

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CHAPTER 7 EXPERIMENTAL INVESTIGATION OF

FRICTION CHARACTERISTICS

7.1 Introduction

The experimental investigation of friction in mechanical systems depends on many factors.

According to Pilipchuk (2002) some of these factors include the material properties and

geometry of sliding bodies, surface finish, surface chemistry, sliding speed, temperature,

normal load, and type of lubrication (Pilipchuk, 2002).

An experimental program was conducted to examine the effect of varying levels of

contamination of water lubricant on the friction coefficient of materials for water-

lubricated bearings. Experiments were conducted at room temperature for the following

ranges:

Sliding speeds, 0.393-1.557 m/s

Applied normal loads, 8-46 N

Contamination, 0-6%.

The experimental study was conducted in accordance with ASTM G115-04 Standard

Guide for Measuring and Reporting Friction Coefficients (ASTM, 2004b).

7.2 Experimental study of the effect of contamination on friction: experimental plan and procedure

This experimental study of friction was conducted to examine the effect of varying

contaminated lubrication regimes on the friction coefficient of materials for water-

lubricated bearings within boundary and mixed lubrication regimes.

The purposes of this experimental study of were to:

Investigate the effect of water contamination on the friction coefficient of materials

used for water-lubricated bearings

Assess material performance under dry and water-lubricated conditions

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Identify the friction mechanism under dry and water-lubricated conditions.

The selected test parameters for the experimental study of friction characteristics are given

in Table 7.1.

Table 7.1 Operational parameters adopted for experimental study

Parameters Experimental Rig

Applied normal force 8 N

17.5 N

26 N

36.5 N

46 N

Lubrication Water-lubricated

Water contamination

0%

1%

2%

4%

6%

Sliding speeds 0.393 m/s

0.767 m/s

1.158 m/s

1.557 m/s

Test duration 1 hour

Disk diameter

Pin diameter

0.3 m

0.01 m

Sand particles’ size 53 – 106 μm

Materials:

Disk

Pin

Stainless steel (AISI 440C)

NF22 (Railko)

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7.3 Experimental study of the effect of contamination on friction: results and discussions

7.3.1 Investigation of friction under water lubrication

Two sets of experiments were conducted for both the dry and the clean water-lubricated

conditions at room temperature, namely:

Friction coefficient versus normal load

Friction coefficient versus sliding speed.

The first set of experiments was conducted at room temperature under dry and clean water

lubrication conditions. A constant sliding speed (0.393 m/s) was selected and the load was

incrementally increased from 8 N to 46 N. The friction force was recorded by measuring

voltage outputs from the strain gauges on the arm, and from this data the friction

coefficients were calculated. This procedure was repeated three times for each sliding

speed of 0.767 m/s, 1.158 m/s and 1.557 m/s under both dry and water lubrication

conditions. The purpose of these tests was to collect base data on the variation of the

friction coefficient with changing normal load under dry and clean water lubrication and to

identify the boundary and mixed lubrication regimes for clean water.

The second set of experiments was also conducted under dry and clean water lubrication

conditions. The sliding speed was increased gradually from 0 m/s to the maximum possible

sliding speed of 3.69 m/s at each load of 8 N, 17.5 N, 27 N, 36.6 N and 46 N, respectively.

The friction force voltage outputs were measured and recorded, and friction coefficients

were calculated. This procedure was repeated for all sliding speeds (0.767 m/s, 1.158 m/s

and 1.557 m/s). The purpose of this second set was to investigate the effect of sliding speed

on the friction coefficient and to enable the identification of the boundary and mixed

lubrication regimes with clean water lubrication.

Figures 7.1 and 7.2 illustrate the experimental results for variations of the friction

coefficient values for the NF22 (Railko) pin against the stainless steel disk. The tests were

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conducted at room temperature under dry lubrication conditions, at 8 N, 17.5 N, 27 N,

36.5 N, and 46 N loads, and at various sliding speeds.

Figure 7.1 Coefficient of friction of NF22 (Railko) material against stainless steel

versus normal applied load under dry conditions

The dependence of the friction coefficient on the relative velocity between sliding surfaces

is complex and depends on the material properties at the interacting surfaces. In dry sliding

contact between interacting surfaces, friction can be modelled as elastic and plastic

deformation forcing microscopic asperities into contact. The friction force can be described

by the following equation reproduced here from Ibrahim (1994a, p. 218, Eq. 15):

F1(Pn)=fas (Aas/At) Pn (7.1)

where:

F1(Pn) is the friction force due to asperity contact (N);

fas is the asperity sliding friction coefficient;

Aas is the real area of contact (m2);

At is the total area of contact (m2);

Pn is the contact load (N).

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According to this equation (Ibrahim, 1994a), increasing the load leads to greater contact

between the asperities and therefore, a higher friction coefficient.

Figure 7.2 Coefficient of friction of NF22 (Railko) material against stainless steel

versus sliding speed under dry conditions

For dry conditions, the friction coefficient increases with increasing load and sliding speed.

The friction–speed curves are non-linear and are dependent on three factors; the material

properties of the sliding surfaces, the normal load, and the medium which occupies the gap

between them. The NF22 (Railko) material has three phases (matrix and organic fibres)

and a solid lubricant. According to Mosleh et al (2002), increasing the applied normal load

increases the friction force due to increasing the area of asperity contact (Mosleh et al.,

2002). The nonlinearity of the friction coefficient–speed curve dependence is in strong

agreement with the literature and is attributed to the creep deformation of the interface

asperities (Martins, 1990). More advanced examples of dry friction models of this

mechanism were described in the literature by Marklund and Larsson (2008) where the

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friction coefficient was found to be a function of speed and temperature in addition to the

normal load (Marklund and Larsson, 2008).

Figures 7.3 and 7.4 illustrate the experimental results of variations in the friction

coefficient values for an NF22 (Railko) pin against a stainless steel disk tested at room

temperature using clean water lubrication, at 8 N, 17.5 N, 27 N, 36.5 N, and 46 N loads,

and at various sliding speeds.

Figure 7.3 Coefficient of friction of NF22 (Railko) material against stainless steel

versus normal load under clean water-lubricated conditions

It can be seen from Figure 7.3 that the value of the friction coefficient decreases with

increasing load and sliding speed. This is the opposite effect to that found under dry

lubrication conditions. The normal load exerts slightly less influence on the friction

coefficient than sliding speed, particularly at high speeds. Marklund and Larsson also

reported that the friction coefficient was not particularly load dependent (Marklund and

Larsson, 2008).

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Figure 7.4 better illustrates the relatively small influence that the normal load has on the

coefficient of friction, particularly at higher speeds. The largest difference in the value of

the coefficient of friction for different normal loads is approximately 0.18 which is at the

lowest sliding speed of 0.393 m/s.

Figure 7.4 Coefficient of friction of NF22 (Railko) material against stainless steel

versus sliding speed under clean water-lubricated conditions

This type of friction behaviour (NF22 (Railko) material against stainless steel) was

explained in the literature by the elastohydrodynamic theory of lubrication and theory of

asperity contact (Unal et al., 2004, Godfrey, 1995, Ibrahim, 1994a). In the zones of real

contact, all types of deformation (elastic, elastoplastic, and plastic) can occur. The friction

coefficient for materials with a high modulus of elasticity changes during plastic contact as

the load increases, as shown in Figure 7.4. This change is due to the interaction between

asperities and deformation processes in the zones of real contact. The maximum friction

coefficient conforms to elastic asperities’ contact and the transition to a state where the

interaction between irregularities begins to affect the deformation processes in the zones of

real contact. The decrease of the friction coefficient due to increasing load is a result of

transition from elastic asperity contact to plastic deformation at the points of contact. The

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increasing load makes asperities in contact plastically deform which may release the solid

lubricant; in conjunction with water, this provides better lubrication by creating a wedge of

lubricant and reducing the friction coefficient as stated in Unal et al (2004).

