QUEENSLAND UNIVERSITY OF TECHNOLOGY
SCHOOL OF ENGINEERING SYSTEMS
WEAR REDUCING ADDITIVES FOR LUBRICANTS CONTAINING SOLID
CONTAMINANTS
SUBHASH CHANDRA SHARMA
B.E. (Mech.), M. Tech. (Mech.)
Principal supervisor: Prof. Doug Hargreaves
Associate Supervisor: Prof. Will Scott
Submitted to Queensland University of Technology for the Degree of
Doctor of Philosophy
2008
ii
BLANK
iii
ABSTRACT
Machines operating in dusty environments, such as mining and civil works, are prone
to premature failure, leading to production losses. To address this problem, this
research project examines the interaction between solid contaminants and the bearing
micro-geometry, in lubricated surface contacts. In particular, it seeks to identify anti-
wear additives that are effective in reducing wear under abrasive conditions, making
machine elements more dirt tolerant.
In general, the influence of antiwear additive is so small that it is difficult to isolate
it. Manufactures often make claims about their antiwear products, which are difficult
to verify. Hence, there is a need to characterising the antiwear additives available
with a well-defined parameter, making it easier for consumers to compare the
efficacy of various additives, and be able to select the most suitable additive for a
given environment.
Effect of micro-geometry parameters such as radial clearance, out-of-roughness and
surface roughness was examined and a Film Shape Factor (FSF) – also termed
gamma ratio – has been proposed for ensuring adequate separation of journal
bearings operating in hydrodynamic lubrication regime, where the out-of-roundness
values are higher than the surface roughness values.
In this research, an experimental study has been conducted on journal bearings, to
examine the influence of five antiwear additives on the bearing wear and micro-
geometry. The test additives were provided by the industry partner without revealing
their chemical identity or composition; however, these included some of the most
commonly used antiwear additives. The tests were performed under three conditions:
pure base oil, base oil containing contaminants, and base oil containing contaminants
treated with five different additives.
The experiments were aimed at choosing one wear measuring technique that
evaluates the performance of an individual additive reliably, and based on this
technique the additives were characterised. To achieve these objectives, a multi-wear
parameter approach (MWPA) was developed, which employed three main wear
iv
measurement methodologies, i.e. weight loss, micro-geometry and particle counts –to
examine the effect of the antiwear additives. Minimum oil film thickness was also
measured to study the lubrication status in the bearing contacts. The MWPA helped
in comparing different wear measuring methods, and in selecting the most reliable
one. This approach also helped in developing short duration wear tests, thereby
saving time, while still getting reliable results without repeating these.
Wear experiments were performed on seven sets of bronze bearings and steel sleeve
shafts. The test contaminant was 16 micron Aluminium oxide Al2O3 powder mixed
in oil with 4% concentration by weight. These solid contaminants were treated with
five different antiwear additives to study their influence on the bearings. Bearings
were operated such that the minimum oil film thickness in the bearing was equal to
the size of the contaminants. These tests were run for a constant sliding distance of
7536m.
The results showed that most of the wear measuring techniques do not suit heavily
contaminated test conditions. However, the out-of-roundness technique proved to be
the most reliable and practical. Based on this technique a methodology was
developed which gave a wear characteristic number (N). A unique value of N can be
derived for each additive, thereby ranking the additives for their efficacy.
The finding of this research provides a better understanding of the methodologies
used for measuring wear in journal bearings subjected to dusty environments, and
examines the efficacy of each one of these. The wear characteristic number (N) can
be used by manufacturers with support from international standards organisations, so
that the users can confidently choose the most appropriate antiwear additive for their
application.
Machines operating in a dusty environment, such as mining industry and civil works
are prone to premature failure with subsequent production losses. In response to this
problem, this research project examines the interaction between solid contaminant
particles and the lubricant film micro-geometry in lubricated surface contacts. In
particular, it seeks to identify lubricant anti-wear additives, which are effective in
reducing wear under abrasive conditions and thus making machine elements more
dirt tolerant.
v
Table of Contents
ABSTRACT ......................................................................................................................... iii
LIST OF PUBLICATIONS.................................................................................................xiv
STATEMENT OF ORIGINALITY.......................................................................................xv
KEY WORDS .....................................................................................................................xvi
ACKNOWLEDGEMENT..................................................................................................xvii
NOMENCLATURE AND SYMBOLS.............................................................................. xviii
CHAPTER-1 .......................................................................................................................................... 1
1. INTRODUCTION ............................................................................................................ 1
1.1 Rationale................................................................................................................................. 1
1.2 Background Research ............................................................................................................. 2
1.3 Project Motivation .................................................................................................................. 2
1.4 Current Scenario ..................................................................................................................... 4
1.5 Knowledge Gaps..................................................................................................................... 4
1.6 Aims and Objectives of the Research ..................................................................................... 4
1.7 Research Methodology ........................................................................................................... 5
1.7.1 Effect of anti-wear additives on bearing wear ................................................................ 6
1.7.2 Effect of micro-geometry on tribological performance .................................................. 7
1.7.3 Characterisation of anti-wear additives........................................................................... 8
1.8 Contribution to the Body of Knowledge................................................................................. 8
1.9 Organisation of the Thesis .................................................................................................... 10
CHAPTER 2 ........................................................................................................................................ 11
2. LITERATURE REVIEW................................................................................................. 11
vi
2.1 Overview ...............................................................................................................................11
2.2 Contaminant Effects on Wear................................................................................................12
2.2.1 A general view of abrasive wear....................................................................................14
2.2.2 Abrasive wear in sliding bearings..................................................................................17
2.2.3 Micro polar lubricant effects in bearing lubrication ......................................................17
2.3 Effect of Solid Contaminants on Journal Bearing Performance ............................................18
2.3.1 Effects of abrasive hardness on wear.............................................................................25
2.3.2 Contaminant motion in lubricated contact .....................................................................27
2.4 Anti-wear Additives and Performance Characterisation .......................................................29
2.4.1 Commonly used antiwear additives ...............................................................................31
2.4.2 Lubricated wear and characterisation of additives.........................................................38
2.4.3 Bearing performance measurement techniques .............................................................40
2.5 Effects of Micro-geometry on Bearing Performance.............................................................41
2.5.1 Roughness effects in hydrodynamic bearings................................................................41
2.5.2 Worn journal bearing analysis .......................................................................................44
2.6 Knowledge gaps ....................................................................................................................48
2.7 Conclusion.............................................................................................................................48
CHAPTER-3 .........................................................................................................................................51
3. EXPERIMENT DESIGN AND DEVELOPMENT..........................................................51
3.1 Overview ...............................................................................................................................51
3.2 Identification of Performance Parameters .............................................................................52
3.2.1 Parameters as measure of energy conservation..............................................................53
3.2.2 Parameters as a measure of bearing life.........................................................................54
3.3 Journal Bearing Design .................................................................................................................. 55
vii
3.4 Contaminant Selection and Characterization........................................................................ 59
3.5 Lubricant and Additive Selection ......................................................................................... 60
3.5.1 Base oil ......................................................................................................................... 61
3.5.2 Additives selection........................................................................................................ 61
3.6 Test Rig and Instrumentation................................................................................................ 63
3.7 Multi-Wear Parameter Approach (MWPA).......................................................................... 66
3.7.1 Weight loss ................................................................................................................... 67
3.7.2 Out-of-roundness .......................................................................................................... 67
3.7.3 Radial clearance measurements .................................................................................... 68
3.7.3.1 Metrological issues .................................................................................................... 72
3.7.4 Gamma ratio for film thickness measurement .............................................................. 76
3.7.5 Bearing Component Roughness.................................................................................... 82
3.7.6 Maximum Wear Depth ................................................................................................. 82
3.7.7 Particle Counts in Oil Sample....................................................................................... 83
3.7.8 Minimum Oil Film Thickness Measurement ................................................................ 83
3.8 Trigonometric Solution of Film Thickness Measurement .................................................... 92
3.9 Test Procedure and Experiment Design................................................................................ 95
3.10 Conclusion .......................................................................................................................... 96
CHAPTER-4 ........................................................................................................................................ 99
4. EXPERIMENTAL RESULTS AND ANALYSIS............................................................. 99
4.1 Overview .............................................................................................................................. 99
4.2 Weight Loss........................................................................................................................ 103
4.3. Out-of-roundness ............................................................................................................... 106
4.4 Radial Clearance................................................................................................................. 113
viii
4.5 Change in Roughness ..........................................................................................................117
4.5.1 Bearing roughness........................................................................................................118
4.5.2 Shaft sleeve roughness.................................................................................................121
4.5.3 Individual additive effects............................................................................................123
4.5.4 Roughness traces of bearing elements .........................................................................128
4. 6 Particle Counts (PC) ...........................................................................................................132
4.6.1 Gravimetric change (PCg) ............................................................................................140
4.7 Maximum Wear Depth (WDmax) .........................................................................................142
4.8 Changes in Minimum Oil Film Thickness (Hmin) ................................................................144
4.9 Comparative Analysis of Techniques and Results...............................................................147
4.9.1 Methodologies .............................................................................................................148
4.9.2 Additive performance ..................................................................................................150
4.10 Conclusion.........................................................................................................................151
CHAPTER 5 .......................................................................................................................................153
5. CHARACTERISATION OF ANTIWEAR ADDITIVES.................................................153
5.1 Overview .............................................................................................................................153
5.2 Wear Measurement by Weight Loss....................................................................................154
5.2.1 Wear Computation from Out-of-roundness Trace .......................................................155
5.2.2 Computation of Cross Sectional Wear Area (CSWA).................................................157
5.2.3 Wear Characteristic Equation ......................................................................................165
5.2.4 Area measurement by Newton Cotes method..............................................................167
5.3 Wear Assessment from Out-of-roundness Traces ...............................................................169
5.3.1 Maximum wear depth ..................................................................................................169
5.3.2 Computed wear volume (V) .......................................................................................................171
ix
5.3.3 Computed weight loss (W).......................................................................................... 172
5.4 Characterisation of Additives ............................................................................................. 174
5.4.1 Wear Characteristic Number (N) ................................................................................ 175
5.5 Discussion on Results ......................................................................................................... 177
5.6 Conclusion.......................................................................................................................... 178
CHAPTER 6 ...................................................................................................................................... 181
6. CONCLUSIONS .......................................................................................................... 181
6.1 Problem Statement.............................................................................................................. 181
6.2 Literature Review ............................................................................................................... 182
6.3 Experiment Design and Development ................................................................................ 184
6.4 Results and Analysis........................................................................................................... 186
6.5 Characterisation of Antiwear Additives.............................................................................. 187
6.6 Research Contributions....................................................................................................... 188
CHAPTER 7 ...................................................................................................................................... 191
7. SCOPE FOR FUTURE WORK ................................................................................... 191
8.0 REFERENCES........................................................................................................... 193
APPENDICES ................................................................................................................................... 203
SUMMARY OF APPENDICES ....................................................................................... 203
APPENDIX A- PUBLICATIONS..................................................................................... 205
APPENDIX –B BEARING FORTRAN PROGRAM........................................................ 215
APPENDIX –C EXAMPLES OF ESDU BEARING OUTPUT)....................................... 225
APPENDIX-D MICROGRAPHS (SURFACE IMAGES ........................................................ 235
APPENDIX-E OUT–OF–ROUNDNESS TRACES)........................................................ 242
APPENDIX F –ROUGHNESS TRACES OF BEARINGS AND SHAFT SLEEVE........... 251
x
LIST OF FIGURES