It is evident that the friction coefficient of NF22 (Railko) material against stainless steel

under clean water lubrication conditions is much more dependent on the sliding speed than

on the normal applied load. According to the Stribeck model, all this experimental work is

within either boundary or mixed regimes of lubrication. Olsson et al (1997) reported that,

the total friction force at contact consists of two components; one is due to asperity contact

(equation 7.1) and the second is due to hydrodynamic viscous shear (Olsson et al., 1997):

F2(V)=Fc V (7.2)

where:

F2 is the friction force due to hydrodynamic viscous shear (N);

Fc is a constant, dependent on lubricant viscosity, area of contact and minimum film

thickness of the lubricant in the hydrodynamic regime (N/m/s);

V is the sliding speed (m/s).

The negative slope (boundary and mixed regimes) of the “Stribeck” model is clearly seen

in the experimental results, shown as Figures 7.3 and 7.4.

7.3.2 Investigation of the friction coefficient and the effect of contamination

The aim of these experiments was to investigate and analyse the effect of water

contamination on the relationship between friction and speed for NF22 (Railko) material

against an AISI 440C stainless steel disk.

Figures 7.5 to 7.8 present the graphs of the experimental results of the friction coefficient

for NF22 (Railko) against stainless steel under contaminated water-lubricated conditions,

at 8 N, 17.5 N, 27 N, 36.5 N, and 46 N loads, at 0.393 m/s, 0.767 m/s, 1.158 m/s, and

1.557 m/s sliding speeds, and at various levels of water contamination of 1%, 2%, 4%, and

6%, respectively.

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Figure 7.5 Coefficient of friction of NF22 (Railko) against stainless steel for 1%

contaminated water lubrication

Figure 7.6 Coefficient of friction of NF22 (Railko) against stainless steel for 2%

contaminated water lubrication

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Figure 7.7 Coefficient of friction of NF22 (Railko) against stainless steel for 4%

contaminated water lubrication

Figure 7.8 Coefficient of friction of NF22 (Railko) against stainless steel for 6%

contaminated water lubrication

These Figures show that the value of the coefficient of friction decreases with increasing

sliding speed and load, but increases with increasing water contamination.

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According to Olsson et al. (1997), contamination is another factor that introduces

complications. The presence of small particles of foreign material between the sliding

surfaces gives rise to additional forces that strongly depend on the size and material

properties of the contaminants (Olsson et al., 1997).

When the results for contaminated water are compared with those for clean water

lubrication, the tribological behaviour can still be described by the “Stribeck” model. At

the same time, it can be seen that the friction coefficient increases with increasing

contamination due to the influence of the abrasive sand particles between interacting

bodies at all sliding speeds. Moreover, the maximum increase in the friction coefficient

was seen at the slowest sliding speed of 0.393 m/s due to asperity contact, as well as the

abrasive sand particles. For further analysis of the effect of water contamination on the

friction coefficient, the slowest sliding speed of 0.393 m/s which resulted in the largest

value of the friction coefficient was chosen. The results are shown in Figure 7.9.

Figure 7.9 Coefficient of friction vs. water contamination of NF22 (Railko) material

(sliding speed=0.393 m/s)

As mentioned above, friction in water-lubricated sliding bodies depends on friction due to

asperities’ contact, lubricant contamination, and hydrodynamic viscous shear friction.

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Whether any nonlinearities due to the effect of water contamination on friction exist is not

yet known.

As seen on Figure 7.9, the friction force due to water contamination is a function of the

degree of contamination (g).

Thus, the total friction force under contaminated water lubrication can be described as

follows:

F=F1 (Pn)+F2 (V)+F3 (g) (7.4)

where:

F is the total friction force (N);

F1(Pn) is the friction force due to asperities’ contact force (N);

F2(V) is the friction force due to hydrodynamic viscous shear force (N);

F3(g) is the friction force due to water contamination (N).

The complex effect of water contamination on friction has not been thoroughly

investigated, however to simplify further investigations of friction due to water

contamination, it is assumed that F3(g) has a linear dependence of friction coefficient

against the level of water contamination (g).

7.4 Conclusions

A new experimental approach using the POD experimental apparatus for the study of the

effect of water contamination on the friction coefficient has been conducted. The

experimental range of sliding speeds was identified as being within the boundary and

mixed (transition) regimes of the “Stribeck” curve.

The effect of water lubrication and water contamination on the pair of materials

comprising NF22 (Railko) composite and stainless steel was experimentally investigated

and the following conclusions can be drawn:

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1. The friction coefficient of NF22 (Railko) composite material decreases when the

applied load is increased due to the material properties (solid lubricating

component and plastic deformation).

2. The friction studies of the composite material (NF22 (Railko)) against an AISI

440C stainless steel disk under various loads, sliding speeds, and water

contamination show an increase in the friction force with increasing water

contamination.

3. It is not yet known if any nonlinearities from the effect of water contamination on

friction exist. As a first iteration, it can be assumed that friction force due to water

contamination increases linearly with an increase in water contamination.

4. For the specific ranges of loads and speeds investigated in this study, the sliding

speed has a greater effect than the applied normal load on the friction coefficient

for this composite material.

The friction in water-lubricated bearings depends on a number of factors such as applied

load, sliding speed, bearing design, and environment conditions. Therefore, the value of

this POD test method is to predict the relative ranking of water-lubricated bearings

materials combinations rather than simulating in service water-lubricated bearings.

However, further extensive experimental and theoretical studies are required to properly

understand the effect of water contamination on friction and the physical implications of

variations in the material properties of composite materials used for engineering

applications.

A mathematical model of contact mechanics between the two friction surfaces is necessary

for the analysis and design of water-lubricated bearings materials under contaminated

water conditions. In this experimental work, a simple empirical linear assumption of

contact mechanics based on experimental results has been proposed (see Equation 7.4) for

water contamination at levels between 1% and 6%.

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The next chapter uses the information from these experiments as a basis for the

investigation of the effect of water contamination on the mass loss and the specific wear

rate of materials used for water-lubricated bearings.

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CHAPTER 8 EXPERIMENTAL INVESTIGATION OF WEAR

8.1 Introduction

In engineering applications of water-lubricated bearings, where the water contains different

types of contaminants including solid silica particles, the use of wear-resistant materials is

required. According to Prehn (Prehn et al., 2005), in addition to the contamination by silica

particles, the water used for lubrication may also be chemically aggressive. For these

reasons, the choice of materials is limited.

An experimental program was undertaken to examine the effect of contaminated water

lubrication regimes on the mass loss and the specific wear rate of materials used for water-

lubricated bearings. This study investigated the sliding response of NF22 (Railko) material

against stainless steel. Experiments were conducted at room temperature for sliding speeds

in the range of 0.393 to 1.557 m/s and applied normal loads of 8 N to 46 N.

The three-body wear mechanism involved is identified as being a function of lubrication,

water contamination, and material properties. The relative wear resistance of samples is

compared using specific wear rate calculations as a function of mass loss, whilst results of

surface microscopy are also presented. The experimental study was conducted in

accordance with ASTM G 99-04 Standard test method for wear testing with a Pin-on Disk

apparatus (ASTM, 2004a).

8.2 Experimental study of wear: experimental plan and procedure

An experimental study of wear was conducted in order to examine the effect of varying

contaminated water lubrication regimes on the mass loss and specific wear rate of

materials for water-lubricated bearings.

The purposes of this experimental study were:

To investigate the effect of water lubrication and contamination on the wear rate of

materials used for water-lubricated bearings

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To assess the material performance under contaminated and non-contaminated

water-lubricated conditions

To identify the significant wear mechanisms for contaminated water-lubricated

conditions.

The parameters used for the experimental study of wear, including specific test conditions

and experimental samples, are provided in Table 8.1.

Table 8.1 Parameters for wear test

Parameters Experimental Rig

Applied normal force 8 N

17.5 N

26 N

36.5 N

46 N

Lubrication Water-lubricated

Water contamination

0%

1%

2%

4%

6%

Sliding speeds 0.393 m/s

0.767 m/s

1.158 m/s

1.557 m/s

Test duration

Disk diameter

Pin diameter

Sand particles size range

1 hour

0.3 m

0.01 m

53-106 µm

Materials:

Disk

Pin

Stainless steel

NF22 (Railko)

This experimental work was conducted at room temperature. The POD test rig was run for

four hours before taking any measurements to ensure full contact between the pin and the

disk surface. For the first run of experiments, a constant sliding speed of 0.393 m/s and a

load of 8 N were used. Each run was repeated three times. After each run, the mass of the

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pin was measured using digital scales and the mass loss recorded. Subsequently, the same

sliding speed was used; the load was increased in steps up to 46 N. The mass losses were

measured and recorded, and the specific wear rates calculated. This procedure was

repeated three times for all sliding speeds.

No measurable mass losses were recorded during wear tests under clean water lubrication.

The specific wear rate was therefore, zero. In the case of clean water lubrication, only

adhesive wear mechanisms are possible and adhesion is significantly reduced by

lubrication. This appears to be due to the specific material properties of the NF22 (Railko)

composite, which has a solid lubricant incorporated into its structure.

8.3 Experimental study of wear: results and discussions

8.3.1 Wear tests

Figures 8.1 to 8.4 present mass loss values for NF22 (Railko) under contaminated water

lubrication conditions at 8 N, 17.5 N, 27 N, 36.5 N, and 46 N loads, at 0.393 m/s,

0.767 m/s, 1.158 m/s, and 1.557 m/s sliding speeds, and at 1%, 2%, 4%, and 6% water

contamination, respectively.

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Figure 8.1 Mass loss for NF22 (Railko) material at different load and speed values

under 1% water contamination

Figure 8.2 Mass loss for NF22 (Railko) material at different load and speed values

under 2% water contamination

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Figure 8.3 Mass loss for NF22 (Railko) material at different load and speed values

under 4% water contamination

Figure 8.4 Mass loss for NF22 (Railko) material at different load and speed values

under 6% water contamination

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CHAPTER 8 Experimental investigation of wear

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From these graphs it is evident that the value of mass loss decreases with increasing sliding

speed and load, and increases with greater water contamination. Olsen et al (1997)

explained this effect as the adhesive wear mechanism is coupled with abrasive wear

mechanism under contaminated water lubrication. The presence of small particles of sand

between sliding surfaces gives rise to additional forces that strongly depend on the size and

material properties of the contaminants (Olsson et al., 1997).

Comparing the results with those of clean water lubrication, it is clear that the wear

mechanism can be described generally by adhesive-abrasive wear. This is where sand

particles are in contact with the surfaces of the pin and the disk, resulting in material being

displaced from the pin as well as the disk. At the same time, the mass loss increases with

increasing water contamination due to the role played by the abrasive sand particles

between contacting bodies. At the same time, it can be seen that the mass loss increases

with increasing water contamination due to the influence of abrasive sand particles

between interacting bodies at all sliding speeds in the range. Moreover, the maximum

increase in the mass loss was seen at the slowest sliding speed (boundary regime of

lubrication) of 0.393 m/s due to interactions between asperities as well as abrasive sand

particles. This speed was therefore, chosen for further analysis of the effect of

contamination on the mass loss. These results are shown in Figure 8.5.

Figure 8.5 Mass loss versus degree of water contamination for NF22 (Railko) material

at a sliding speed of 0.393 m/s

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CHAPTER 8 Experimental investigation of wear

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This graph suggests that, at low load, more sand particles are drawn into the contact area

which leads to greater mass loss. It is expected that with a change in particle size, the

character of this plot could be changed significantly. Figure 8.5 also shows that, the mass

loss due to water contamination is a nonlinear function of the degree of contamination. To

investigate the effect of water contamination as well as the type of contaminant, further

theoretical and experimental work is required. A theoretical three-body wear model is

necessary for determination of the wear mechanism under water contamination for further

stability analysis and design.

The task of obtaining a theoretical three-body wear model for these types of contact is

difficult due to the complexity of this mechanism with input variables of abrasive as well

as adhesive wear. It is therefore recommended for future work.

In the case of contaminated water lubrication, a combination of adhesive and severe

abrasive wear mechanisms occur. As shown in the experimental results, the mass loss

decreases with increased sliding speed and applied load, and increases with greater water

contamination. This can be explained by the boundary and mixed regimes of “Stribeck”

model of hydrodynamic and three-body wear mechanism (Torrance, 2005).

A micrograph of the pin’s surface before and after the full cycle of experiments is shown in

Figure 8.6.

a) before b) after

Figure 8.6 Micrograph of pin’s worn surface before/after a full cycle of experiments,

at a magnification of X500

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CHAPTER 8 Experimental investigation of wear

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Despite the severe level of contamination, the images are similar and no visible structural

changes to the pin’s surface or embedded silica particles are evident. Torrance (Torrance,

2005) suggested that the process of wear depends on the mechanics of the contact between

the two surfaces entraining the abrasive particles, on particle size and percentage of

contamination, and often on chemical interactions between the wearing surface and the

surrounding media. It was reported in WÄRTSILÄ (2007) that it has generally been

assumed that the presence of organic fibres and solid lubricant in composite materials

makes it wear resistant.

Figure 8.7 presents the micrograph of the same worn surface of the pin. This micrograph

was obtained using an optical stereomicroscope. As seen from this Figure, the NF22

(Railko) material consists of two visible phases (matrix and organic fibres) creating good

wear resistance for this material.

Figure 8.7 Micrograph of pin’s worn surface after a full cycle of experiments, at a

magnification of X100

The micrographs of wear traces on the paired worn pin’s surface and stainless steel disk

are shown in Figures 8.8 and 8.9, respectively. Both samples were thoroughly cleaned,

using ethanol and compressed air, before microscopic examination.

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As reported by Moore (1978), for abrasive particle contact with very low loads, the contact

will be predominantly elastic. “Such contacts may result in material removal by surface

molecular mechanisms or from surface films or by Hertzian fracture of brittle materials”

(Moore, 1978). As the load on an angular abrasive particle increases, contact on both

ductile and brittle materials will involve plastic deformation to a greater extent leading to

surface damage.

Figure 8.8 Micrograph of pin’s worn surface after a full cycle of experiments

Figure 8.9 Micrograph of stainless steel disk’s worn surface after a full cycle of

experiments, at a magnification of X100

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CHAPTER 8 Experimental investigation of wear

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It is clear that in spite of the good wear resistance of NF22 (Railko) material, the three-

body abrasion wear mechanism could significantly damage both surfaces in a severely

contaminated water-lubricated environment although there are no visible structural

changes to the pin or sand particles embedded in the composite pin or steel disk.

8.3.2 Specific wear rate calculations

To analyse the volume of plastically-displaced material, a specific wear rate needs to be

calculated.

Unal et.al (2004) reported the following calculation procedure which was adopted for this

experimental study. The sliding wear data reported here should be the average of at least

three runs. The average mass loss was used to calculate the specific wear rate (Unal et al.,

2004):

K0=∆m/L*F*ρ (m2/N), (8.1)

where:

∆m is the average mass loss (kg);

L is the sliding distance (m);

F is the applied load (N);

ρ is the density of the material (kg/m3).

L=v*t (m), (8.2)

where:

v is the sliding speed (m/s);

t is the duration of the test (s).

Figures 8.10 to 8.13 illustrate the specific wear rate values calculated for each test sliding

speed and normal load under contaminated water-lubricated conditions, at 8 N, 17.5 N,

27 N, 36.5 N, and 46 N loads, at 0.393 m/s, 0.767 m/s, 1.158 m/s, and 1.557 m/s sliding

speeds, and at 1%, 2%, 4%, and 6% water contamination, respectively.

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CHAPTER 8 Experimental investigation of wear

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Figure 8.10 Specific wear rate for NF22 (Railko) material at different load and speed

values with 1% water contamination

Figure 8.11 Specific wear rate for NF22 (Railko) material at different load and speed

values with 2% water contamination

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CHAPTER 8 Experimental investigation of wear

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Figure 8.12 Specific wear rate for NF22 (Railko) material at different load and speed

values with 4% water contamination

Figure 8.13 Specific wear rate for NF22 (Railko) material at different load and speed

values with 6% water contamination

It is clear that the sliding speed variation has little influence on the specific wear rate of the

NF22 (Railko) material, but the value of the specific wear rate increases with increasing

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water contamination. As mentioned above, the most significant variation of specific wear

rate versus water contamination and applied load was seen at the lowest sliding speed as

displayed in Figure 8.14.