Figure 2.1 Classification of abrasive wear, Misra & Finnie (1980) .......................................15 Figure.2.2 Effect of abrasive hardness on wear rate, Czichos (1978) ..................................26 Figure.2.3 Particle Motion in bearing contact (William and Hyncica (1992) ......................28 Figure 2.4 Clearance ratio and film thickness relationship Chu (1974).................................45 Figure 2.5 Out of roundness magnified part of the edge, Bagnel (1978) ...............................46 Figure 3.1 Bearing and journal drawing.................................................................................56 Figure 3.2 SEM micrograph of Aluminium Oxide particles .................................................60 Figure 3.3 EDAX elemental analysis of Al2 O3 .....................................................................61 Figure 3.4 Test rig assembly ..................................................................................................62 Figure 3.5 Loading System ....................................................................................................64 Figure 3.6 Oil Circuit .............................................................................................................65 Figure 3.7 Multi-Wear Parameter Approach (MWPA)..........................................................68 Figure 3.8 Talyrond 100.........................................................................................................69 Figure 3.9 Metroscope for bearing ID measurements ............................................................70 Figure 3.10a Bearing ID profile by Vernier measurements .................................................73 Figure 3.10b Bearing ID profile by Sigmascope....................................................................73 Figure 3.10c Bearing ID profile by Hole-test-gauge..............................................................74 Figure 3.10d Bearing ID profile by Metroscope ....................................................................74 Figure 3.10e Concentric bearing and shaft sleeve diameter graphs .......................................75 Figure 3.11a Oil film thickness between the surfaces ............................................................77 Figure 3.11b Oil film thickness based on composite roughness ............................................77 Figure 3.12 Film thickness based on composite out-of-roundness concept ...........................79 Figure 3.13 Film thickness based on out-of-roundness concept ............................................80 Figure 3.14 Quant Alert..........................................................................................................84 Figure 3.15 Flow chart ...........................................................................................................87 Figure 3.16a Probe calibration fixture....................................................................................90 Figure 3.16b Calibration setup ...............................................................................................91 Figure 3.17a Calibration chart of probe 1 ..............................................................................91 Figure 3.17b Calibration chart of probe -2.............................................................................92 Figure 3.18 Geometrical representation of film thickness measurement ...............................94 Figure 4.1 Weight loss in bearings .......................................................................................103 Figure 4.2a After Test A2 bearing surface (X50)................................................................104 Figure 4.2b Micrograph of bearing surface after Test A7 (X100) .......................................105
xi
Figure 4.3 Weight loss in shaft sleeves................................................................................ 105 Figure 4.4 Change in out-of-roundness of bearings............................................................. 107 Figure 4.5a Bottom end out-of-roundness before Test A2 ................................................. 109 Figure 4.5b Bottom end out-of-roundness (convex graph) after Test A2........................... 109 Figure 4.5c Top end out-of-roundness before Test A3....................................................... 110 Figure 4.5d Middle position out-of-roundness before Test A6........................................... 110 Figure 4.5e Bottom end out-of-roundness before Test A6 ................................................ 111 Figure 4.5f Top end out-of-roundness of bearing after Test A6 ........................................ 111 Figure 4.5g Middle position out-of-roundness of bearing after Test A6 ............................ 112 Figure 4.5h Bottom end out-of-roundness Test A6 ............................................................ 112 Figure 4.6 Shaft sleeve trace inside the bearing out-of-roundness trace............................. 113 Figure 4.7 Change in radial clearance of bearings.............................................................. 115 Figure 4.8 Changes in bearing element geometry................................................................ 116 Figure 4.9a Change in bearing circumferential roughness.................................................. 119 Figure 4.9b Change in bearing transverse roughness.......................................................... 120 Figure 4.10a Change in shaft sleeve circumferential roughness........................................ 121 Figure 4.10 b Change in shaft sleeve transverse roughness................................................ 122 Figure 4.11a Roughness effects after Test A1 ................................................................... 123 Figure 4.11b Roughness effects after Test A2..................................................................... 124 Figure 4.11c Roughness effects after Test A3 .................................................................... 125 Figure 4.11d Roughness effects after Test A4................................................................... 126 Figure 4.11e Roughness effects after Test A5 .................................................................... 126 Figure 4.11f Roughness effects after Test A6..................................................................... 127 Figure 4.11g Roughness effects after Test A7.................................................................... 127 Figure 4.12a Bearing circumferential roughness before Test A5 ...................................... 128 Figure 4.12b Bearing circumferential roughness after Test A5........................................... 128 Figure 4.12c Bearing transverse roughness before Test A5 ............................................... 129 Figure 4.12d Bearing transverse roughness after Test A5 .................................................. 129 Figure 4.12e Shaft sleeve roughness before Test A5.......................................................... 130 Figure 4.12f Shaft sleeve roughness after Test A5 ............................................................ 131 Figure 4.12 g Shaft sleeve transverse roughness before TestA5 ....................................... 131 Figure 4.12 h Shaft sleeve transverse roughness after Test A5 ......................................... 132 Figure 4.13 Comparison of change in counts for different tests ......................................... 134 Figure 4.13a Change in counts after Test A1...................................................................... 135 Figure 4.13b Changes in counts after Test A2.................................................................... 136 Figure 4.13c Changes in counts after Test A3 .................................................................... 137
xii
Figure 4.13d Changes in Counts After Test A4 ..................................................................137 Figure 4.13e Changes in counts after Test A5 ....................................................................138 Figure 4.13f Change in counts after Test A6 ......................................................................138 Figure 4.13g Changes in counts after Test A7 ....................................................................139 Figure 4.13h Change in total particle count .........................................................................139 Figure 4.14 Comparison of wear particles changes.............................................................140 Figure 4.15 Changes in total weight of contaminants ..........................................................142 Figure 4.16 Changes in maximum wear depth....................................................................143 Figure 4.17 Reduction in minimum oil film thickness........................................................146 Figure 5.1 Wear profile of a worn bearing ...........................................................................155 Figure 5.2 Roundness measurements locations....................................................................156 Figure 5.3 Wear zone shape .................................................................................................157 Figure 5.4 Out-of-roundness ‘before test trace’ .................................................................159 Figure 5.5 Out-of-roundness ‘after test trace’ .....................................................................160 Figure 5.6 Computed out-of-roundness shape of a worn bearing ........................................161 Figure 5.7 Actual trace of a test bearing with redrawn shape .............................................162 Figure 5.8 Wear depth measurement at different nodes (wndn)...........................................163 Figure 5.9 Wear Characteristic Equation for Test A2..........................................................167 Figure 5.10 Wear Characteristic Equation for Test A3........................................................167 Figure 5.11 Comparison of maximum wear depth..............................................................171 Figure 5.12 Comparison of computed weight loss and measured weight loss.....................174
xiii
LIST OF TABLES
Table 2.1: Sources of Solid Contaminants, Dwyer-Joyce (1993).......................................... 13 Table 2.2 Wear reducing properties of various lubricants, Zheng (1986) ............................. 32 Table 2.3 Results for infinitely wide bearing (Christensen 1969-70) .................................... 42 Table 3.1 Sample of test parameters ...................................................................................... 57 Table 3.2 Hardness measurements on bearing and shaft sleeves........................................... 58 Table 3.3 Antiwear additive properties.................................................................................. 63 Table 3.4 Bearing ID measurements...................................................................................... 71 Table 3.5 Bearing ID measurements: statistical analysis...................................................... 72 Table 3.6 Roughness and out-of-roundness data of test bearings .......................................... 81 Table 3.7 Operating parameters and experiment design........................................................ 96 Table 4.1 Initial measurements before the tests ................................................................... 101 Table 4.2 Experimental results ............................................................................................ 103 Table 4.3 Rise in particle counts of different sizes ............................................................. 133 Table 4.4 Changes in measured minimum oil film thickness ............................................ 145 Table 4.5 Comparison of performance of antiwear additives .............................................. 150 Table 5.1 Computed maximum wear depth data ................................................................. 164 Table 5.2 Comparison of maximum wear depth.................................................................. 170 Table 5.3 Wear Volume....................................................................................................... 173 Table 5.4 Wear Coefficients of antiwear additives.............................................................. 177
xiv
LIST OF PUBLICATIONS
1. Sharma, S., C. and Hargreaves, D (2001), “Effect of Solid Contaminants on Journal Bearing Performance”, World Tribology Conference, Vienna, pp.1-4
2. Sharma, S. C., Hargreaves, D. and Scott, W., (2004), Influence of Errors in Measuring the Radial Clearance of Journal Bearing Performance. 1st International Conference on Advanced Tribology, Singapore. pp.1
3. Sharma, S., Hargreaves, D., Scott, W., (2008), “Journal bearing metrology and manufacturing issues”, 9th Global Congress on Manufacturing and management (GCMM 2008), 12-14 November 2008, pp (paper accepted).
4. Sharma, S., Hargreaves, D., Scott, W., (2008), “Characterisation of antiwear additives”, 9th Global Congress on Manufacturing and management (GCMM 2008), 12-14 November 2008, pp. (abstract accepted- paper to be published in The Journal of Computational materials and Surface Engineering).
5. Sharma, S., Hargreaves, D., Scott, W. (2008), “Characterisation of additives using out-of roundness traces”, 2nd International Conference on Advance Tribology 2008 (ICAT 2008), 3-5 December 2008, Singapore (paper accepted)
xv
STATEMENT OF ORIGINALITY
The work contained in this thesis has not been previously submitted to meet
requirements for an award at this or any other higher education institution. To the
best of my knowledge and belief, the thesis contains no material previously
published or written by another person except where due reference is made.
Signed ……………………………….
(Subhash Chandra Sharma)
Date ……………………………..
xvi
KEY WORDS
Journal bearing, Contaminants, Antiwear additives, Worn journal bearings, Bearing
metrology, Hydrodynamic lubrication, Sliding bearing wear, Micro-geometry,
specific film thickness.
xvii
ACKNOWLEDGEMENT
I would like to express deep gratitude to my principal supervisor, Prof. Doug
Hargreaves for his consistent academic and administrative support. I am thankful to
my former principal supervisor and current Associate Supervisor, Professor Will
Scott who introduced me to the field of contamination in lubrication and motivated
me to find innovative experimental techniques to solve the problems encountered
throughout the project.
I am thankful to our industry partners Fuchs Australia and specifically Mr. Phil
Leeming who supported the project throughout its duration by supplying additives
and information related to industrial applications.
I am thankful to Mr. David McIntosh for providing me full technical support and a
jovial environment in the laboratory during the testing work. I am thankful to Mr.
Terry Beach and Mr. Mark Haynes for helping me in fabrication and instrumentation
work required for the test rig. Thanks are due to Mr. David Allen, Mr Steve Behari,
Mr Erwin Schilling and Mr. Alf Small for their technical support in instrumentation
and testing time to time.
I am thankful to Prof. Eric Hahn, University of New South Wales for his valuable
suggestions in modelling and to Prof. Nalin Sharda, of Victoria University for the
useful discussions in presenting the results and giving the final shape to this thesis.
Finally, thanks to my wife Pallavi, sons Vyom and Vihang who supported me
emotionally to complete this work.
I would like to dedicate this thesis to my late parents Shri Nathu Ram Sharma,
Shrimati Bhagwan Devi Sharma and grand mother Shrimati Rani Devi for
inculcating in me the spirit of higher learning, over and above any other worldly
gain.
Subhash Chandra Sharma
xviii
NOMENCLATURE AND SYMBOLS
ASTM = American Society for Testing and Materials
ASME = American Society of Mechanical Engineers
Al2O3 = Aluminium Oxide
ASI = American Standards Institute
ha = Contaminant average height/diameter (microns)
C = Radial clearance (m)
ΔC = Change in radial clearance (microns)
ha = Contaminant average size (microns)
CSWA = Cross Sectional Wear Area (mm2)
D = Bearing diameter (m)
e = Eccentricity (microns)
e o = Eccentricity at no load (microns)
Eox = Eccentricity at no load in the direction of Probe ‘X’
Eoy = Eccentricity at no load in the direction of Probe ‘Y’
ESDU = Engineering Science Data Unit
f0 = Wear depth at the first node (microns)
f1 = Sum of all the WD’s at odd nodes (i.e. 3, 5, 7,….17), (microns)
f2 = Sum of all the wd’s at even nodes (i.e. 2, 4, 6…16), (microns)
F = Frictional force (N)
FSF = Film Shape Factor
Ha = Hardness of the abrasive (Kgf/mm2)
Hb = Hardness of the bearing
Ho = Film thickness at no load (ho/RB)
h = Oil film thickness at any angular position (microns)
xix
hmin = Minimum oil film thickness in the bearing contact (microns)
Δh = Nodal distance along the outer crescent (microns)
HB = Brinell hardness
HRB = Rockwell hardness at ‘B’ scale
HRC = Rockwell hardness at ‘C’ scale
Hmin = Minimum oil film thickness (microns) (hmin/RB)
Δhmin = Change in minimum oil film thickness (microns)
Hxps = Measured minimum oil film thickness – start of test (microns)
Hxpm = Measured minimum oil film thickness – during test (microns)
Hxpe = Measured minimum oil film thickness – end of test (microns)
H esdu = Minimum oil film thickness from ESDU 84031 chart (microns)
H = Non dimensional film thickness (at any location) (H=h/RB)
hhdl = Film thickness value required for λ = 10 , (microns)
Hhdl = Film thickness required for γ = 10 , (microns)
IDmax = Maximum ID of bearing (mm)
ISO = International Standards Organisation
ID = Internal diameter
K ratio = Film thickness to particle size ratio (‘K’ ratio), K= hmin/ha)
l = Sliding distance (m)
L = Bearing length (m)
L = Bearing length in mm
L/D = Bearings length to diameter ratio
Lc = Cut off length (microns)
MWPA = Multi-wear parameter approach
N = Speed (rpm)
OD = Outer diameter of shaft sleeve (m)
xx
ΔOD = Change in Shaft sleeve OD (microns)
OR = Out-of-roundness (microns)
ΔORb = Change in out-of-roundness (microns)
ΔODs = Change in outer diameter of Shaft sleeve shaft (microns)
Orb = Out-of- roundness of bearing (microns)
Ors = Out-of –roundness of shaft sleeve (microns)
Ori = Localised or instantaneous out-of-roundness (microns)
Or1 and Or2 = Out-of-roundness of surface 1 and 2 respectively (microns)
Pi,j = Non dimensional pressure at i,j node (ph2/ηoL.Us)
Px = Distance measured with Probe X (microns)
Py = Distance measured with Probe Y (microns)
Pox = Distance measured with Probe ‘X’ at no load and full speed (microns)
Poy = Distance measured with Probe ‘Y’ at no load and full speed (microns)
Pc = Particle count (counts /ml)
Pcg = Gravimetric particle (g/l)
ΔPc = Change in particle count (counts/ml)
Rsk = Roughness skewness
RB = Bearing radius (m)
Rs = Shaft sleeve radius (m)
Rb = Bearing circumferential roughness (microns)
ΔRb = Change in bearing circumferential roughness (microns)
Rs = Change in shaft sleeve circumferential roughness (microns)
ΔRs = Change in Shaft sleeve roughness circumferential (microns)
Rbt = Bearing transverse roughness (microns)
xxi
ΔRbt = Change in bearing axial or transverse roughness (microns)
Rst = Shaft sleeve transverse roughness (microns)
ΔRst = Change in transverse roughness of shaft sleeve (microns)
Ra = Average roughness (microns)
Rp = Highest peak (microns)
Rt = Maximum upward excursion (microns)
Rq = Root mean square roughness (microns)
S = Sommerfeld number (S = (ημ/Wap).(Wap/C)2 )
SEM = Scanning Electron Microscope
SF = Scale Factor
t = Time (s)
Τ = Cross Sectional Wear Area (m2)
Us = Bearing velocity (m/s)
V = Wear volume (m3)
V = Computed wear volume (mm3)
W = Normal load (N)
Wz = Load component in Z direction
W = Non dimensional load wz/(ηoub). (C/RB)
W = Computed weight loss (mg)
Wapp. = Applied load (N)
Wcal = Non dimensional Load calculate (N)
WCE = Wear Characteristic Equations
wndn = Computed wear depth at nth node (microns)
WD = Computed maximum wear depth (microns)
WZW = Wear Zone Width (microns)
Wx = Load component in x direction (N)
xxii
Wy = Load component in y direction in thesis z direction (N)
Wb = Bearing weight (g)
ΔWb = Measured weight loss in bearing (mg)
Ws = Weight of shaft sleeve (g)
ΔWs = Measured weight loss in shaft sleeve shaft (mg)
WD = Computed maximum wear depth (microns)
WDmax = Measured maximum wear depth (microns)
ω = Cross sectional wear area (CSWA), (mm2)
X = Linear coordinate in the direction of circumference (mm)
X1 = Coordinates in space as distance from the tip of the X probe (m)
Y1 = Coordinates in space as distance from the tip of the Y probe (m)
y = War depth calculated at X distance using WCE (mm)
β = Angle subtended by SN and NB lines
ρ = bearing material density g/cm3
ε = Eccentricity ratio (εce
= )
φ = Angle MNS considered in probe geometry (degrees)
Φ = Attitude angle assumed (degrees)
Ψ = Calculated attitude angle (degrees)
η = Oil viscosity (Pa.s-1)
ηo = Inlet oil viscosity (Pa.s-1)
λ = Lambda ratio or Film parameter or specific film thickness (hmin/σcomp)
θ = Angle MNB in probe geometry (in this thesis)
σ1 and σ2 = RMS roughness values of surface 1 and 2 (microns)
γ ratio = Film Shape Factor (FSF)
σb and σs = RMS roughness of bearing and shaft sleeve (microns)
1
CHAPTER-1
1. INTRODUCTION
1.1 Rationale
Bearings are used for transmitting forces between machine components in relative
motion. One of the objectives of a machine designer is to incorporate efficient
bearings that minimise power consumption due to friction, and achieve longer life in
face of wear. A lubricant layer between the mating surfaces of a bearing often helps
in reducing frictional force, and thus, minimises wear. However, the machines
working in dusty environment are prone to higher wear rate; hence the effect of
contaminants in bearings lubrication is an important topic of study.
Machines working in dusty environment often fail prematurely and incur high
maintenance costs. To minimise such failures, operators change the lubricant
frequently, thus aggravating the hazardous waste disposal problem, which is a threat
to the ecological system. If machines can be designed to be more dust tolerant,
bearing life will be extended and thus reduce hazardous waste material (Scott and
Hargreaves, 1991). Thus, tribologists face simultaneous challenges of energy
conservation, environmental protection and machine reliability.
Bourne (2002) reported on the British army exercise in the Saif Sareea deserts of
Oman in year 2001. In this exercise, sixty-six Challenger-2 tanks were deployed at a
cost of A$230 million, but half of the tank engines seized due to dust. During the
Iraq war, Offley (2003) reported that Jessica Lynch was captured as her gun jammed
due to CLP lubricant failure in dusty environment. During the very first dust-storm
more than 200 infantry vehicles and 70 helicopters were disabled, and, as a result,
coalition forces faced heavy casualties. To conclude, failure of lubricants can lead to
loss of life as well; therefore, research into the performance of lubricants and wear
reducing additives in dusty environments is more important than ever before.
2
1.2 Background Research
Given the importance of lubricants in smooth machine operation, substantial research
has been conducted on the effect of contaminants in journal bearings. Such research
has emphasised the importance of using clean oils and given strategies for
controlling the contaminants. The most widely applied strategy is the application of
filters. Duchowski (1998) recommended that the filtration requirements for journal
bearing should be more stringent, and that ISO 4406 cleanliness level must be adhere
to using the ISO 16/14/12 cleanliness code; additionally, filtration requirements for
six micron contaminants (β6 ratio) should also be met. However, as the costs of
filtration increase exponentially with cleanliness, unduly high cleanliness is not
always economically viable.
Another strategy to minimise wear is to use antiwear additives, to mitigate the effect
of contaminants. Many such additives are commercially available; however, little
information is supplied by manufacturers about their efficacy. Manufacturers often
make unsubstantiated claims that users have no way to confirm.
However, in most machine components the wear is so small that it is difficult to
distinguish between the improvement due to additives, and the effect of operating
parameters such as speed, load and misalignment. Rowe (1980) has conducted
research to characterise additives in dust free environments. Furthermore, he used
elasto-hydrodynamic concentrated lubricated contacts. He suggested that further
research be conducted in this area. However, since the 1980s not much research has
been carried out on the efficacy of antiwear additives. Consequently this research is
filling a long standing knowledge gap.
1.3 Project Motivation
Professor W. Scott at Queensland University of Technology received requests from
the mining industry to investigate the effect of contaminants on oil change period.
Subsequently, Hirstch and Scott (1980) conducted an experimental study on journal
bearing lubricated with oil containing contaminants and found that bearing journals
with rough surfaces do not wear as rapidly as their smoother counterparts. A logical
explanation for this phenomenon – as given by Hirstch et. al. – is that the
3
contaminant particles take a preferential path through the lubricant film (i.e. valley-
to-valley of the rough surfaces) and hence reduces three-body abrasive wear. Martin
(1991) also supported Hirstch’s hypothesis, but could not provide any empirical
evidence. These researchers concluded that the surface topography (roughness /
waviness) of mating surfaces is important in almost every facet of machine
operation; nonetheless, there is no clear-cut evidence as to which topography works
the best. This is partly due to the difficulty in uniquely defining the micro-geometry
in quantitative terms. A "smooth is the best" attitude has developed which results in
the pursuit of expensive finishing processes, which are not only uneconomical but
may, in fact, impair the part performance.
Later, the mining industry also raised the issue of efficacy of the additives for
selecting appropriate additives for dusty applications. As a result this project was
developed as an Australian Postgraduate Award (Industry) (APAI) with the support
of Fuch Australia lubricant company.
After reviewing the literature and analysing the problem it was realised that rolling
element bearings used in the machines are either sealed or shielded, whereas sliding
bearings are often exposed, and hence need special attention. Since journal bearings
are the most widely used sub category of sliding bearings, these were chosen for this
study. This resulted in an experimental study on the effect of solid contaminants on
the wear of journal bearing and specifically on its micro-geometry, which includes
surface roughness, roundness, wear-depth and radial clearance.
Furthermore, there is a need to develop a systematic methodology for determining
the efficacy of the antiwear additives. Though, the tribological performance of an
additive can be judged by its ability to save energy and resist wear; it is easier to
measure wear than frictional losses. Therefore, this research project has placed
greater emphasis on wear measurements.
There are several antiwear additives available in the market; these additives are sold
separately or are blended with lubricants. The chemical composition of these
additives is mostly confidential. Five antiwear additives were supplied by Fuchs
Australia for this research, as these additives are used commercially. Some of these
additives are commercial products, while the others are experimental products with
4
proprietary composition. This experimental research is aims to using materials and
operating conditions as close to field conditions as possible.
1.4 Current Scenario
Moon (2007) has recently reported the detrimental effects of lubricants containing
solid contaminants. He concludes that solid contaminants reduce the life of
tribological components by 15%, cause 35% of downtime, and 82% of the wear.
Moon emphasises that particles smaller than the minimum oil film thickness in the
bearing contact do not cause much harm; however, the particles equal to the size of
the film thickness are highly detrimental.
Maru (2006) has studied the effect of antiwear additives on lubricants containing
solid contaminants on a tribometer in rotary and reciprocating motions. But his study
was confined to boundary lubrication regimes, and the focus was on the wear
mechanisms rather than characterisation of lubricants.
The available literature indicates that no significant research has been conducted on
oils containing solid contaminants treated with antiwear additives since 1980s, hence
this research is timely.
1.5 Knowledge Gaps
While much research has been carried on bearing wear, the literature review revealed
the following knowledge gaps:
• The effect of solid contaminants treated with antiwear additives on journal bearing wear has not been fully studied.
• Characterisation of antiwear additives based on their efficacy for dusty applications under hydrodynamic lubrication has not been carried out.
• The effect of solid contaminants on the bearing micro-geometry, and its effect on the bearing’s tribological performance is not well understood.
• There is no standard numerical parameter for classifying the performance of antiwear additives operating in dusty hydrodynamic lubrication conditions.
1.6 Aims and Objectives of the Research
The main aims of this research project are to conduct experiments on journal
5
bearings lubricated with oil containing solid contaminants treated with antiwear
additives, and study the following aspects:
a) The effect of contaminants –treated with antiwear additives– on journal bearing
wear
b) The effect of antiwear additives on the wear of journal bearings operating with
oils containing solid contaminants
c) The effect of change in micro-geometry on bearing’s tribological performance
d) Characterisation of additives using the most suitable wear measuring technique
The research objectives (deliverables) were derived from a literature review in
consultation with the project’s industrial partner (Fuchs Lubricants Australia). The
main objectives are as follows:
1. Compare the tribological performance of: a) journal bearings lubricated with
pure base oil, b) base oil containing solid contaminants, and c) oils containing
solid contaminants treated with different antiwear additives.
2. Determine the effect of solid contaminants on wear and micro-geometry of a
journal bearing.
3. Evaluate different wear measurement techniques for their suitability to
identifying a methodology for characterising the antiwear additives.
4. Study the effect of micro-geometry on the tribological performance by
measuring the change in minimum oil film thickness.
5. Characterise antiwear additives using a unique number, and rank them for
their efficacy.
1.7 Research Methodology
Before conducting the wear tests on the test bearings, micro-geometry parameters
and their effect on lubrication were examined. This required verification of micro-
geometry parameters including their metrology. Micro-geometry parameters i.e.
6
out-of-roundness, surface roughness and radial clearance were measured carefully
before conducting the tests. While measuring the radial clearance it was found that
ID of the bearings varies at different locations along the circumference. This is
mainly due to out-of-roundness, and hence, it is proposed that out-of-roundness
should also be specified along with the radial clearance, just like cut-off length is
specified along with the surface roughness. Further investigations revealed that the
out-of-roundness values are higher than the surface roughness values. This resulted
in proposing a new design parameter called Film Shape Factor (FSF) or gamma ratio.
This research has three main components that are:
1) The efficacy of anti-wear additives on bearing wear
2) Effect of micro-geometry on tribological performance
3) Characterisation of anti-wear additives
1.7.1 Effect of anti-wear additives on bearing wear
The research problem predicates the need for experimental studies that measures the
effect of antiwear additives on bearings lubricated with oil containing solid
contaminants. Therefore, a series of wear tests were planned on the pairs of 40mm
ID bronze bearings and steel shaft sleeve. The bearings were designed using
Engineering Science Data Unit method (ESDU 84031) and were tested under
simulated dusty conditions.
Wear tests generally suffer from poor repeatability. If factors influencing the wear
are not controlled then the wear results can vary by as much as a factor of 10 or more
(Bayer, 2004). The experiments were designed for best utilisation of the available
resources, which led to the following strategic decisions:
• The test should be conduct for short duration without repeating them.
• The wear results obtained from different methods need to be compared to
find out the level of accuracy of each method, and select the most reliable
method to obtain reliable results.
• Environment and procedures must be consistent, because the tests are not to
7
be repeated.
• Test must be conducted for K=1, sliding distance (l) = 7536m, and all other
operating and environmental parameters must be kept the same.
A performance parameter selection process based on weight loss, micro-
geometry and particle counts called multi wear parameter approach (MWPA)
was developed, this comprised fourteen parameters. Wherever appropriate
multiple observations were recorded, and average of these was used for analysis.
The following six measurement parameters were used in the MWPA:
1. Weight loss
2. Out-of-roundness
3. Roughness
4. Radial clearance
5. Wear depth
6. Particle count
1.7.2 Effect of micro-geometry on tribological performance
The effect of change in micro-geometry on the tribological performance was also
measured, by recording the change in minimum oil film thickness. However, the
tribological performance of test bearings was affected not only due to change in
micro-geometry but also as a result of: a) combined effect of change in bearing
geometry, b) influence of antiwear additives, and c) the oil flow restrictions due to
concentration of the solid contaminants in and around the bearing contact.
The change in minimum oil film thickness was recorded at three intervals i.e. at the
beginning, middle and at the end of each test. Each measurement technique was
examined critically, merits and demerits of these techniques were compared, and the
most suitable measurement technique was chosen for characterising the antiwear
additives.
Bearings were designed using ESDU 84031 method and seven test bearing sets were
fabricated for this experimental study. First test used pure base oil, and the second
test used oil containing 4% (by weight of 16 micron) Al2O3 powder. Subsequent, five
8
tests used contaminated oil treated with antiwear additives.
To make sure that the measurement system used in this study works satisfactorily,
the results obtained with pure base oils were compared with the values predicted by
the on-line ESDU A 9305 software program, as well as by using a FORTRAN
program based on the algorithm proposed by Pai and Mazumdar (1992).
1.7.3 Characterisation of anti-wear additives
Wear test results were analysed for accuracy of the measurement technique as well as
for the efficacy of the antiiwear additives. The out-of-roundness technique was found
to be the most suitable. Using this technique, an antiwear additive characterisation
method was developed. In this method the out-of-roundness traces were used for
computing the weight loss, and by using this weight loss a number called wear
characteristic number (N) was derived. This number represents the efficacy of an
additive and hence users can select an additive for their applications using this
number.
1.8 Contribution to the Body of Knowledge
In this experimental research, various wear measurement techniques were evaluated
to examine their suitability for studying wear in bearings lubricated with oil
containing solid contaminants, and treated with antiwear additives. Effect of solid
contaminants with and without additives was studied on the wear of journal bearing
components; and specifically on the micro-geometry of the bearings. The effect of
change in micro-geometry on minimum oil film thickness was also studied. Different
wear measurement techniques were compared and the best technique was chosen for
characterising the antiwear additives – specifically for dusty applications.
Out-of-roundness was found to be the most reliable and suitable micro-geometry
parameter for characterising antiwear additives. This parameter was used to develop
a method for characterising the antiwear additives. A wear characteristic number (N),
was derived to rank the additives based on their efficacy.
9
The main contributions of this research to the existing body of knowledge include:
• Developed a clear understanding of the effect of antiwear additives on the
wear of a journal bearing lubricated with oils containing solid contaminants;
and in particular, on the change in bearing micro-geometry.
• Established a process for comparative analysis of different wear measuring
methods –for characterising antiwear additives.
• Developed a novel geometrical method for improving the precision of oil film
thickness measurement –using floating proximity probes.
• Demonstrated the inadequacy of the current metrological practices for
measuring the radial clearance in journal bearings; and as a result predicated
the need to review the current bearing design heuristic –that the lambda ratio
should be close to 10.
• A Film Shape Factor (FSF or gamma ratio) has been proposed as a design
parameter for ensuring adequate separation of bearing surfaces of a journal
bearing working in hydrodynamic lubrication regime where out-of-roundness
value is higher than the surface roughness.
• Derived wear characteristic equations (WCE) to obtain a unique wear
signature for individual antiwear additives.
• Derived a wear characteristic number (N) to characterise antiwear additives
and facilitate the selection of the most suitable antiwear additive for a given
dusty environment.
These contributions to the existing body of knowledge for the tribology of machine
elements will help in extending the life of journal bearings, minimise downtime, save
energy, benefit the environment –by reducing hazardous waste of contaminated oils–
and help lubricant users to select the most suitable antiwear additive for dusty
applications.
10
1.9 Organisation of the Thesis
This thesis comprises seven chapters and six appendices. A brief introduction to each
chapter is given in the following:
Chapter 1 gives an introduction to the thesis, including problem statement, objectives
and contribution to the body of knowledge.
Chapter 2 provides a detailed literature review concerning the various aspects of this
research and highlights knowledge gaps.
Chapter 3 details the experimental design and tests set up. It explains the issues
related to metrology of radial clearance, limitations of the current concept of specific
film thickness and a proposed gamma ratio for hydrodynamic bearing design. The
chapter also gives the details of a proposed geometrical method for measuring oil
film thickness with proximity probes mounted on a floating bearing housing.