Figure 8.14 Specific wear rate versus degree of water contamination of NF22 (Railko)

material for a sliding speed of 0.393 m/s

Unal et al. (2004) stated that the wear process involves fracture, tribochemical effects, and

plastic deformation. Variation to the applied normal load and contamination lead to

transitions between regions dominated by each of these processes and commonly give rise

to changes in the specific wear rate. Furthermore, this result is closely related to structural

characteristics and chemical effects occurring in the frictional processes, as well as to

transfer film formation on the counterfaces (Unal et al., 2004).

For the NF22 (Railko) material tested in this experimental investigation within the sliding

speed range from 0.393 m/s to 1.558 m/s, the speed and normal load had less influence on

the specific wear rate than changes to the level of water contamination. The main reason

for the specific wear rate increase is the complex three-body wear mechanism, including

abrasive and adhesive wear.

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The main purpose of these wear tests was to analyse the performance of samples of NF22

(Railko) bearing material in highly abrasive operational conditions, when forced against an

AISI 440C stainless steel.

This resultant knowledge provides a base for further investigation into a three-body wear

mechanism under contaminated water conditions. This could lead to the development of a

more advanced three-body nonlinear wear model for water-lubricated bearings and other

engineering applications.

8.4 Conclusions

The effect of contaminated water lubrication on the wear mechanism of the pair of

materials comprising NF22 (Railko) composite and AISI 440C stainless steel was

experimentally investigated. The following conclusions can be drawn:

1. Mass loss and the specific wear rate of NF22 (Railko) composite material decreases

as the applied normal load increases. This is due to plastic deformation of the

contacting surfaces.

2. Wear studies of the composite material NF22 (Railko) against the AISI 440C

stainless steel disk under various normal loads, sliding speeds, and water

contamination show no significant mass loss under clean water lubrication due to

the material properties of the composite.

3. The mass loss and specific wear rate increase nonlinearly with the increase of water

contamination due to the complex adhesive-abrasive wear mechanism.

4. For the specific range of applied normal load, sliding speed, and level of water

contamination explored in this experimental study, water contamination has a

stronger effect on the mass loss and specific wear rate than the normal load and

sliding speed. This can be explained by the strong influence of abrasive nature of

the wear mechanism.

5. The wear in water-lubricated bearings depends on a number of factors such as

applied load, sliding speed, bearing design, and environment conditions. Therefore,

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CHAPTER 8 Experimental investigation of wear

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the purpose of this test method is to predict the relative ranking of water-lubricated

bearings materials combinations, but not to simulate wear of in service water-

lubricated bearings.

Thus, these experimental results represent a basis for further extensive experimental and

theoretical studies of the wear mechanism and the physical implications of a variation in

the material properties of composite materials used for water-lubricated engineering

applications.

The effect of water contamination and magnetic damping on the vibration and its

correlation to wear was experimentally investigated and results are presented and discussed

in the next chapter.

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CHAPTER 9 Experimental investigation of vibration–wear relationship

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CHAPTER 9 EXPERIMENTAL INVESTIGATION OF

VIBRATION–WEAR RELATIONSHIP

9.1 Introduction

Friction, vibration, and wear in machines are primarily due to the dynamic friction forces

acting in machinery. Friction-induced vibrations and wear on contact surfaces depend on

the combination of contact materials, load, sliding speed, lubrication, and contamination

(Krishna Kumar and Swarnamani, 1997).

Propeller shaft bearing systems on ships and submarines can make use of different types of

water-lubricated systems. A number of selection requirements have been developed and

provided in manufacturers’ catalogues and design manuals (THORDON, 2006,

WÄRTSILÄ, 6/09/2007, Litwin, 2009). These criteria include bearing pressure, sliding

speed, size, type of lubrication, surface finish, lubricant flow, fittings, and machine

tolerances. The essential factors to be considered when designing and selecting water-

lubricated bearing materials are possible vibration, noise, power loss, and lifespan. The

manufacturers of most bearings provide detailed life-load analysis where bearing

calculations consider design requirements such as load, lubricant type, and speed of

rotation (sliding speed). According to Chowdhury and Helali (2007), the reduction of wear

and vibration depends on interfacial conditions such as applied load, geometry, relative

surface motion, sliding speed, surface finishes, type of materials, system rigidity,

lubrication, and lubricant contamination (Chowdhury and Helali, 2007). Nevertheless, the

relationships between these main factors (friction, wear, vibration, and contamination) in

water-lubricated bearings are unknown and have generated limited attention in the

literature.

In boundary and mixed regimes occurring in water-lubricated bearings, high three-body

contact pressures elastically and plastically deform the bearings’ contact surfaces. This

leads to surface fatigue. With improvements in bearing materials and surface preparations,

surface damage due to the vibration–wear relationship from water contamination has

become one of the main challenges associated with the performance of water-lubricated

bearings. Water contains different types of contaminants, which may include solid silica

particles and swear debris material particles as a consequence of the wear itself.

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According to Maru (2007), vibration measurement is commonly used for condition

monitoring of rotating machinery, being relatively easy to use. In this context, this chapter

presents an experimental investigation into the effect of water contamination on the

vibration–wear relationship for water-lubricated bearings materials using a modified POD

test rig. The contacting characteristics of sliding bodies were considered in order to

determine the existence of correlations between tribological aspects and vibration

phenomena (Maru et al., 2007b).

The ASTM G 99-04 Standard test Method for wear testing with a Pin-on Disk apparatus

was used to collect the wear data (ASTM, 2004a).

9.2 Operational conditions and experimental methods

An experimental program was conducted to examine the effect of varying contaminated

water lubrication regimes on the vibration–wear relationship of materials for water-

lubricated bearings operating in boundary and mixed regimes. This study investigated the

lubricated sliding response of NF22 (Railko) material against stainless steel. Experiments

were conducted at room temperature for sliding speeds 0.393 m/s and 1.557 m/s under an

applied load of 8 N. To investigate the effect of damping on the relationship between

vibration and wear, a strong magnet was used as a magnetic field damper. Experiments

were repeated three times for each run for both damped and undamped conditions.

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Figure 9.1 POD test rig equipped with a magnet

Figure 9.1 shows the POD test rig equipped with a strong magnet mounted on the rig base.

The magnet is independent from the sliding load cell containing the pin. The magnet was

positioned within a millimetre of the load cell for the investigation of the vibration–wear

relationship under damped conditions.

An experimental study of the effect of water contamination and damping on the vibration–

wear relationship was conducted to examine the effect of varying contaminated water

lubrication regimes and damped and undamped conditions on mass loss, specific wear rate,

and the frequency characteristics of the induced vibrations of water lubricated bearings

materials.

The purposes of this experimental study were to:

Investigate the effect of water lubrication and contamination on the relationship

between vibration and the wear rate of materials for water-lubricated bearings

under boundary and mixed regimes of lubrication

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Investigate the effect of damping of vibrations on the relationship between

vibration and the wear rate of materials for water-lubricated bearings materials

Assess the materials’ performance under water-lubricated conditions

Identify the effect of damping on the wear mechanism under water - lubricated

damped and undamped conditions.

The technical parameters adopted for the experimental study, including specific test

conditions and experimental samples, are shown in Table 9.1.

Table 9.1 Technical parameters adopted for the vibration–wear experiments

Parameters Experimental Rig

Applied normal force 8 N

Lubrication conditions Water-lubricated damped

Water-lubricated undamped

Water contamination 0%

0.5%

1%

2%

Sliding speeds:

Boundary regime

Mixed regime

0.393 m/s

1.557 m/s

Calibration coefficient

Test duration

Disk diameter

Pin diameter

Sand particles size range

9.0 N/volt

2 hours

0.3 m

0.01 m

53-106 µm

Materials:

Disk

Pin

AISI 440C Stainless steel

NF22 (Railko)

This experimental work was conducted at room temperature. The preparation procedure

included polishing the surfaces of the pin and the disk using various grades of wet and dry

silicon carbide paper down to 1200 grit.