Chapter 4 presents the results and their detailed analysis, which is used to identify the
most suitable wear measurement technique for characterising antiwear additives. It
also provides a discussion of the influence of each anti-wear additive on the wear of
test bearings.
Chapter 5 gives the methodology for characterising the additives, and the wear
characteristic number for selecting an additive based on its antiwear performance.
Chapter 6 presents the main conclusions of this research project.
Chapter 7 suggests directions for future research.
Appendices A to F are summarised and included at the end of the main thesis body.
11
CHAPTER 2
2. LITERATURE REVIEW
2.1 Overview
This research project required an in-depth knowledge of various components of the
problem, which included contaminants, wear measurements, antiwear additives,
miro-geometry and characterisation of antiwear additives. Thus the literature review
was guided by the following three main areas connected to each other:
a) The effect of contaminants –treated with antiwear additives– on journal bearing
wear
b) The effect of change in micro-geometry on bearing’s tribological performance
c) Characterisation of additives using the most suitable wear measuring technique
The literature review was further classified, and a search was conducted on the
following topics:
• Contaminants’ effect on wear
• Effect of solid contaminants on journal bearings
• Tribological performance and criteria for its measurement
• Micro-geometry parameters and their effect on bearing lubrication
• Antiwear additives and their characterisation
The literature on antiwear additive gives more emphasis on additive chemistry, only
limited information is available on additive characterisation. In this context, literature
on applications of commonly used antiwear additives was reviewed.
Appendices A to F are summarised and included at the end of the main thesis body.
12
2.2 Contaminant Effects on Wear
Bearings are usually prone to solids contamination at the lubricated contacts and they
tend to damage the bearing. In fact even the cleanest of lubricants carry some
contaminants as it is impossible to remove all particulate matter, which passes
through the filter Khonsari et al. (1999). The need for cleanliness of lubricants was
emphasised by Douglas (1989) who suggested that cleanliness standards need to be
reviewed from time to time in view of changing requirements of machinery. Godfrey
(1989) highlighted the relationship between lubricant cleanliness and bearing life.
Dufrane et.al. (1983) reported on a turbine bearing survey where 54% failures were
attributed to contamination of lubricants.
Duchowski (1998) defined a contaminant as an unwanted foreign element or
substance that can have adverse effects on the lubrication system of an operating
machine. These adverse effects are; efficiency, service life and reliability of the
machine components. Contamination can take place in either form of matter such as
gas, liquid or solid. Gases normally dissolve in the liquid lubricant or form bubbles;
liquids are either directly dissolved or form an emulsion but solids are always found
in the form of suspended particles. In the present study only contamination of the
lubricant by solid particles is considered.
Dwyer-Joyce et al. (1990) has indicated that lubricant contamination may be by the
three forms of matter: solid, liquid and gas either individually or in combination.
There are three main sources of solids contamination in machinery:
1. Implanted contaminants: The contaminant particles are introduced at the time of
manufacturing including the process itself, assembly, handling, packaging, and
transportation.
2. Ingested Contaminants: Particles may enter the system from environment
through seals or air filters etc.
3. Self-Generated Contaminants: These are mainly due to wear of mating surfaces
in the lubricated contacts. There may be different forms of wear mechanisms
responsible for the generation of these contaminants. Though rare, it also
includes erosive wear mechanism, which is caused by high flow rate of lubricant
13
through pipes and seals.
Knowing the concentration, size distribution, and material properties such as
geometry, hardness, etc., the severity of contamination can be assessed in a
lubrication system. It is easy to determine the first two parameters in comparison to
hardness and material type of an individual particle. The hardness of particles can be
assumed to be the hardness of the bulk material, which is reported in any standard
handbook. The material type can be determined with the help of energy dispersive X-
ray micro analysis and Infra Red (IR) emission (Kuhnell (1992)) or any other
technique used for elemental analysis. These methods are complex in nature but well
accepted by modern researchers.
As shown in Table 2.1, Beghini et al. (1992), Dwyer-Joyce et al. (1990), Dwyer-
Joyce (1992) and Hamer et al. (1989) have reported the following solid contaminant
concentration results after analysis of used lubricant collected from different type of
machines.
Table 2.1: Sources of Solid Contaminants, Dwyer-Joyce (1993)
Investigator Sample Source
Concentration g/l
Size range μm
Materials present
Dwyer-Joyce(1983)
Paper mill 0.4 0-150 Cu, Sn, silicates
Dwyer-Joyce (1983)
Motor vehicle sump
1-2 0-250 Al, Cu, Fe, Fe, Pb
SKF ERC (Beghini 1992)
Various bearing lubricants
0.3-1.5 0-250 Fe, Al, Cu, Sn, SiC, sand
Dwyer-Joyce (1993)
Aircraft gas turbines
1-2 0-200 Fe, C, silicates
Fuchs Australia *
Mining industry
2-5g/l 0-200 Depending upon the mining industry
* Data from Fuchs Australia (personal communication)
Effect of contaminants on bearings can be correlated with abrasive wear. It is
important to understand abrasive wear in general and its effect on different
14
tribological elements.
A bearing may wear out faster due to presence of solid contaminants present in the
oil. The gradual wear affects the bearing performance in terms of load carrying
capacity and subsequently accelerates the increase in friction and wear. These effects
have been very recently studied by Fillon (2002).
2.2.1 A general view of abrasive wear
Friction and wear are inevitable in tribological components of machines. The
Institution of Mechanical Engineers, U.K., has defined wear as;
“The progressive loss of substance from the surface of a body brought about by
mechanical action. Usually it reduces the serviceability of a body, but it can be
beneficial in its early stages.” Dwyer- Joyce (1993).
There are different types of wear mechanisms (Halling (1975), Rabinowicz (1965)).
Archard (1953) introduced the first mathematical equation to quantify out wear of
metals as follows:
V= kW l/3H (2.1)
Where ‘V’ is the wear volume, ‘K’ is known as wear coefficient, ‘W’ is normal load,
‘l ’ is sliding distance and ‘H’ is hardness of the substrate. The model was developed
for adhesive wear but Rabinowicz (1965) modified it for abrasive wear and modified
the above equation to;
HWlV
πθcot
= (2.2)
Where θ is the angle of the cone used as an indenter and all other terms are same as
in Equation 2.1. This equation was derived for an abrasive wear process, where an
indenter of a cone shape of radius r and cone angle θ was used. In this model k/3 is
replaced by (cot (θ)/π).
Rabinowicz et al. (1961) carried out an experimental study on three-body wear with
different sizes of grits and found that smaller grits lead to low wear rates. Later
15
Rabinowicz (1965) investigated a threshold limit of grit size above which wear did
not increase when silicon carbide was used as an abrasive mixed with oil and water
in a sliding bearing.
Misra and Finnie (1980), Misra and Finnie (1981), Misra and Finnie (1982) Misra
and Finnie (1983) published a series of research papers based on several three-body
abrasive wear experiments and concluded that a critical particle size affects the
abrasive wear severely. Misra and Finnie (1980) divided abrasive wear into three
groups; two-body, three- body and two and a half body wear. They further
subdivided three-body wear as shown in Figure 2.1
Figure 2.1 Classification of abrasive wear, Misra & Finnie (1980)
They defined open three-body wear system as wear occurring in two opposing
surfaces which are apart whereas closed three-body wear occurs where wear particles
are trapped between the surfaces, such as bearings lubricated with an oil containing
solid contaminants. Similarly stressed situations are those where particles are crushed
into small pieces as in the case of ball mills. In the case of low stress, wear particles
16
do not break into small particles.
Larsen-Basse (1975) found that moisture is responsible for an increase in wear in
presence of silicon carbide and aluminium oxide abrasives in a sliding contact up to
65%.Chandrasekaran et al. (1985) studied the effect of abrasive particles on wear in a
four ball test machine under boundary lubrication regimes and determined the size of
grits on scuffing in the contact.
Fodor (1980); Fodor (1987); Fodor and Ling (1987) studied the effects of abrasive
particle size on internal combustion engine components and demonstrated that by
using correct filtration system, up to 4% energy can be saved and the oil change
period can be extended up to 10000 km. Odi-Owei and Roylance (1987) performed
some experiments with a four-ball test machine under sliding conditions. They
observed that the contaminants were embedded in the inlet zone on the ball surfaces
and restricted the oil flow, which results in ploughing action. Concentration and
particle size both were damaging in the boundary lubrication regime as adhesive
wear dominated the abrasive wear and leads to scuffing. They presented a simple
mathematical model to predict friction under abrasive containing lubricant
conditions.
In an experimental study on a pin-on-disk machine, Mehan (1988) compared the
wear behaviour of different materials and coatings under contaminated lubricant
conditions. The study revealed that 2g/ l concentration of Al2O3 particles in white
mineral oil caused less wear. Furthermore, cemented tungsten carbide (WC) cermets
as well as WC coating displayed less than that for chromium-plated pins when
rubbed against cast iron in an uncontaminated lubricant. The study also revealed that
the increase in surface roughness from 0.15 micron to 0.8 micron caused 10 times
more wear.
The geometry of abrasive particles influences the wear pattern. Defining the
geometry of their irregular shape has been a challenge for researchers. A number of
researchers have suggested numerical methods and imaging techniques to determine
the shape of the particles mathematically Kuhnell (1992).
17
2.2.2 Abrasive wear in sliding bearings
Solid contaminants are the enemy of lubricated surfaces of both sliding and rolling
bearings. They not only increase friction between the two moving surfaces but also
cause accelerated wear. The lubrication mechanism and especially the ratio between
lubricant film thickness and contaminant size is important in determining the wear
severity in bearings (Douglas, 1989).
Journal bearings normally operate in hydrodynamic lubrication regime where fluid
properties influence the fluid flow through the contact. Particles mixed with
lubricants in any form influence the flow and hence the oil film thickness, which is
directly related to their performance (Das, et. al., 2004).
2.2.3 Micro polar lubricant effects in bearing lubrication
The effect of particles mixed with lubricant has been a major concern in bearing
performance for practicing engineers since they understood hydrodynamic
lubrication theory Allen and Kline (1971). These particles may be long chain
polymers due to additives, dust particles, or other solid contaminants. The concept of
micro-polar effects has also been applied to bearings lubricated with oils containing
excessive solid contaminants in colloidal form and it was found that under dynamic
loading their stability improved Das et al. (2004).
Viscosity is a major criterion for satisfactory bearing performance. This topic has
been discussed by several bearing designers such as; Eringer (1964), Allen and
Kiline (1971) and Khonsari and Brewe (1989). These studies have who contributed
significantly to lubrication theory where large quantities of polymeric additives are
used.
A common assumption in bearing design is that lubricants are Newtonian. However
this is not the case when polymeric substances are mixed with the lubricant. Many
researchers, Prakash and Sinha (1975),Prakash and Sinha (1977), Mahanti (1976) ,
Zaheeruddin and Isa (1978), Tipei(1979), Sinha et al.(1981), Prakash and Kumar
(1987) have done extensive work in the field of journal bearings. They recognized
that the effect of synthetic lubricants can be significant depending upon the
concentration of solid particles in the lubricants. They found that polymers in the
18
lubricants affect the friction, viscosity and load carrying capacity. Other researchers
Khonsari and Kim (1989), Dai and Khonsari (1992) have presented theoretical
models to incorporate micropolar effects in journal bearings under different
operating conditions. 3-dimensional roughness effects were considered by Tsann-
Rong (1996).
The theory of micropolar lubrication is governed by continuum mechanics. The
continuous media is considered as a set of structured particles possessing individual
mass and velocity. The definition by Eringer (1964) a simple micro-fluid is a
viscous medium whose behaviour and properties are affected by the local motion of
the particles in its micro-volume. This theory may be applied to contaminant effects
in journal bearings. But Ronen and Malkin (1981) considered that the concentration
of solid contaminants is usually too low to be worth considering. This resulted from
an experimental study of performance evaluation of a journal bearing lubricated with
solid contaminated fluids.
In recent years experimental results show that the Newtonian viscous lubricant
blended with small amount of long chained additives improve lubricant properties.
Furthermore Allen and Kilne (1971) demonstrated that the micro-continuum model
can be used for first approximation for modelling bearings lubricated with the oil
containing dirt and metal particles. Das (2002) studied the effect of contaminants of
different length micro-polar fluids on journal bearing performance.
2.3 Effect of Solid Contaminants on Journal Bearing Performance
Journal bearing contamination is very common because these bearings are normally
not sealed like rolling element bearings. The mechanism of contaminant ingestion
into journal bearing contacts is different from rolling element bearings mainly due to
difference in the order of minimum oil film thickness in their contacts. The order of
minimum film thickness is 10-100 times higher in journal bearings in comparison to
rolling element bearings.
Sliding and rolling element bearings operate under thick film lubrication called
hydrodynamic (HDL) and Elasto-hydrodynamic lubrication (EHL) regimes
19
respectively and each exhibit low friction and wear. Though journal bearings are
much simpler in construction than rolling element bearings, the theory of their
lubrication mechanism is equally complex. An enormous amount of research on
bearing design has contributed to the development of high performance bearings.
McKee (1927) conducted a study on a journal bearing set-up to determine the effect
of contaminants on bearing performance. Friction was measured as a function of
bearing characteristic number (ZN/P), where Z is viscosity, N is speed and P is
specific load. He observed that solid contaminants are responsible for increased
coefficient of friction and bearings made of softer materials wear faster than the hard
material shafts.
Roach and Mich (1950) reported that in year 1945 Rosenfield investigated the effect
of particle size on bearing wear and found that the particles of the size of minimum
film thickness in the bearing are the most dangerous for elevated wear in bearings.
Roach and Mich (1950) conducted a study to examine the effect of solid
contaminants mixed with oil on temperature rise in journal bearings which is a
crucial parameter for monitoring the satisfactory performance of a bearing. The rise
in temperature was observed for different material combinations. They also
investigated the influence of solid contaminants on other bearing performance
parameters such as wear and embedability of the material pair. They identified three
basic situations with regard to the ratio between particle size and minimum film
thickness: the size of the abrasive particles is larger, smaller or equal to the height of
the film thickness in the contact. These can be mathematically represented by ‘K’
ratios;
K = 1, K > 1, and K < 1, where K = ha/hmin, where ha = height of the abrasive
particle and hmin = minimum lubricant film thickness.
The following conclusions were drawn from his study:
1. Particles of size smaller than the minimum film thickness in the contact are not
harmful to journal bearings.
2. Operating bearings at a high flow rate and a small l /d ratio where l is the length
20
of the bearing and d is the diameter of the journal is a better way to avoid
embeddability problem.