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Two sets of tests were conducted under both: damped and undamped conditions. The

duration of each test was one hour for vibration-wear data acquisition. Prior to the

commencement of the tests, the POD test rig was run for an hour before taking any

measurements. This was done to ensure that there was full contact between the pin and the

disk surfaces. During all vibration and wear tests, the vibration parallel to the sliding

direction was measured and the friction-induced vibration voltage output signal from the

bending arm strain gauges was recorded using a USB-1408FS data acquisition device.

For the first run of experiments, a constant sliding speed of 0.393 m/s (boundary regime)

and a load of 8 N under clean water lubrication in undamped conditions were used. During

the test, a friction-induced vibration voltage output signal from the bending arm strain

gauges was recorded at a sampling rate of 0.01 sec/sample. Before and after each run, the

mass of the pin was measured and the mass loss calculated and recorded. At the same

sliding speed, clay free sand was introduced as the contaminant. The friction-induced

vibration voltage output signal and mass losses were measured and recorded, and the

friction force, Welch power spectrum densities, and specific wear rates were calculated and

analysed. This procedure was repeated for a sliding speed of 1.557 m/s (mixed regime).

The experimental runs were repeated three times at each condition and the average value

for the three runs was recorded for each data point.

To investigate the effect of damping on vibration and wear, the same set of experiments

was conducted using a damping magnet. Since it is generally known that the specific wear

rate depends on multiple variables such as load, sliding distance, sliding speed, property of

materials, and type of lubrication and contamination, the purpose of these two sets of

experiments was to collect data for further analysis of the effect of water lubrication, water

contamination, and damping on the vibration–wear relationship.

At sliding speeds of 0.393 m/s and 1.557 m/s the two different lubrication regimes

(boundary and mixed) were characterised by the behaviour of the friction force versus

time, the nature of pin surfaces as examined under the scanning electron microscope, and

the specific wear rate measured as mass loss per unit of sliding distance. The effect of

increasing water contamination and damping on the lubrication regimes was also

characterised.

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9.3 Experimental study of effect of water contamination on vibration–wear relationship: results and discussions

9.3.1 Friction force–time analysis

Figure 9.2 shows typical damped and undamped friction force–time graphs for the NF22

(Railko) material in clean water and various levels of contaminated water lubrication. All

tests were conducted at a sliding speed of 0.393 m/s conditions, under an 8 N load and

under 0%, 0.5%, 1%, and 2% water contamination, respectively. The measurements were

obtained and recorded at the final stage of the two hours duration test.

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a) Contamination – 0.0%, sliding speed – 0.393 m/s

b) Contamination – 0.5%, sliding speed – 0.393 m/s

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c) Contamination – 1%, sliding speed – 0.393 m/s

d) Contamination – 2%, sliding speed – 0.393 m/s

Figure 9.2 Calculated friction force (N) versus time for undamped and damped

conditions under load 8 N and sliding speed 0.393 m/s

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Under clean water lubrication (see Figure 9.2 a), the process of friction can be described by

an adhesive mechanism of interaction under the boundary lubrication regime (Aronov et

al., 1983). As the water contamination is increased above a certain critical value, which in

this case is 0.5 %, a transition occurs from an adhesive to an adhesive-abrasive friction

process. This is characterised by the appearance of higher-amplitude unstable low-

frequency vibration shown in Figure 9.2 b), c), d). These unstable low-frequency

vibrations, due to the solid silica particles, are responsible for the formation of the three-

body abrasive interaction coupled with an adhesive mechanism of friction. The typical

average friction force value does not change significantly and is within the range of 2-3 N.

At the same time, damping significantly decreases the amplitude of vibrations. The

maximum vibration suppression was observed with clean water lubrication and a 0.393 m/s

sliding speed. This was explained in the literature by “disabling of adhesive wear

mechanism by damping of the amplitude of vibration” (Maru et al., 2007a, Maru et al.,

2007b).

Figure 9.3 shows typical damped and undamped calculated friction force–time graphs for

the NF22 (Railko) material in clean and various levels of contaminated water lubrication.

In this case the tests were conducted at a sliding speed of 1.557 m/s conditions, under an 8

N load and under 0%, 0.5%, 1%, and 2% water contamination, respectively. The

measurements were obtained at the final stage of the 60-minute duration test.

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a) Contamination – 0.0%, sliding speed – 1.557 m/s

b) Contamination – 0.5%, sliding speed – 1.557 m/s

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c) Contamination – 1%, sliding speed – 1.557 m/s

d) Contamination – 2%, sliding speed – 1.557 m/s

Figure 9.3 Calculated friction force (N) versus time for undamped and damped

conditions under load 8 N and sliding speed 1.557 m/s

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The friction can be described by the adhesive mechanism of interaction in the mixed

(transition) lubrication regime (Aronov et al., 1983), when clean water lubrication is used

(as shown in Figure 9.3 a). Comparing these results with those for the low sliding speed of

0.393 m/s results (shown in Figure 9.2 a), the reduction of vibration is due to a mixed

lubrication regime when a lubricant film is formed and resulting in reduced adhesive

interaction. When 0.5 % contamination was introduced, the complex three-body

mechanism of friction begins to play a significant role and an increased amplitude of

vibration was observed (Litwin, 2009, Meuter, 2006). A transition occurs from an adhesive

to an abrasive friction process, characterised by the appearance of high-amplitude unstable

low-frequency vibrations as seen in Figure 9.3 b), c), d) when the water contamination was

increased above a critical value of 1.0%. A similar effect was described my Maru et al.

(2007a) and (2007b). Similar to the low sliding speed, these unstable low-frequency

vibrations appear result of the solid silica particles producing for a three-body interaction

coupled with a mixed regime of lubrication. As the highest sliding speed is in the mixed

lubrication regime, the typical average friction force was lower compared to the low

sliding speed results and was within the range of 1-2 N. At the same time, damping

significantly decreased the amplitude of vibrations. The maximum vibration suppression

was observed for clean water-lubrication. This again can be explained by the disabling of

the adhesive wear mechanism by the damping of the amplitude of vibration as reported in

Maru et al. (2007a) and (2007b).

A significant reduction of vibration by damping was observed at 2% contamination and

1.557 m/s sliding speed. Greater amplitudes of vibrations with a normal load are due to

mixed lubrication at a high sliding speed and a higher level of water contamination when

solid silica particles are dragged between the interacting sliding surfaces of the pin and

disk. Vibration is reduced by damping and enables a stable gap between the two surfaces to

be formed during mixed lubrication (Chowdhury and Helali, 2007).

9.3.2 Power spectral density analysis

Welch power spectral density analysis (periodogram method) is used to estimate the power

of signal (vibration) at different frequencies. This method is based on converting the time

signal into frequency domain using Fourier transform.

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Welch power spectral density analyses were conducted to establish the conditions for low-

frequency vibration generation. The effects of undamped and damped conditions, water

lubrication and contamination, sliding speeds, and friction force–time histories were

analysed.

Figure 9.4 presents the Welch power spectral density values data for NF22 (Railko) for

clean and contaminated water lubrication under an 8 N load, at a 0.393 m/s sliding speed,

and at 0%, 0.5%, 1%, and 2% water contamination, respectively, with and without

damping. The measurements were taken during the final stage of a 60-minute duration test.

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a) Contamination – 0.0%, sliding speed – 0.393 m/s

b) Contamination – 0.5%, sliding speed – 0.393 m/s

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c) Contamination – 1%, sliding speed – 0.393 m/s

d) Contamination – 2%, sliding speed – 0.393 m/s

Figure 9.4 Welch power spectral densities for damped and undamped conditions

under load 8 N and 0.393 m/s, sliding speed

Figure 9.4 presents the boundary regime of lubrication with different levels of

contamination. It is clearly seen that damping reduces the dominant resonance frequency

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of the system, typically in the 12-14 Hz frequency band. As the water contamination is

increased above a certain critical value, which in this case is 0.5%, the suppression of

resonance frequencies becomes more difficult. As mentioned above, transition occurs from

an adhesive to an adhesive-abrasive friction process, characterised by the appearance of

high-amplitude unstable low-frequency vibrations (Figure 9.4 b), c), d)). These unstable

low-frequency vibrations appear as soon as the solid silica particles become responsible for

the formation of the three-body interaction coupled with the adhesive friction mechanism.

As a result of the contamination, the intensity of vibration was increased by approximately

20 db/Hz. This type of vibration (low frequency and high intensity) at a low sliding speed

was generated due to the boundary regime of lubrication (Bhushan, 1980).

Figure 9.5 presents the Welch power spectral density values data for NF22 (Railko) with

clean and contaminated water lubrication under an 8 N load, at 1.557 m/s sliding speed and

at 0%, 0.5%, 1%, and 2% water contamination, respectively, with and without damping.

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a) Contamination – 0.0%, sliding speed – 1.557 m/s

b) Contamination – 0.5%, sliding speed – 1.557 m/s

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c) Contamination – 1%, sliding speed – 1.557 m/s

d) Contamination – 2%, sliding speed – 1.557 m/s

Figure 9.5 Welch power spectral densities for undamped and damped conditions

under load 8 N and 1.557 m/s, sliding speed

Figure 9.5 represents the Welch power spectral densities in a mixed regime of lubrication

with different levels of contamination. It is clearly seen that damping significantly reduces

the dominant resonance frequencies of the system as well as self-induced low-frequency

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vibrations, typically up to 15 Hz frequency band. At high sliding speeds, a few

power/frequency peaks were observed. These frequencies appear to be directly associated

with the sliding system (mixed lubrication) and excited by an adhesive-abrasive

mechanism of interaction coupled with the mixed regime of lubrication, which is similar to

that reported by Bhushan (Bhushan, 1980).

Overall, the level of power increases with increasing contamination due to the active

abrasive three-body interaction in undamped conditions. According to Torrance (2005) and

Younes (1993), the introduction of damping significantly decreases the intensity of

vibrations with a similar tendency of the power spectral densities to increase with

increasing water contamination (Torrance, 2005, Younes, 1993).

As seen from the results, low-frequency vibrations (up to 15 Hz) play the most significant

role in the vibration–wear relationship under these experimental conditions. No high-

frequency vibrations were recorded during this set of experiments.

9.3.3 Vibration–wear analysis

For further vibration–wear relationship analysis, the root mean square (RMS) needs to be

calculated and analysed as it is one of the most important characteristics of vibration

analysis. The RMS parameter of the vibration signal was taken as a rough indicator of the

average level of vibration (Maru et al., 2007b, Maru et al., 2007a).

In the case of a set of n values , the RMS value is given by equation

9.1:

, (9.1)

where:

xn is a value of friction force recorded at a rate of 0.01 samples/sec;

n is the number of recorded samples.

According to previous results of calculated friction force – time data and Welch power

spectral densities analysis, the low frequency (LF) band is much more sensitive to the

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CHAPTER 9 Experimental investigation of vibration–wear relationship

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operational conditions, being significantly affected by water contamination and damping

(Maru et al., 2005).

Table 9.2 presents the RMS calculated values for NF22 (Railko) in water-lubricated

conditions (clean and contaminated water lubrication), under an 8 N load, at 0.393 m/s and

1.557 m/s sliding speeds, and at 0%, 0.5%, 1%, and 2% water contamination, respectively,

with and without damping.

Table 9.2 Calculated RMS acceleration values at different sliding speeds, lubrication,

and contamination conditions, under an 8 N load

Contamination,

%

Sliding speed,

m/s

Calculated RMS acceleration values

Undamped Damped

Difference

in RMS,

%

0 0.393 2.85443 1.562226 45.0

1.557 1.187191 0.748596 36.9

0.5 0.393 7.927545 4.666254 41.1

1.557 3.31157 1.445673 56.3

1 0.393 6.258963 3.179442 49.2

1.557 4.55773 2.067589 54.6

2 0.393 2.657814 2.168903 18.4

1.557 3.142426 1.927 38.7

Comparing the values of the RMS differences (∆RMS, %) between undamped and damped

conditions, it is seen that under a high level of water contamination (2%) the effect of

damping is much lower for both sliding speeds. At the same time, with an increasing

sliding speed, the RMS values reach a maximum at 0.5% contamination for a low sliding

speed and at 1% contamination at a high sliding speed. This is due to both adhesive and

abrasive wear mechanisms, which are in turn due to the transition from a boundary to a

mixed regime.

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Figure 9.6 illustrates the final experimental RMS values for NF22 (Railko) against the

stainless steel disk with various degrees of water contamination (0%, 0.5%, 1%, and 2%,

respectively), under undamped and damped conditions under a load of 8 N.

Figure 9.6 Calculated RMS values for NF22 (Railko) material at normal and damped

load (8 N)

These results show the differences in vibration levels for measured operational conditions,

obtained from tests performed under clean and contaminated lubrication. It can be seen in

Figure 9.6, that vibrations in the overall low-frequency bands are greatly affected by

applied operational conditions, especially by damping (which decreases the vibration for

all levels of contamination and at all sliding speeds), the level of contamination, and

sliding speed. For a low sliding speed, the damping is most effective and the highest value

(peak) of the RMS was recorded for 0.5% water contamination. With an increasing sliding

speed, the peak of vibration (RMS) decreased and also moved to the 1% water

contamination point. This is due to the transition from a boundary to a mixed regime (more

lubricant appears between interacting surfaces). It can be seen that vibrations in low-

frequency bands are greatly affected by operational conditions, especially by sliding speed

(a regime of lubrication), which is in strong agreement with results from the literature

(Maru et al., 2007a, Maru et al., 2007b, Maru et al., 2005). Figure 9.6 also shows that

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vibration is significantly affected by damping, especially at a low-speed range. This is due

to the lower frequency band of vibration generated.

Some of the main observations are highlighted below. As it was discussed by Maru et al.

(2007), the same trend is observed in all operational conditions in which the vibration level

increases with an increase of contaminant concentration and then decreases for higher

concentrations up to the limit value (Maru et al., 2007a). Another observation concerning

the graph (see Figure 8.6) is related to the comparison between levels of contamination. It

is seen that an increase in vibration is seen at low contamination levels. This shows a

significant change in the dynamic behaviour of the pin-disk system which is connected to

the dependency of the vibration–wear relationship on contamination. It is recommended

that these observations be investigated further.

For wear results, no measurable mass losses were recorded during tests under clean water-

lubricated conditions. In the case of clean water lubrication, only the adhesive wear

mechanism is possible, which is significantly affected by the type of material and lubricant

used. This is due to the specific material properties of the NF22 (Railko) composite which

contains a solid lubricant in its structure.

Table 9.3 presents the average mass loss values for NF22 (Railko) under water-lubricated

conditions, at 0%, 0.5%, 1%, and 2% water contamination, under an 8 N load and at

sliding speeds of 0.393 m/s and 1.557 m/s, respectively, with and without damping.

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Table 9.3 Average mass loss (g) at different sliding speeds, lubrication, and

contamination conditions, under an 8 N load

Lubrication

Contamination,

%

Sliding speed,

m/s

Average mass loss, g

Undamped Damped Difference

in W, %

Water

0 0.393 0.0000 0.0000 0.00

1.557 0.0005 0.0000 0.00

0.5 0.393 0.0043 0.0004 -3

1.557 0.0061 0.0105 +42

1 0.393 0.0064 0.0048 -33

1.557 0.0097 0.0216 +45

2 0.393 0.0118 0.0054 -55

1.557 0.0177 0.0334 +53

Comparing the values of mass loss (∆W, %) between undamped and damped conditions, it

is seen that at a low sliding speed, the damping leads to a mass loss reduction due to the

boundary regime of lubrication. At a high sliding speed, the value of mass loss increases

due to a more aggressive abrasive wear mechanism coupled with the mixed regime of

lubrication. It is also seen from Table 9.3 that the value of mass loss increases accordingly

with increased water contamination.

Figure 9.7 presents the final experimental results of calculated specific wear rates of NF22

(Railko) against the stainless steel disk under water-lubricated conditions at various

degrees of water contamination (0%, 0.5%, 1%, and 2%), respectively.

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Figure 9.7 Specific wear rate for NF22 (Railko) material for undamped and damped

conditions

The results show two different vibration–wear relationship trends exist: one for low

(boundary lubrication), and one for high (mixed lubrication) sliding speeds for the

conditions used.