3. Presence of abrasives in a lubricated contact is responsible for high friction
resulting in more resistance to achieve higher speeds without temperature rise.
Running such bearings with a larger film thickness is advisable to achieve longer
bearing life and low frictional losses.
In order to be more realistic, Rylander (1952) considered different types of solid
contaminants and examined their effect on the size of the contaminant in comparison
to minimum oil film thickness in the bearing and found that:
• The concentration of graphite particles reduced the lubricant flow rate in the
contact,
• Concentration of contaminants and coefficient of friction are linked,
• A hard shaft wears faster than the soft bearing because wear particles embedded
in the soft bearing surface act as a cutting tool, and
• Particles smaller than the minimum film thickness do not harm the bearing
performance as they pass through the minimum clearance without creating any
obstruction in flow of the lubricant through the contact.
Broeder and Hejnekamp (1965) measured the wear distribution (around the
circumference) in a hydrodynamic bearing with stationary radial load. Analysis of
Broeder and Hejnekamp results showed that significant liner wear occurs at the
circumferential locations where the local film thickness is less than the abrasive
particle size and that the wear rate is more intensive at locations of smaller oil film
thickness. They also observed different types of wear mechanisms such as pitting,
ploughing, and cutting in the test bearings.
Elwell (1977) used steel-weld spatters and fly ash as abrasive contaminants and
observed that the smaller particles liberated from the shaft as a result of large
particles striking the shaft with high force were more damaging than the large
particles themselves. Fly ash does not harm the surfaces but it creates a lapping
21
action and makes surfaces smooth. However, this happened at only low speeds. At
high speed, the wear was high and surfaces were rougher.
Several lubricated wear models were developed by Rowe (1967); Rowe (1970);
Rowe (1986) for different types of additives in different wear modes. However the
performance of individual additives was presented mathematically without
considering the effects of abrasive particles. Ronen et al. (1980) prompted by the
high frequency of engine bearing failures on vehicles operating in the deserts of
Israel, conducted an experimental study on journal bearings and confirmed the
findings of Broeder and Heijnekamp (1965) in their preliminary studies.
Hirstch et al. (1980) considered roughness effects in their experimental studies on
bearings lubricated with oils containing solid contaminants. They used rough and
smooth bearings with particles of different sizes and confirmed that large abrasive
particles caused a greater amount of wear than the smaller particles. They also
observed a unique phenomenon of journal wear; in the presence of abrasive particles,
initially there was low wear in rougher journals as compared to smoother ones. The
rough journals became smoother during the wear process when the abrasive particle
size was smaller than the minimum film thickness. However, the bearing surface
roughness increased in the vicinity of minimum film thickness in presence of
contaminants in the lubricant. Their finding of low initial wear with the rough
journals is an important finding. They explained this phenomenon by a preferential
path theory. According to preferential path theory, particles find an easy way to flow
through valleys of rough surfaces in the contact zone, hence three-body abrasive
wear does not occur and that is why the wear rate is lower in the shaft at the initial
stage. Frith and Scott (1993) and Martin (1991) also supported this phenomenon but
they could not provide reliable experimental evidence.
Ronen and Malkin (1981); Ronen and Malkin (1983); Ronen et al. (1980) continued
their experimental work using contaminants with automotive filter testing such as
engine crank case sludge and Arizona dusts. Initially they confirmed that wear of
journal and liner under dynamic loading conditions depends upon the minimum film
thickness where large particles are primarily responsible for the wear. They also
emphasised that so called ‘clean oil’ contained wear particles which may be system
generated or ingested from other sources not taken into account initially. In
22
another part of their experimental study they considered hardness of journal and
liners and studied six different combinations of journals and bearings and observed
that smaller shaft to liner hardness ratio causes relatively more liner wear and less
shaft wear. In their first mathematical model Ronen and Malkin (1981), they
explained that abrasive particles are either locked due to embedding in the soft
bearing material or they may roll. The shaft-to-liner hardness ratio of 3-4 is more
dangerous than higher hardness ratios. By controlling the minimum film thickness,
partial embedding of the particles can be avoided. They also concluded that
irrespective of size and hardness, contaminants increase the friction in bearings.
Their experimental work is quite comprehensive in nature and gives a solid base for
further research in this area.
In another experimental study on journal bearings, Dufrane (1983) and Dufrane and
Kannel (1989) confirmed that hydrodynamic bearings lubricated with clean oil
should be operated at a lambda ratio greater than 10, where lambda ratio is the ratio
of the minimum oil film thickness and composite roughness of the surfaces in
contact. They also warned that this ratio is not enough for bearings lubricated with
solid contaminants.
Contaminant effects on journal bearings were also examined by other researchers
Dufrane et al (1983), Watanabe et al. (1985), Dufrane and Kannel (1989). They used
five different grades of SiO2 abrasive particles with different hardness ratios (bearing
to abrasive). Their observations confirmed the findings of other researchers that wear
tends to be serious when particle size approaches minimum film thickness.
Prakash and Kumar (1987) studied the effect of contaminant size on unfilled PTFE
bearings, phosphor bronze and Al-Si alloy bearings under boundary lubrication. The
experimental study revealed that smaller particles (7-micron cut-off) caused
maximum wear in phosphor bronze whereas large size (25-micron cut-off) particles
were harmful for PTFE bearings. In general 15-micron particles caused maximum
wear in metallic bearings in comparison to 7 and 25-micron particles. Their results
are valid only in the boundary lubrication regime and therefore cannot be applied to
hydrodynamic bearings.
Fodor (1987), Fodor and Ling (1987) used Air Cleaner Fine Test Dust (ACFTD) and
23
silica particles and studied the effect of their size and concentration on friction and
wear of hydrodynamic bearings. They found that friction and wear increase with
increase in size and concentration of the particles. They designed new filters based
on their previous studies on I.C. engines in a contaminated environment and
demonstrated that the oil change periods in these engines can be extended
significantly.
Xuan (1989) conducted experiments on a journal bearing test rig and found the effect
of different hardness ratios between journal, bearing and abrasives on wear of
individual bearing parts. They defined the limits of hardness ratios to achieve low
wear in the bearings.
Effect of abrasives and flash temperature considerations in contact zone are also
important from a scuffing point of view. Khonsari and Wang (1990) studied this
problem in depth and developed a model by which critical temperature and critical
abrasive size can be calculated for a given bearing by a governing equation. Their
model is very basic in nature and useful for further research on scuffing in bearings.
The following conclusions drawn from their studies are useful for research into the
abrasive wear of bearings.
• The size of particles, the hardness of the overlay and penetration depth of the
solid particle into the overlay influences the scuffing of the bearing,
• The flash temperature rises with rise in hardness of the overlay, and
• At higher speeds the critical size of the particle beyond which the chance of
scuffing is remote approaches the order of the minimum film thickness.
In continuation of their research, Khonsari et al. (1999) demonstrated that a particle
partially embedded in the bearing overlay can cause scuffing if the flash temperature
exceeds a critical value. They emphasized that the other factors which may influence
the scuffing are operating speed, hardness ratio, diameter of the particle, thermal
diffusibility of the slider and orientation of the particle when they are embedded
within the surfaces. The study led to a theoretical model in which the filter rating
required to minimise the failure rate can be determined. However this model is
24
rudimentary and based on many assumptions and therefore requires experimental
confirmation.
The effect of speed on friction and wear was studied by Wlkstrom et al. (1993). They
used bronze and babbitt journal bearings with lubricants containing quartz and iron
particles and found that friction did not increase as significantly as the wear. At low
speeds the shaft surfaces were smoothed due to wear. They also confirmed that a
hard particle embedded in the overlay acted as a cutting tool for the journal.
Researchers Narayanan et. al. (1995) have shown that solid contaminants are not
only the foreign particles or wear debris but are also found in the form of additives.
Most lubricants contain different additives which are introduced to enhance certain
characteristics of the lubricant. Added to these, there exist undesirable contaminants
in the lubricant which along with additives form a dilute suspension in the oil. An
important effect of the solid particles suspended in the oil is to produce a thickening
effect. Also, it is found that there is an increase in the viscosity in the part of the
lubricant film which is in the vicinity of the journal and bearing surfaces due to
adhesion and other surface phenomena which have not been fully explained. This
results in a lubricant film with variable viscosity. Also, non-uniform distribution of
the solid suspended particles builds up concentration gradients which result in mass
transfer of these particles predominantly across the film thickness from the journal
surface, which results in a reduction of the thickening effect. In their study they
found that the results reveal enhanced load-carrying capacity and improved stability
characteristics for the micro polar lubricant as against the single-phase Newtonian
lubricant.
Among recent studies, Duchowski's (1998) experimental findings are very important
from the point of view of filtration requirements in journal bearings. He
demonstrated experimentally that the damaging particles are of the size approaching
minimum film thickness or bigger. He stressed that the existing filtration practices
need to be reviewed. In the light of his experimental findings he recommended that
the users of journal bearings (5.1 to 20.3 cm in diameter) should maintain cleanliness
level on the ISO 4406 scale below 16/14/12 and filter elements rated at 6 micron
(β6> 200) or finer must be used to minimise failures in journal bearings.
25
Din and Kassfeldt (1999) performed experiments on a journal bearing in mixed
lubrication regime where self-generated solid contaminants were mixed with
environmentally friendly lubricants. The results showed that even though friction
was high, wear rate was low. The results obtained were influenced more by the
lubricant rather than the contaminant itself.
In a recent experimental study Mizuhara et al. (2000) showed that speed plays a vital
role in the wear of bearings when solid contaminants are present in the lubricant.
The increased friction due to contaminants when operating at low speeds disappears
at high speeds. The rate of change in friction reduces at higher concentrations.
Moreover, the friction depends upon the number of particles entering the contact.
Velocity dependence is mainly governed by the interfering time (the duration of the
particles staying in the contact) and load supported by the particles. A theoretical
model has been developed based on these findings.
2.3.1 Effects of abrasive hardness on wear
Hardness of the mating surfaces and abrasive particles influences the wear of bearing
elements severely and the wear of an individual component depends upon the
hardness ratios of the three bodies in contact. Researchers have examined the effects
of hardness of different constituents in the bearing contact and recommended some
hardness ratios for journals and bearings operating with lubricants containing
abrasives. The experimental studies conducted by Roach and Mich (1950), Rylander
(1952) were of a very basic in nature. Subsequently Broeder (1965) investigated the
tribological wear of HV 700 hardness mating with various bearing sleeves. The
mating pair was lubricated with oil contaminated by silicon carbide grain of an
average diameter of 21 micron. The wear of the shaft that mated with same hardness
HV 700 was three times less than the shaft that was mated with bronze sleeve with
HV80 hardness.
An important study in this area was carried out by Czichos (1978) who suggested a
systems approach to solve tribological problems and proposed three wear regimes
based on the hardness ratios between abrasive hardness Ha and hardness of the metal
surface subjected to wear Hm. These regimes are shown in Figure 2.2. Truscott
(1972) confirmed these models by experimental studies. The experiments were based
26
on hydraulic systems and different critical ratios of particle to surface hardness ratios
for the three regimes of wear were determined.
Rigney (1994) demonstrated experimentally that in three-body abrasive wear, cutting
and deformation co-exist. He demonstrated that when hardness of the materials
changes, the ratio of cutting to plastic deformation also changes. The theoretical
model developed by him satisfies the experimental findings.
Figure.2.2 Effect of abrasive hardness on wear rate, Czichos (1978)
Xuan et. al. (1989) performed an exhaustive experimental study on a journal bearing
test rig and established relationships between the journal hardness-Hj, bearing
hardness-Hb and abrasive hardness-Ha. They conducted wear test for Hb/Hj ratios
0.75, 0.6 and 0.3. The bearing and shaft pairs were subjected to four types of
abrasive particles, with Hj/Ha ratio ranging from 0.14 to 2.75. From these
experiments empirical constants of the wear function were obtained. The critical
hardness ratio and the wear coefficients were also analysed. They concluded from
their study that the wear rate in journal bearings is maximum, when Hb/Hj ratio is
higher than one and Hj/Ha ratio is lower than one. The wear rate is the lowest when
Hb/Hj is low and Hj/Ha is high.
27
2.3.2 Contaminant motion in lubricated contact
It is well known that under dry condition sliding motion causes more friction than
rolling motion. In a three body wear situation where solid particles are suspended in
oil, it is difficult to monitor the motion of these particles. A theoretical model was
presented by Ronen and Malkin (1983) for three body wear. This model was
rudimentary in nature; nonetheless it gave a new direction to other researchers to
explain the motion of abrasives in a lubricated contact. Berthier and Godet (1989)
have defined velocity and friction zones in sliding wear, and Fang et. al. (1991, 1992,
1993) gave a seminal theory entailing the motion of particles in three-body wear.
William and Hyncica (1992) extended this work by performing abrasive wear
experiments using abrasives in a foil bearing. They observed the motion of abrasive
particles when the size of particles was small in comparison to lubricant film
thickness. They also found that these particles tumble and leave erosive wear marks
on the surface and at higher ratios of abrasive size to film thickness, grooving of
surfaces occurred. The critical particle size to film thickness ratio in sliding contacts
is given by:
βsec/ =hD (2.3)
Where; D, h and β are defined in the Figure 2.3.
As shown in Figure 2.3, a particle of height D with grit angle β (always less than 90
and greater than 45 degree angle) and bigger than the film thickness gap h is
entrapped. The transition typically occurred for a ratio of about two. Using the force
equilibrium and geometry of a single abrasive grit they modelled the tumbling and
ploughing processes caused by the abrasives and showed the conditions of a particle
to be trapped in the soft body cutting the harder body. The model also demonstrated
how a harder material might wear more than its softer counter face.
28
Figure.2.3 Particle Motion in bearing contact (William and Hyncica (1992)
Some researchers Fang et al. (1993); Fang et al. (1992); Fang et al. (1991); Kragelski
et al. (1992); Kragelski (1982) investigated the particle motion in a lubricated contact
and a new relationship for the motion of a particle in the contact zone was proposed
Fang et al. (1993):
V = αV1 ± βV2 where α + β= 1 ( 2.4)
21
2
HHH+
=α and 21
1
HHH+
=β (2.5)
Where V1 and V2 are the velocities of surface one and two respectively, and V is the
linear velocity of the particle with respect to the stationery surface. Factors α and β
are derived from the hardness ratios H1 and H2 of the two surfaces in relative motion
as shown in Equation 2.5. The sum of factors α and β is equal to one, thus their
individual value is always less than one. The relationship indicates that the velocity
of the particle is mainly governed by the harder material. This model was developed
considering a spherical abrasive particle moving at velocity V relative to the
stationery lubricated surface. Volumes of the deformed components of the abrasive
particles as well as the surfaces were calculated based on the geometry of the
particles.
Fang et al. (1991), Fang et al. (1992), Fang et al. (1993) developed a new apparatus
and conducted an extensive study on the motion of abrasive particles. In three body
wear they found a relationship in terms of coefficient of friction and particle
29
dimensions, which helps to determine the type of motion occurring (sliding or
rolling). The sliding motion leaves scratches on the surfaces whilst rolling motion
creates pits. Wilkstrom et al. (1993) conducted some experiments at low speed and
low concentration (0.02%) of solid contaminants and found that the friction rise in
journal bearings was very low, and the surfaces were smooth due to lapping effect.
They emphasised the effect of the critical contaminant size to minimum oil film
thickness ratios (K ratios) on the wear of the bearing liner and established that K
ratios equal to one causes severe wear.
2.4 Anti-wear Additives and Performance Characterisation
The use of additives in lubricants began as early as 19th century when fatty oil and
sulfur were added to mineral oils to acquire higher load carrying capacities from
lubricants Booser (1983). As such lubricants perform several functions other than
keeping two mating surfaces of a machine component apart. The most important of
all is to reduce wear and friction to meet the tribological goals, which means
conservation of energy and material. The other functions are needs based and vary
from one application to the other such as corrosion, anti-oxidant, extreme pressure
(EP), viscosity index improvers (VI), pour point depressant, anti-foams,
detergent/dispersant etc. The world’s total lubricant production of more than 40
million metric ton Khorramian et al. (1993) all contain one additive or another.
Additives are chemicals added to a carrier lubricant to harness any one or more
properties mentioned above.
In hydrodynamic bearings, film thickness to roughness ratio (lambda ratio) is more
than four to ensure that there is no asperity interaction in the lubricated contact.
Hence ideally there is no wear in this regime of lubrication. However, this cannot be
assured due to two main reasons: firstly, machines start and stop, thus speeds vary
from zero to maximum during start and vice versa when stopping. Secondly,
unintentional changes in operating conditions do occur. In both cases it is the
minimum oil film thickness which is affected. These operating conditions, which can
affect the film thickness, are load, temperature, misalignment, contaminants, undue
vibrations, etc. Under these situations, bearings are lubricated with either no
lubricant film or only a few molecular layers thick as in boundary lubrication regime,
30
Dorinson and Ludema (1985).
In an experimental study of vane pump wear Tao and Appeldoorn (1969)
investigated the effects of oleic acid and stearyl amine on different metal parts of the
pump. They observed that oleic acid performed best on iron oxide, which is basic in
nature, whilst stearyl amine is best on silica which is acidic in nature. No reduction in
abrasive wear by oleic acid was observed when a steel ball was loaded against a
grinding wheel. The experimental study revealed that the reactivity of additives
depends on the acidic or basic nature of the mating surfaces. They also observed that
the abrasives coated with antiwear additives do not adhere to surfaces and remain
loose in the contact thus preventing cutting action.
Metal-to-metal contact may occur due to low film thickness as a result of change in
operating parameters or due to entrapment of solid contaminants in the lubricant. As
explained earlier, the main sources of solid contaminants are either self-generated
wear particles or environmental dust particles. Metal-to-metal contact can be reduced
by adding film forming antiwear agent to the carrier lubricant which helps to reduce
friction between the abrasive and the contacting surfaces and prevent adherence
between them.
Friction is directly responsible for wear of metals. There are two types of friction –
surface-to-surface friction and fluid friction. Surface-to-surface friction is due to
asperity contact between the two mating surfaces and fluid friction is caused by the
resistance to motion between the molecules of the fluid. Friction reducing additives
better known as friction modifiers help to reduce the coefficient of friction, which is
the best indicator of low wear. Viscosity index and pressure-viscosity coefficient are
two important properties of the lubricant responsible for low fluid friction.
The other factors responsible for low friction are Lansdown (1982):
• material combination and their miscibility in each other
• their solubility in base oil
• atomic size of metals contained in lubricants
• valency of elements
• molecular structure of materials
31
• electrochemical activity
• intermolecular forces
There are many antiwear additives available in the market but very few are available
in pure chemical form (generic). Most are sold in the form of an additive package
comprised of different chemicals to achieve desired properties of a lubricant. Each
constituent of the package is committed to perform its intended duty. Additive
technology is very complex and package design is the domain of chemical engineers
and chemists. Efforts have been made to find wear coefficients of several antiwear
additives and base oils Rowe (1970), Klaus and Bieber (1965); Forbes and Battersby
(1974); Forbes (1974) A number of wear models by Rowe (1967), Rowe (1970) and
Groszek (1962) under lubricated conditions were developed where the role of
temperature in adsorption and chemical reactivity of additives were taken into
account.
Din and Kassfeldt (1999) used environmental friendly rapeseed with synthetic ester
oil in a journal bearing lubricated under contaminated conditions. They recorded
high friction in these bearings but wear rate was very low. The possible explanation
for these results was that the abrasive particle size contained in the oil was large
compared to the film thickness in the bearings, which operated under mixed
lubrication.
2.4.1 Commonly used antiwear additives
The most commonly used additive packages in industry are identified by the name
of the prominent chemical group they contain, the most popular antiwear additives
known by the name of their chemical groups are: phosphate esters, sulfurized olefins,
sulfurized sperm oil, metal dithiophosphates, borates, phosphites and metal
dithiocarbamates. Properties of these additives are reviewed in brief to develop a
basic understanding of additives in this section.
Phosphate esters
The general formula of phosphate ester is O=P(OR)3 where R is alkyl group having
4 to 20 carbon atoms and a molecular weight between 200 to 1000 daltons. Zheng et
al. (1986) have measured the values of coefficient of friction and scar diameter on
32
a Falex test machine of some of the lubricants and compared them with O-tricresyl
phosphate ester (TCP). These lubricants are: molybdenum dithiophosphate
(MoDTP), Zincdithiophosphate (ZDTP) and sulphurised olefins (SO). The results
revealed that TCP is the second best wear reducing additive as shown in the Table-
2.2.
This synthetic ester has the advantage of being neutral, ash less and stable at high
temperature. As far as their effectiveness in reducing wear is concerned, it was
reported that a 2% (by weight) sample of triaryl phosphate (TAP) in base oil
produces a scar diameter of 0.43 mm and load wear index of 25.2 under the test
conditions specified by ASTMD2266 (2004) and ASTMD2596 (2004) standards
respectively. Warne (1985) reported that TCP is hazardous for health, which prevents
its further use as an antiwear additive.
Table 2.2 Wear reducing properties of various lubricants, Zheng (1986)
Sulfurised olefins (SO)
Sulfurised olefins form a sulphide film on the lubricated surface and act similar to
other antiwear additives like MoDTP, TCP and ZDTP but their coefficient of friction
is low Zheng (1986). The SO mainly act as an extreme pressure additive. However,
they also act as an antiwear and antioxidant. Due to the presence of sulphur, this
additive is useful for high temperature applications but free sulphur poses a risk of
forming acids when in contact with copper or its alloys.
Sulfurized sperm oil
Sulfurised sperm is well known for increasing the load carrying capacity of
lubricated contacting surfaces, but it is not generally used alone. By adding
chemicals like chlorine or sulphur compounds, they form iron sulphide and sulphur-
chlorinated fatty material which are softer and act like a solid lubricant. These added
compounds help to form a good lubricating film on a surface due to the affinity of
33
the polar part of the molecules. The sulphur and chlorine react better under pressure
and elevated temperatures; therefore at higher loads prove to be better load carrying
agents. When these additives are used with ZDTP they act as effective friction
modifiers.
The use of sulphurised sperm oil was very common before 1970 but now they are
banned. Their use has been replaced by certain alkyl esters. As stated above, the
effectiveness of this additive increases many fold when added with chlorine and
sulphur compounds but they pose a threat in terms of corrosion of the surfaces due to
formation of hydrochloric and sulphuric acids.
Metal dithiophosphates
The use of lead salts in lubricants has been known to be toxic, hence lead
dithiophoshate has been replaced with other compounds Jiusheng (2003). Currently
metal dithiophosphates are popular, as they are better in terms of their antiwear and
toxicity properties. These are available in different forms of dithiophosphates such as
dialkyl, diaryl and alkyl type of compounds. Some of the most commonly used
additives of this family are discussed below:
Zincdialkyldithiophosphate (ZDDP)
ZDDP additives are very popular in industry today, not only as antiwear additives
but their antioxidant and anticorrosive properties against copper-lead bearings are of
equal importance for the users Klamann (1985). These additives are also used in
combination with other additives such as TCP, amine salts of phosphate esters,
sulphur fats and olefins. These were originally used in automotive lubricants as an
antioxidant but their usefulness as anti-scuffing and antiwear additives was identified
later from experience in service. The antiwear activity of these additives is an
important factor which needs to be considered when temperatures are high, Jahanmir
(1986). Kulczycki (1994) studied the effect of viscosity on ZDDP tribological
performance and characterised the additive according to their thermal stability.
Though the mechanism by which ZDDPs reduce wear and the process by which wear
occurs in their presence is not well understood, Bell et al. (1992). Marina et al.
(1997) have investigated the ZDDP film formation mechanism in relation to a stable
34
film deposition of colloidal polyphosphate material. It was also found that at low
loads and temperatures, wear reduces two orders of magnitude whereas at increased
loads and temperature it increases by one order of magnitude, Jahanmir (1987).
Keeping in view the application of antiwear additives in fire hazard situations such
as: mining and metal working industry, aqueous antiwear additives were developed.
Most antiwear agents were oil soluble agents, which were dispersed in water using
surfactants. Two such long chain slats of zinc; polyoxyethelyne glycol phosphate and
polyethylene glycol thiophosphate (ZOP and ZTP) developed by Feng et al (1995)
showed good wear resistance in fire resistant lubricants. However, in the presence of
water, ZDDP is less effective as an antiwear additive due to hydrolysis.
Winer (1967) indicated that zinc is the secret to ZDDP’s performance; however the
dithiophosphate part of the molecule is equally, if not more, important. This additive
can be classified in two groups: neutral and basic. For primary and secondary ZDDP,
the former is mainly responsible for thermal stability whereas secondary ZDDP is
more effective as an antiwear additive. In fact it decomposes to hydrogenperoxides
and thus inhibits oxidation and modifies the surfaces in contact. In an experimental
study Liston (1992) has demonstrated that three types of ZDDP discussed above can
be chosen as either EP, antiwear or antioxidant additive depending upon the
application.
A number of researchers Khorramian et al. (1993) and Kapsa (1981) have
investigated antiwear properties of ZDDP of different types in combination with
different types of other additives such as detergents and antioxidants. They found
that these additives hamper the effectiveness of ZDDP.
Researchers have used Scanning Electron Microscope (SEM) and Scanning Auger
Microprobes (SAM) test rig; Marina (1997), to find out the antiwear mechanism of
ZDDP. They observed that it reacts with the contacting surface and forms a thin
layer. In another study the antiwear property of ZDDP was attributed to adsorption of
these additives, forming a layer of a friction polymer (Kawamura ,1983).
High contents of phosphorus in ZDDP can cause deterioration problems in the
presence of other additives. Researchers recommend that in the case of crank case
35
oils where catalysts are present, the percentage of phosphorus needs to be below 0.1
% by weight, Khorramian, Iyer et al. (1993)
In an experimental study Glaeser (1992) concluded that sub-micron wear particles
found in many lubricated wear conditions represent a large surface area in a
lubricated wear situation, which can become a major factor in the depletion of the
lubricant antiwear additive.
In an experimental study Kano et al (2003) have found that ZDDP forms a higher
shear strength film than MoDTP. They also found that the film thickness depends
upon the type of additive, its concentration, the contact temperature and the rate of
heat removal due to wear. The antiwear property of these additives is associated with
the film thickness formed and scuffing occurs when the protective films have been
worn off.
Molybdenum dithiophosphate( MoDTP )
It was in 1878 that molybdenum was finally and clearly distinguished from its base
metal but it was readily available with reasonable purity in 1918. The real turning
point came in 1934, when low friction properties of molybdenum disulfide were
evident. Bell and Findlay (1941) reported the first successful operation of this
additive. A number of reviews have been published on this topic of growing interest,
for example Winer (1967) and Lansdown (1999) who recently published a book on
this subject. Mo S2 is used in dry powder, dispersion and compound chemical forms.
It is very widely used in the space industry as a solid lubricant and as a dispersed
additive in greases and oils for automotive and industrial applications. There are
many other forms in which MoS2 can be used, such as burnished coatings, sputtered
films, bonded films, composites and pastes.
According to Bell and Findlay (1941) the coefficient of friction of MoS2 varies with
several factors such as gaseous environment, humidity, temperature, load and purity,
state of orientation and consolidation of the film. It is not possible to incorporate
these parameters simultaneously while measuring the coefficient of friction
experimentally. There is controversy about repeatability of coefficient of friction
values reported by experimenters. However, the friction is at its lowest for fully
36
ordered surface films in dry air or vacuum at high load and highest for randomly-
oriented films in the presence of water vapour or certain other vapours at low load.
This proves that MoS2 is less effective in moist conditions. However, it is almost
impossible in most cases to compare the results published by different investigators.
MoDTP is the most popular form of MoS2 to be used as a friction modifier in
industrial and automotive industry. Experimental studies by Zheng (1986) have
demonstrated that in engine oils, fuel consumption can be reduced by 2-5% with the
use of MoS2. This additive is better known as a friction modifier than an antiwear
additive. MoDTP decomposes with time and forms a MoS2 film on the contacting
surface, which is a low shear strength film and hence reduces friction significantly.
However, under dynamic conditions, these films are effective when, the contact
pressure between the two surfaces is below the critical pressure.
In Falex tests, Zheng (1986) found that the oil-soluble sulfurized oxymolybdenum
di-(2-ethylhexyl)-phosphorodithioate (MoDTP) has excellent friction-reducing
behaviour. He also observed that when MoDTP is used as an additive, the asperity
tips of the rubbing surfaces wear gradually and many micro-terraces, occur on the
surfaces, thus the rubbing surface becomes smoother
Antimony dialkyldithiophosphate (SbDTP)
The basic structure of SbDTP is similar to ZDDP and MoDTP, except that Sb
replaces the metal part of the molecule. This additive has better antiwear and
antioxidant properties (Khorramian et al, 1993) than ZDDP and MoDTP. It can also
act as an EP additive but requires a higher number of carbon molecules in the
lubricant to ensure its solubility.