The adhesive wear mechanism plays the major role for the low speed range (boundary

regime) when zero or very little lubricant film is presented between the interacting

surfaces, even if it is coupled with abrasive wear. The reduction of vibration by damping

does not allow sand particles to move between the sliding surfaces and make a strong

three-body contact. This leads to a decrease of surface three-body interaction and, as a

consequence, the specific wear rate decreases.

At high speed (mixed lubrication), there is mostly an abrasive wear mechanism due to

lubricant being forced between the interacting bodies. This allows more sand particles to

be drawn into the contact area. Damping reduces vibration, but maintains a stable gap

between the pin and the disk which allows contaminated lubricant to be more aggressive

(Maru et al., 2007a).

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In spite of much lower mass loss and insignificant differences in surface damage between

undamped and damped conditions, the maximum increase in the specific wear rate was

obtained at the slowest sliding speed (0.393 m/s) in undamped conditions. The clearest

pictures of significant difference in the pin’s surface damage were found at the high-speed

range. The microscopy of the pin surface under undamped and damped conditions at a high

sliding speed (1.557 m/s) after a full cycle of experiments is shown in Figure 9.8.

a) Pin surface before test

b) Undamped condition, contamination – 0.5% c) Damped condition, contamination – 0.5%

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d) Undamped condition, contamination – 1% e) Damped condition, contamination – 1%

f) Undamped condition, contamination – 2% g) Damped condition, contamination – 2%

Figure 9.8 Microscopy of the pin’s worn surface before and after a full cycle of

experiments at a high sliding speed of 1.557 m/s, at a magnification of X30

Regardless of vibration suppressions by damping and different levels of contamination, the

photos in Figure 9.8 show greater surface damage for damped conditions than for

undamped conditions. As discussed previously by Torrance (2005), the process of wear

depends on the mechanics of the contact between the two surfaces employing the abrasive

particles, on particle size and contamination, and often on chemical interactions between

the wearing surface and the surrounding media (Torrance, 2005). At a high sliding speed

range, when a film of lubricant is formed and damping suppresses vibration, more sand

particles move between the pin and the disk due to the mixed regime of lubrication. The

presence of an increased quantity of sand particles between the sliding surfaces gives rise

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to additional forces that strongly depend on the size and material properties of the

contaminants (Olsson et al., 1997).

9.4 Conclusions

In this chapter, water-lubricated bearings materials were experimentally investigated in

order to study the effect of water contamination on vibration and the correlation of

vibration to wear. The methods of vibration analysis (Welch power spectral analysis and

RMS versus specific wear rate) were effective in characterising the trends in the vibration–

wear relationship due to water contamination.

The effect of water contamination on the vibration–wear relationship was recognisably

different in nature from that of the level of contamination. The level of vibration (RMS)

increased with the contamination level, and then stabilising at a specific limit. At the same

time, when water contamination was increased, the specific wear rate also increased.

Comparing the results for undamped and damped conditions, it can be seen that damping

can reduce vibration and wear due to adhesive wear at the lower range of sliding speeds,

which represents the boundary lubrication regime. At the same time, it is apparent that due

to the mixed lubrication regime, mass loss and specific wear rate increase by adding

damping at the higher range of speeds. The vibrations are suppressed by damping; sand

particles are free to move into the contact area between the pin and the disk due to the

influence of abrasive wear, sliding speed, and the formation of a lubrication film between

the interacting surfaces.

It is expected that, with a change of particle size, the character of the vibration–wear

relationship could be changed significantly due to the prevailing nature of wear

mechanisms. The results obtained can be used for further investigation into the effect of

damping, water contamination, and the type of contaminant on the vibration–wear

relationship. Further theoretical and experimental work and a theoretical three-body

vibration–wear model for simulation of the vibration–wear relationship under damping and

water contamination for stability analysis and design can be undertaken based on the

results presented here. The task of obtaining a theoretical three-body vibration–wear model

for this type of contact is difficult due to the complexity of this mechanism, which includes

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abrasive wear as well as adhesive wear. This task can be the centrepiece for a future

theoretical study.

As mentioned above, the specific wear rate in water-lubricated sliding bodies depends on

the vibration–wear relationship due to asperities’ contact, vibration, and abrasive

interaction between contacting bodies and the contaminant. As seen from the results, the

vibration–wear relationship due to undamped or damped conditions and water

contamination is a nonlinear function of the degree of contamination. The character of this

function is still not known, nor is it known how it can be numerically depicted. The present

methodology can be adopted for further experimental and theoretical investigations which

need to be undertaken to identify the complex vibration–wear relationship for different

types of materials, contaminants, and operational conditions.

It is well known that the vibration–wear process involves fracture, tribochemical effects,

and plastic deformation. Transitions between regions dominated by each of these

commonly give rise to changes in the vibration–wear relationship with undamped or

damped conditions and contamination (Song, 2008). Furthermore, these results are closely

related to structural characteristics and chemical effects occurring in frictional processes,

as well as transfer film formation on the counterface as described in the literature (Unal et

al., 2004, Maru et al., 2007b, Bryant and York, 2000, Ling Wu et al., 2010, Maru et al.,

2007a, Krishna Kumar and Swarnamani, 1997, Chowdhury and Helali, 2007). For the

NF22 (Railko) material tested in this experimental investigation and for sliding speeds of

0.393 m/s and 1.557 m/s, the sliding speed and load values as well as contamination have

shown significant influence on the vibration–wear relationship. They represent a complex

function of all of these variables. Friction-induced vibrations have been found to be a

function of sliding speed, operational (undamped or damped) conditions, and load. They

are produced by a complex mechanism of adhesive-abrasive wear at the bearing interfaces.

According to Krishna Kumar and Swarnamani (1997), this is due to the failure of

lubrication and other wear modes that can be considered to be dependent on the nature of

the surfaces and their physical properties (Krishna Kumar and Swarnamani, 1997). Those

physical properties can include surface roughness, hardness, and elastic modulus.

The main purpose of this vibration–wear experimental study was to analyse the influence

of vibration damping and water contamination on the vibration–wear relationship for NF22

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(Railko) bearing material in low and highly-abrasive conditions against a stainless steel

counterface material.

The effect of damping and water contamination on the vibration–wear relationship of a

pair of materials comprising an NF22 (Railko) composite and AISI 440C stainless steel

was experimentally investigated and shows very interesting results. In general terms, the

data obtained can be used to explained this relationship in terms of the complex three-body

adhesive-abrasive mechanism of the interaction of interacting materials, as a function of

load, sliding speed, sliding distance, and lubrication (Krishna Kumar and Swarnamani,

1997).

The following conclusions can be drawn from these experimental results:

1. Specific wear rate of NF22 (Railko) composite material decreases when damping is

applied at a low sliding speed during a boundary lubrication regime.

2. Specific wear rate of NF22 (Railko) composite material increases when damping is

applied at a high sliding speed during a mixed lubrication regime.

3. Vibration–wear experimental studies of the composite material (NF22 [Railko])

against an AISI 440C stainless steel disk under constant load and either undamped

and damped conditions, at low and at high sliding speeds, show no mass loss under

clean water lubrication.

4. Vibration increases and then decreases and specific wear rate increases with greater

water contamination. The highest value of RMS for the low sliding speed was

recorded at 0.5% contamination and for the high sliding speed at 1% water

contamination.

5. For the specific applied load of 8 N, at low and high sliding speeds, and with the

water contamination explored in this experimental study, damping had a stronger

effect on specific wear rate. It is a function of sliding speed and a result of the

change in lubrication regime between selected sliding speeds.

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6. The vibration-wear relationship in water lubricated-bearings depends on multiple

factors that include applied load, sliding speed, bearing design, and environmental

conditions. Thus, the value of this POD test method is to predict the relative

ranking of materials pairs and not for modelling vibration-wear for in service water-

lubricated bearings.

7. Future experimental studies must include an examination of the effect of active

magnetic dampers to facilitate adequate water-lubricated bearings materials

performance.

Thus, the POD experimental methodology and the results presented can be used for further

extensive experimental and theoretical studies of the vibration–wear relationship of various

combinations of materials with different material properties used for water-lubricated

engineering applications.