Gold dihydrocarbylphosphorodithioate (AuDPD)
AuDPD prolongs the antiwear property of lubricating oils as reported by Khorramian
et al (1993). The salts deposit close strongly bonded metallic coatings of gold
directly on the metal surface being lubricated as a fixed, almost permanent, solid
lubricant. These solid films formed on the surfaces prevent the wear of the substrate
when moving one with respect to the other. It has been found that gold
37
phosphorodithioate is superior to the corresponding zinc salt.
Phosphite Compounds
Phosphites are a derived form of phosphorous compounds and have proved to be
good antiwear agents. However, the mechanism of their wear reducing property is
not yet well understood. Researchers observed that phosphates or phosphites
containing long chain alkyl groups, formed smaller wear scar diameters than those
containing branched or aryl groups in a wear test (Khorramian et.al., 1993).
Metal Dithiocarbomate (DTC)
It is observed in automotive applications that in the presence of some catalysts,
phosphorous causes deterioration of the surfaces. In order to overcome this problem
dithiocarbomates (DTC) of some metals were developed. The general chemical
formula of these chemicals is (R2NCS2)x M where M represents Metal such as
Mo,Zn, Sb, Pb and R is alkyl group of 1-22 carbon atoms and x is the integer 1,2 or 3
depending upon the metal used. The main advantage of DTC’s is that they contain
reduced or no phosphorous and some are excellent copper corrosion inhibitors.
The antiwear mechanism of these additives is explained by Yuji and Gondo (1989)
MoDTC forms a film-composed mainly of MoS2. The coefficient of friction
decreases with increase in the percentage of MoS2 on the surface film. MoDTC has
better film forming capability than MoDTP and is therefore superior. A lubricant
containing ester of a polycarboxylic acid and glycol or glycerol with the selected
metal DTC derivative enhances friction-reducing properties even further. Zn DTC
helps inhibit corrosion and wear more effectively when combined with detergents.
Borates
Various types of borates are used, mainly as EP additives. However there are a
number of borate additive compositions which are useful as friction modifiers,
antioxidants and corrosion inhibitors. In general borates compounded with little or no
phosphorous are preferred as antiwear additives. In low phosphorous containing
lubricants it is preferable to use boron-containing compounds derived from hydroxy-
38
containing ester, Kawamura (1983) and Khorramian et al (1993).
Borate containing additives are limited if they are to be used in the presence of
moisture as they crystallise and form solid hard granules. These hard granules act as
abrasive particles and cause severe wear of the surfaces. Experience has shown that
alkali metal borate dispersion slowly forms solid deposits on the shafts near the seals.
Thus during the motion, seals get abraded causing leakage of lubricant which may
lead to severe damage. Borates also create a compatibility problem when used in the
presence of other additives such as phenates, sulphurised fats and ZDDP.
In a study manometer zinc borate of 20-50 micron, antiwear, antifriction properties
have been examined by Dong and Hu (1998). It was found that this additive gave
low wear and friction coefficient and a significant increase in load carrying capacity.
He examined the surfaces under SEM and found that Diboron trioxide, FeB and FeB2
films were formed which helped to reduce the friction between the surfaces.
2.4.2 Lubricated wear and characterisation of additives
The primary function of a liquid lubricant is to reduce friction and wear. The
secondary functions are to carry wear debris and heat away from the contacting
surfaces. However, the acceptable liquid lubricant must have many other properties
which may be tailored by selecting a suitable additive package. A lubricant prevents
metal-to-metal contact between the asperities of two mating surfaces either by
forming a thick film in the hydrodynamic or elastohydrodynamic regimes, surface
films by physical or chemisorption, tenacious surface films by chemical reaction or
by prior treatment of the surfaces with some low friction material to form a solid film
or coating. The total load on the bearing surfaces is shared by the lubricant film,
surface film and by metal to metal contact depending upon the type of lubrication
mechanism, Thompson (1971).
Significant research has been carried out on the mechanism of surface interaction
between lubricants and additives. Many additives react with metal surfaces and the
energy of desorption of the surface reaction product becomes the controlling factor in
the effectiveness of the additive. There is a sufficient evidence Groszek (1962) that
energy of adsorption and desorption is responsible for the effectiveness of polar
39
additives in preventing wear. Wear results for graphite in the presence of various
gases also were described by Rowe (1967). Later, the model was extended to the
competitive adsorption of lubricant additive and base oil molecules Rowe (1970).
Researchers were attracted towards wear in 1950s when first mathematical
relationship Archard (1953) was established between wear volumes and operating
parameters; load, sliding distance (l)and hardness as shown in Equation 2.1.
There have been controversies about this equation as this was preliminary or
rudimentary in nature. The same model has been extended with modified wear
coefficient ‘K’ as wear in boundary lubrication regime. There is a school of thought
who feels that the friction in the boundary regime needs to be considered in
conjunction with viscous shearing of the substance where asperity contact does not
exist. Hence the lubricated ‘Km’ value needs to be some fixed fraction of K value of
dry sliding. The modified reactive lubrication is commonly used for lubricated wear
and is a modified Archard’s equation:
m
m
PK
DV α
= (2.6)
Where comparing in Archard’s Equation 2.1; Pm is the hardness, and
KK m =α.
0UtRT
EXs
e ⎟⎠⎞⎜
⎝⎛−
=α (2.7)
Xe = diameter of area associated with an adsorbed lubricant molecule
to = fundamental time of oscillation of the molecule in adsorbed state
E = energy adsorption
U = sliding speed
Ts = surface temperature
R = gas constant
The new components wear faster because initially there are more asperity contacts
and the temperature is not enough to let the additive film form, Peterson (1980).
40
There is no standard method developed to characterise an additive. However,
depending upon the purpose, an index may be derived of the additive. This may not
be possible for an antiwear additive because different antiwear behaviour is difficult
to predict. However lubricated wear of different metal pairs has been expressed in
terms of a wear coefficient, Peterson (1980).
Wear coefficients for different lubricants and additives have been reported in the
Wear Control Handbook (Peterson and Winer, 1980) with data obtained from a four
ball test. The values of K vary from 10 -6 to 10 -8. The pressure and speed for the test
conditions were far greater than that for sliding bearings so the values may not be
suitable for journal bearings. Blau (1997) reviewed 70 years research on the subject
of wear. He concluded that wear is broad in nature and the process itself very
complex with the variables chosen being either insufficient or not directly relevant to
the designer.
2.4.3 Bearing performance measurement techniques
Literature revealed that there are various methods for reporting the tribological
performance of machine components. Frictional losses and the loss of material due to
wear are the two main tribological performance aspects. There are various indicators
by which the performance of a machine can be represented, such as; temperature rise,
vibration level, and pressure drop in oil feed system. Tribological performance of a
machine can be determined either by loss of energy due to friction between the
machine components, or by the loss of material due to wear in them Peterson and
Winer (1980).
The wear in a journal bearing relates to its life, and there are several methods by
which this wear can be expressed. Weight loss is a direct method by which wear in
bearings can be expressed. However, there are several other methods by which wear
in bearings can be expressed qualitatively, such as: change in bearing dimensions
such as radial clearance, change in surface roughness, wear particle count, and wear
scar diameter. However, minimum oil film thickness is one of the key bearing
performance parameters that gives the tribological history of a lubricated contact,
and closely correlates to the wear and friction status in the bearing Peterson and
41
Winer (1980).
In recent years, a number of non-conventional methods have been used for
determining the wear performance of a bearing, especially in bearings lubricated
with oils containing solid contaminants. Yuan et. al., (2004) have used vibration
analysis to correlate the concentration of the contaminants with the bearing vibration
level. Peng et.al. (2005) have also studied the problem on a worm gear test rig and
monitored and compared the wear debris analysis results with the vibration level.
They found that there is a strong relationship with the mode of wear and wear
severity with the vibration spectra. In a recent experimental study Maru and Castillo
et. al. (2007) used vibration measurement technique for correlating the damage in
rolling element bearings, due to solid contaminants. They recorded a change in root
mean square (rms) values in high-band frequency range of 600-10,000 Hz,
depending upon the type of contaminants and the wear damage. The tests were
simulated on a tribometer.
2.5 Effects of Micro-geometry on Bearing Performance
Micro-geometry of a journal bearing refers to radial clearance, surface roughness and
the out-of-roundness. These parameters change as wear progresses in the bearing.
The size of the minimum oil film thickness and change in micro-geometry are of the
same order of magnitude. The size of solid contaminants varies and many particles
are of the same order of magnitude as minimum oil thickness or micro-geometry
parameters such as: out-of-roundness, roughness. The effect of solid contaminants of
the size close to minimum oil film thickness is very harmful on bearing’s wear and
lubrication.
2.5.1 Roughness effects in hydrodynamic bearings
A theoretical consideration of effect of roughness in hydrodynamic bearings was first
studied by Tzeng and Saibel (1967) and subsequently by Christensen (1970). The
motive was to understand the effect of straighted roughness on film thickness, load
carrying capacity and friction in bearings. Roughness on the bearing surfaces was
considered one dimensional and modified film thickness (HT) in the Reynolds’
equation was substituted as:
42
δ+= nT hH (2.8)
Where hn is the nominal film thickness and δ is the excursion of the profile from the
centre line which is a random variable and for new surfaces, the probability density
function was considered to be a Gaussian distribution. Tzeng and Saibel (1967)
predicted that the load carrying capacity of such bearings is high and frictional force
in an infinitely long bearing is low. They compared smooth and rough bearings with
film thickness defined as h and hn respectively and presented their theoretical
solution for transverse roughness case. Later Christensen (1970) extended Tzeng’s
model for the longitudinal roughness case. The conclusions from their studies are
given in Table 2.3.
Table 2.3 Results for infinitely wide bearing (Christensen 1969-70)
The modifications of Reynolds’ equation did not yield a simple equation applicable
for rough surface. This renders these theories impracticable and difficult to apply in
engineering design and practice. Patir and Cheng (1978) have proposed a general
theory for an arbitrary surface pattern.
A number of researchers developed algorithms and numerical methods to either
simplify the solution or to predict the film thickness more precisely than previous
workers for different roughness orientations (longitudinal, transverse or random).
The concept of a moving and stationary roughness Christensen (1970), Tzeng and
Saibel (1967) and Tonder (1986) was also introduced by researchers but the
predictions did not fully agree with experimental results produced by them.
However, theoretical models of Patir and Cheng (1978) proved to be more practical
than the others. The salient features of this model are:
Any roughness pattern on the lubricated surface reduces the lubricant film thickness,
43
the film thickness decreases with an increase in load but the effect of the roughness is
independent of the load, the position of the roughness does not influence the bearing
performance (stationary or moving roughness), longitudinal and transverse roughness
pattern have the same effect on the bearing performance for a given size, and the
influence of roughness on the lubricant film did not depend on bearing internal
clearance or speeds.
In late 80’s some researchers Mitsuya et al. (1989); Mitsuya and Ota (1991) gave a
different concept of film thickness measurement by determining a different datum
line for measuring the film thickness and showed that they had better agreement in
their
experimental and theoretical results. Their studies were related to gas bearings; hence
their research did not gain recognition for conventional bearings.
Zielinski (1997) agreed to film thickness measurement method in rough bearings
approved by Mitsuya et al. (1989) and Mitsuya and Ota (1991) Furthermore, he
developed a new model based on the concept of material removal from the rough
surfaces.
2.5.1.1 Concept of Film Thickness and Material Removed from the Original
Surface
The approach used by Zielinski (1997) differs from other researchers due to the
selection of a datum line from which film thickness in rough lubricated contact is
measured. Existing surface parameters such as Ra and γ cause an ambiguity as two
surfaces having the same Ra values can be significantly different. Therefore these
parameters cannot be reliably used. Asperities above the centre line or mean line
cause restriction to flow whereas valleys below the line enhance the flow (provided
there is a pressure gradient).
2.5.1.2 Modified Bearing Surfaces
In a new concept in design to protect bearings from solid contaminants a two-
component surface layer was manufactured. Applying this concept to a journal
bearing, a soft layer of metal was applied as a corrugated coating on the surface of a
44
journal. This geometry helped in excluding the contaminants from the bearing
contact. In a recent attempt Sep and Kucaba-Pietal (2001) and Sep (2004)
demonstrated that having a layer on the journal with soft metal with microgrooves
bearings can be made up to five times more wear resistant. The component surface
layer design of bearings is based on a concept of preferential path suggested by
Hirstch et al. (1980).
2.5.2 Worn journal bearing analysis
Journal bearings ideally operate in hydrodynamic lubrication regime and so they do
not wear. However, at the start when speed is close to zero, theoretically the film
thickness in the contact zone is zero and hence metal-to-metal contact occurs and
causes wear in the bearing. Normally the wear rate in such bearings is low and the
extent where the bearing is considered to have failed takes a long time. Under ideal
conditions it may last for more than ten years. Problems related to misalignment,
improper lubricant selection, abusive operating conditions and hostile environment
may cause the bearing to fail much earlier than expected. Duckworth (1957)
analysed wear in journal bearings. Forrester (1960) derived similar conclusions.
The change in micro-geometry of a bearing due to wear affects its performance and
causes a reduction in the load carrying capacity and the minimum oil film thickness.
The literature review revealed that very limited work has been done to study the
effect of change in geometry such as wear groove and change in radial clearance.
Previous studies have been performed on the transition from hydrodynamic to
boundary lubrication as speed of a six-inch bearing were reduced in the laboratory,
Elwell (1977) They recommended that the film thickness in the bearing needs to be
ten times the composite roughness of the bearing and the shaft.
Radial clearance determines the bearing performance characteristics.
In their work Chu and Kay (1974) on journal bearings found that the optimum
clearance has a direct link with the bearing diameter. A theoretical trend of minimum
oil film thickness to a radial clearance ratio plotted by them is shown in Figure 2.4.
Film thickness vs bearing radial clearance graphs can be plotted for various values of
bearing parameter (ηU/W). These graphs tend to have the shape depicted in Figure
45
2.4. Safe operation of a bearing is expected only when the minimum oil film
thickness in the bearing is above 95% of the highest value obtained from this graph
can be expected when the operating minimum oil thickness is 95% to 100% of the
optimum minimum oil film thickness. Higher or lower values of radial clearance than
the optimum values may cause a reduction in film thickness.
The metrology associated with the roundness of a bearing affects the effective inside
diameter (ID) and outside diameter (OD) of the bearing and journal respectively.