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CHAPTER 10 Conclusions

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CHAPTER 10 CONCLUSIONS AND RECOMMENDATIONS

10.1 Summary

The main aim of these experimental investigations was to develop new methods for the

experimental investigations of the tribological characteristics of water-lubricated bearings

materials using Pin-on-Disk test rig.

A review of the literature showed that the effect of water contamination on water-

lubricated bearing materials and systems performance is highly complex. It depends on

many factors such as the chemical reactivity and corrosion characteristics of the materials

to water and any dissolved salts it may contain and the abrasive nature of soil and sands

entrained in the water. Limited attention has been given in the literature in terms of

analysing and predicting the effect of water contamination on tribological characteristics of

materials for the design and modelling of water-lubricated bearings. Previous experimental

studies, conducted at the School of Mechanical Engineering, at the University of Adelaide,

were analysed, and used as a starting point for this research.

To measure the friction coefficient, vibration, and wear due to the effect of water

lubrication and water contamination, the operational conditions and test rig design

requirements were identified, a new experimental approach was developed, and a specific

Pin-on-Disk (POD) experimental test rig was designed and built. This experimental

apparatus enabled the measurement of the local friction coefficient and vibration as well as

wear investigation for the material combinations used in water-lubricated engineering

applications.

Under the initial experimental requirements, the pair of materials chosen was an NF22

(Railko) composite and stainless steel (AISI 440C).

In accordance with the initial requirements and operating conditions, the POD test rig was

built and ordinary tap water was introduced as a lubricant. To model contamination, fine

sweeping sand with a particle sizes range between 53-106 µm was chosen as a contaminant

and added to the lubricant.

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CHAPTER 10 Conclusions

145

Experimental investigations of friction and wear were conducted under water lubrication

with a contamination range of 0% to 6% by weight for an applied load range of 8 N to 46

N and a sliding speed range from 0.393 m/s to 1.557 m/s.

The vibration–wear relationship was experimentally investigated under clean and

contaminated water lubrication with a contaminant range of 0% to 2%. For an applied load

of 8 N, the sliding speeds were 0.393 m/s and 1.557 m/s and undamped and damped

conditions were applied.

An analysis of friction, vibration, and wear test experimental results was conducted and

showed the significant effect of water contamination and damping on the friction

coefficient, vibration, wear, and calculated specific wear rate. It was found that there was a

significant increase in the friction coefficient, vibration, and specific wear rate at the

slowest sliding speed of 0.393 m/s. This is due to the boundary regime of lubrication, the

adhesive-abrasive wear mechanism, and specific material properties of NF22 (Railko)

material. At the same time, wear can be decreased by the damping of vibration at slow

sliding speeds.

The microscopy analysis of the pin’s surface was undertaken before and after wear

experiments and confirmed that wear resistance can be significantly improved by adding

organic fibres and solid lubricant to composite bearing materials. Two components of

complex wear mechanisms (adhesive and abrasive) were considered and analysed. It was

shown that material mass loss and specific wear rate depend mostly on abrasive wear.

As mentioned above, boundary and mixed regimes are the reasons for many tribological

problems such as vibration, power loss, and excessive wear. The simulation of friction in

boundary and mixed regimes is difficult due to their strong and complex interdependency.

To acquire an adequate prediction of material performance, for water-lubricated

applications, friction, vibration, and wear, and theoretical models in the boundary and

mixed lubrication regimes are still required and could be the next stage of further water-

lubricated bearing modelling development. As a result of the analysis presented here, it is

suggested that additional components responsible for water contamination and damping

should be added to existing theoretical friction models. The three-body wear model can be

modified, based on the results obtained, as a response to the abrasive nature of water

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CHAPTER 10 Conclusions

146

contamination. In addition, this new POD experimental approach can be used during

development and selection of new materials for water-lubricated bearings.

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CHAPTER 10 Conclusions

147

10.2 Conclusions

In this experimental study, the effect of water contamination on the friction behaviour of a

pair of materials comprising a composite and stainless steel was experimentally

investigated. The following conclusions can be drawn from the results regarding the

friction coefficient, vibration, and wear for water-lubricated bearings materials:

Limited attention has been given in the literature to the effect of water

contamination on friction, vibration, and wear in water-lubricated bearings

materials and systems.

The friction coefficient of NF22 (Railko) composite material against AISI 440C

stainless steel decreases, when the load and sliding speed are increased under clean

water lubrication. This is a result of boundary and mixed regimes of lubrication and

the elastic and plastic deformation between the surfaces.

No mass loss was observed under both dry friction and clean water lubrication for a

composite material (NF22 [Railko]) against an AISI 440C stainless steel disk under

constant load, undamped and damped conditions, and low and high sliding speeds.

Mass loss and specific wear rate increase nonlinearly with an increase in the water

contamination level. In this experimental work, a simple empirical linear model of

contact mechanics, based on experimental results, has been offered for the range of

1% to 6% levels of water contamination as a first iteration.

For the specific range of loads and sliding speeds investigated in this experimental

study, load has a stronger effect than sliding speed on the friction, vibration, and

wear behaviour of composite material.

Damage to the frictional surface is indicative of boundary and mixed lubrication

regimes, as well as three-body abrasive wear. It is caused by the breakdown of the

lubricant film and by the formation and shearing of the adhesive junctions, as well

as the significant influence of abrasive three-body contacts.

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CHAPTER 10 Conclusions

148

Under a boundary lubrication regime, the specific wear rate of NF22 (Railko)

composite material decreases when damping is applied at a low sliding speed.

Under a mixed lubrication regime, the specific wear rate of NF22 (Railko)

composite material increases when damping is applied at a high sliding speed.

For the vibration–wear relationship, it was found that the greatest value of RMS for

a low sliding speed was recorded at 0.5% water contamination and for a 1.557 m/s

sliding speed at 1% water contamination on account of the transition from a

boundary to a mixed lubrication regime.

For the specific applied load of 8 N, the low and high sliding speeds, and the water

contamination explored in this experimental study, damping has a stronger effect on

the vibration–wear relationship which depends on sliding speed and, as a result, on

the lubrication regime.

The tribological characteristics of water-lubricated bearings materials depend on

factors such as applied load, sliding speed, bearing design, and environmental

conditions. Therefore, the value of these POD test methods is to predict the relative

ranking of water-lubricated bearings materials combinations during the design,

development, or selection process and not to simulate all in-service water-

lubricated bearings.

However, the analysis, the experimental results, and the experimental approach

presented here can be used for further extensive experimental and theoretical

studies for design simulation, and for the further development of a theoretical

model for the adequate prediction of the physical implications of a variation in the

operational conditions, material properties, and technical requirements of material

combinations used for water-lubricated engineering applications.

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CHAPTER 10 Conclusions

149

10.3 Recommendations for future work

A proper understanding of the tribological behaviour of water-lubricated contact surfaces

for the design and development of new water-lubricated bearings is important. It is well

known that the effect of water contamination on the performance of water-lubricated

bearings is a complex mechanism and that it is extremely difficult to model. The results

from this experimental investigation allow certain recommendations to be made that will

be helpful in establishing future theoretical and experimental work to develop further

design approaches and theoretical models for reliable and quiet water-lubricated bearings.

Further progress in the theoretical and experimental study of friction in water-lubricated

bearings materials and systems will require the following:

Experimental and theoretical investigation of the effect of water contamination on

friction-induced vibration and wear. A set of experiments needs to be conducted to

further investigate the effect of water contamination on friction-induced vibrations.

Sources and frequency characteristics of noise need to be identified and analysed.

Further developments of theoretical nonlinear models of friction, vibration, and

wear are required to predict the effect of contamination on material performance in

water-lubricated bearings.

A new theoretical three-body model of friction, vibration, and wear using contact

mechanics is necessary for the analysis and design of water-lubricated bearings

under water contamination. It is still not known if any nonlinearity is due to the

effect of water contamination.

Proper material selection for experimental work. For example, all materials need to

be either properly analysed and/or all material property specifications should be

available in detail or determined.

Operational conditions, including the water lubrication regimes and surface

roughness, should be fully identified.

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APPENDIX A: PIN-ON-DISK ASSEMBLY DRAWINGS

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