The bearing or shaft may appear to be round – it may have a constant diameter, when
measured with the micrometer, but when the shape is greatly enlarged, scale it could
be as shown in Figure 2.5, Bagnel (1978). It is hard to achieve a perfectly circular
bearing or shaft. Depending upon the roundness a reference centre can be achieved
with the radii in all the directions varying and thus the radial clearance is different at
each point along the periphery. It is important to know the accuracy of the
measurements and its influence on performance. Since bearings operate on a film
that is a few microns thick the out of roundness needs to be within tight limits.
Figure 2.4 Clearance ratio and film thickness relationship Chu (1974)
46
Bearing failures due to various reasons have been studied by Katzenmeier (1972)
who found that improper lubrication is the major factor responsible for bearing
failures. Greenwood (1970) reviewed the research work of Tallian (1964), and
suggested that the minimum oil film thickness in a contact needs to be at least three
times higher than the value predicted by Tallian’s model. The reason given was the
electrical and thermal contacts which could lead to wear at the film thickness
obtained from the design process. Gradner (1978) also studied the effect of bearings
operating on low Sommerfeld number. In a similar study, Mokhtar (1978) presented
the effect of start and stop speeds on journal bearing wear.
Figure 2.5 Out of roundness magnified part of the edge, Bagnel (1978)
The effect of wear geometry was studied by Dufrane and Kannel (1989); Dufrane
(1983).They showed that, in a steam engine bearing, wear takes place under the load
point and suggested a geometrical model taking into account the worn region of the
bearing for calculations of pressure and film profile based on Reynolds’ equation.
They stated that an optimum amount of wear exists beyond which the altered
geometry would accelerate the wear. They suggested that the wear groove size and
shape depend upon the maximum wear depth δ0. They suggested two models for
simulation of the shape of the worn bearing region as follows:
47
⎥⎦
⎤⎢⎣
⎡⎟⎠⎞
⎜⎝⎛+−= θδδ cos1 0
CC (2.9)
( )θδδ cos10 −−=CC
(2.10)
Where δ0 is the wear depth at an angular coordinate θ and ‘C’ is the radial clearance
and δ is the non-dimensional wear depth. They also considered the effect of pressure
on viscosity of the oil. Both the models can be used with minor change in precision
value. However, the shape of the wear pattern may be the guiding factor for selecting
the model to achieve precision.
Hashimoto et. al.(1986), and Dufrane et. al.(1983) analysed the influence of the wear
defect on the pressure field and also on the eccentricity ratio. They showed that wear
defect damages bearing stability and that low L/D ratio bearings were less sensitive
to a defect. Vaidyanathan (1991) studied four bearings with different geometries for
their influence on parameters such as friction, pressure and Sommerfeld number.
Scharrer (1991) demonstrated that small wear defects have only slight influence on
dynamic coefficient of a hydrostatic bearing. The stability of a bearing was also
studied by Tanaka (1994) and Kumar and Misra (1996). They studied the
hydrostatic bearing stability and extended the work of Suzuki to show that the wear
defect decreases the stability even at light loads. In another study in the same year
Kumar and Misra (1996 b) analysed noncircular journal bearings operating in
turbulent flow and showed that the wear defect increases the flow rate and friction
and reduces the load capacity of such bearings.
In subsequent studies Mizuhara et al. (2000) extended the research by showing that
the friction increases in the bearing lubricated with oil containing solid particles
however the effect disappears as speed increases.
In a recent study Fillon (2002) has considered wear grooves of different radial
clearance and maximum wear depth ratios and determined pressure distribution, film
profile, temperature rise, oil flow rate, power loss and film thickness profile. He
concluded that worn bearings do not always present disadvantages but may have
some advantages such as reduced temperature rise as the geometry of such bearings
48
may come close to a lobed bearings. His solution is purely theoretical without any
experimental support.
2.6 Knowledge gaps
Thorough literature review was carried out to contextualise the aims and objectives
of this research project within the current state of art in this area. This ensured that
the research carried out in this project adds to the existing body of knowledge. The
following knowledge gaps were identified through literature review:
• Lack of detailed study on the effect of solid contaminants treated with
antiwear additives on journal bearing wear
• Need for characterisation of antiwear additives based on their efficacy for
dusty applications under hydrodynamic lubrication
• In-depth knowledge of the effect of solid contaminants on the bearing micro-
geometry, and its effect on the bearing’s tribological performance
• Lack of any standard numerical parameter for classifying the performance of
antiwear additives operating in dusty hydrodynamic lubrication conditions
2.7 Conclusion
The literature review included areas of research such as: wear mechanism,
contaminants types, sources and their impact on wear, journal bearing design,
lubrication and wear, bi-polar lubricants, bearing micro-geometry, antiwear
additives, and the effect of roughness on bearing lubrication.
The main focus of the literature review was on the performance of journal bearings
lubricated with oil containing solid contaminants treated with antiwear additives. The
reviewed literature covered knowledge on issues such as types of contaminants, their
morphology, concentration and hardness. Furthermore, it covered their impact on
journal bearing wear.
Literature on bi-polar lubricants explained the current hydrodynamic theory and its
application to contaminated bearings. Significance of oil film thickness and latest
measurement techniques were also reviewed in for oils containing solid contaminants
49
treated with antiwear additives.
However, researchers have not studied thoroughly the effect of additives on bearings
lubricated with oil containing solid contaminants. The effects of change in micro-
geometry parameters such as out-of-roundness and radial clearance due to
contaminants treated with antiwear additives have also not been studied extensively.
Thus the literature review was useful in refining the research problem by making use
of existing knowledge base and finally in preparing a strategy to meet the aims and
objectives of this research.
50
BLANK PAGE
51
CHAPTER-3
3. EXPERIMENT DESIGN AND DEVELOPMENT
3.1 Overview
Experiment design comprised the road map for the experiment setup and procedures
used in this project, and this led to the development of new practical and theoretical
methodologies for improving bearing design and metrology.
This chapter comprises of two parts, in the first part all the experimental
requirements were identified; including identification of performance parameters,
procurement of materials, instruments, instrumentation and their setting up, and also
some preliminary test results. In the second part, a preliminary study on micro-
geometry of bearings and issues related to them were investigated. This resulted in
identifying some problems and finding their solutions.
Journal bearings are one of the most commonly used tribological components that
are not sealed or shielded. This results in solid contaminant ingress in the lubricated
contacts of the bearing. These contaminants cause wear and as a result micro-
geometry of the bearing changes, which results in poor tribological performance and
reduced bearing life. These problems are more severe when a machine works in a
dusty environment. The use of antiwear additives is common in the industry to
enhance the performance and the life of such bearings. Although, antiwear additive
manufacturers claim superiority of their product, no established method exists by
which their claims can be verified.
After reviewing the literature and analysing the problem, it was concluded that there
is a need to study the effect of solid contaminants treated with antiwear additives on
journal bearings, and specifically, on their micro-geometry – i.e. surface roughness,
roundness and radial clearance. Furthermore, there is a need to develop a systematic
methodology for determining the efficacy of antiwear additives.
To achieve these objectives, the following factors had to be considered (i) bearing
52
design (ii) types of antiwear additives (iii) amount and types of contaminants, (iv)
antiwear additives, (v) operating parameters, (vi) wear measurement techniques, and
vi) available resources such as test rig, time, money, expertise, and measuring
equipment.
There are no standard criteria to select methods and techniques for measuring wear
performance of journal bearings, hence a suitable criterion for defining the
performance of a journal bearing was developed. Parameters were identified to
represent the wear performance of the journal bearing. These parameters included:
weight loss, change in micro-geometry, and change in particle count.
Bearing micro-geometry has been used in this research for tribological performance
testing. The main micro-geometry parameters used include: out-of-roundness, radial
clearance and wear depth; the effect of these parameters on oil film thickness has
been used as the tribological performance measure.
This chapter presents the experimental details including bearing design, selection of
input and output parameters, test rig modifications and instrumentation, selection of
additives, metrology of micro-geometry, wear measurement techniques and
measurement of minimum oil film thickness, to examine the combined effect of
contaminants, additives and micro-geometry. Strategy for collecting and analysing
data was also prepared as a part of experimental design.
3.2 Identification of Performance Parameters
There is no single parameter by which the performance of a bearing can be
measured. The tribological performance of a journal bearing ultimately relates to its
capacity to conserve energy and materials. These can be measured by the friction
forces and the wear in the bearing under the stated operating conditions. The
tribological performance of a machine component can be compared with other
components under the same operating conditions rather than measured in absolute
terms. A machine may comprise of several bearings so that it is difficult to accurately
measure the frictional loss in an individual bearing. Similarly the amount of wear in a
bearing is usually so small that it is difficult to predict the life of a bearing. Moreover
there are several methods by which wear in a bearing can be measured but there is no
53
surety that the results obtained from different methods will produce the same
magnitude of wear. Another difficulty is that the repeatability of test results is poor.
Thus a multi-wear parameter approach where several wear measurement parameters
have been recorded and compared has been adopted to address this problem. Due to
the reasons stated it is also not possible to measure the friction and wear directly but
some parameters have been identified that represent the tribological performance of
the journal bearings subjected to test in this study.
3.2.1 Parameters as measure of energy conservation
The direct method of assessing the energy losses in an operating bearing is to
measure the friction force in the bearing. Attempts were made to measure the friction
force in an operating bearing but it was concluded that unless a sophisticated system
with air bearing support is used to measure the small change in friction the
distinction between the friction forces in the bearing under the influence of different
antiwear additives cannot be made.
Since friction force measurement in a bearing is a complex procedure, common
practice is to measure and compare parameters such as temperature rise, oil pressure
drop, vibration and noise. The temperature rise in a bearing indicates that the friction
in a bearing is high. However, the rise in temperature may be due to viscous shearing
or due to metal to metal contact and it is hard to ascertain the contribution of each
factor individually. Reduction in inlet pressure indicates that the load carrying
capacity of a bearing has reduced if all other operating conditions are kept the same.
This may be due to increase in radial clearance or wear in the bearing which in turn
influences the bearing performance in two ways; a reduction in bearing pressure and
an increase in lubricant viscosity due to cooling of bearing as a result of excessive
side leakage. Vibration and noise measurement are commonly used condition
monitoring tools. Change in vibration mode or level in journal bearings is mainly due
to misalignment, severe wear marks on the bearing surface, entrapment of a large
solid particle or transfer of fluctuating forces from the surrounding area. It is difficult
to identify the contribution of each of the influencing parameters.
Performance testing of bearings requires exactly the same operating conditions for all
additive cases including inlet oil temperature (40 0C) and oil feed pressure 100 kPa.
54
Therefore neither of these two parameters could be used as performance measures.
Since contribution to vibration and noise may be from sources within or without the
bearing components these may prove to be a very complex method for investigating
the minor changes due to additives.
With cognisance of above limitations and those of the test rig used for the
experimental study, other fundamental aspects of journal bearing lubrication were
examined and this helped to identify the parameters used to represent the bearing
performance.
A bearing operating in the hydrodynamic lubrication regime, where specific film
thickness (lambda ratio) is between four and ten, an optimum oil film thickness is the
key to the desired performance of a bearing. A change in friction force, rise in
temperature or drop in bearing inlet pressure are directly linked to the oil film
thickness in the bearing contact zone. Change in performance parameters occur
when there is a change in oil film thickness at the contact. A bearing will run
satisfactorily as long as an optimum specific film thickness at the contact exists.
Although a temporary change in operating conditions such as inclusion of solid
contaminants, change in viscosity or temperature may compensate the effect of one
parameter over another, the final impact on bearing performance is the oil film
thickness in the contact. Hence the change in minimum oil film thickness in a
bearing can represent the influence of an individual antiwear additive or of
contaminants on the performance. The minimum oil film thickness is a small
quantity but with current techniques it can be reliably measured in a bearing with an
accuracy of a few microns. Though the measurement of oil film thickness in a
bearing of an operating machine may not be practical, it can be easily achieved in a
laboratory setup.
3.2.2 Parameters as a measure of bearing life
A tribological component is said to have failed when it is unable to perform its duty
to the accepted level of performance. The geometry of a journal bearing is the key to
achieve desired oil film thickness. When the geometry changes due to wear are
beyond an acceptable limit the bearing cannot carry the rated load and is considered
to have failed. A critical geometrical parameter of a bearing is the radial clearance. A
55
larger clearance than the design requirements results in a thinner oil film at the
contact thus increasing the friction and temperature rise with a subsequent further
reduction in film thickness. When the film thickness reduces to an extent where
metal to metal contact occurs, significant wear takes place and the radial clearance
increases further. This causes the film thickness to keep reducing with hydrodynamic
lubrication being replaced with the mixed then boundary regimes until seizure due to
scuffing takes place. A bearing may be able to form hydrodynamic film at low loads
but if it cannot sustain the rated load it is said to have failed. Thus the rate of wear in
a bearing correlates directly to the bearing life.
There are several methods for measuring wear and the most appropriate ones have to
be chosen. It is a well known fact that the wear tests have poor repeatability and
reproducibility and hence the results differ from one test to another, depending on the
test equipment. It may be difficult to establish which method gives the most reliable
results. A difficulty is that the amount of wear is very small in comparison to the
specimen size in which case the resolution of the measuring instrument becomes of
paramount importance. In such cases it is helpful to utilize more than one
measurement method and by comparing the results of one with another the reliability
of the measurement technique be established. It was decided therefore to explore
several wear measurement techniques for the bearing components. In this case the
selection of wear measuring techniques depended on the availability of resources
also. The wear measurements were categorised in the following three groups:
1. Weight loss: weight loss in both bearing elements
2. Change in micro-geometry: such as; out-of-roundness, roughness and
radial clearance, maximum wear depth
3. Particle counts: including wear debris weight
3.3 Journal Bearing Design
A requirement was that the bearing had to fit the existing test rig. A bronze bearing
with a steel shaft sleeve 40 mm nominal internal diameter (ID) and 40 mm nominal
shaft outer diameter (OD) with L/D ratio 1 was designed using the ESDU method
ESDU84031 (1996).
56
Figure 3.1 Bearing and journal drawing
BE
AR
ING
M
ATE
RIA
L LG
2 B
S 1
400
57
The dimensions of the bearing with tolerances are shown in Figure 3.1. A sample of
bearing design parameters is shown in Table 3.1 and hardness measurements on
bearing and shaft sleeve samples are shown in Table 3.2.
Table 3.1 Sample of test parameters
No. Parameter Value
1 Oil viscosity 0.042 Pa s (40 oC)
2 Inlet temperature 40 oC
3 Oil feed pressure 1 bar
4 Required min. oil film thickness 16 μm
5 Contaminant (Al2O3) size 16μm (nominal)
6 Required K value Close to 1
7 Sliding distance (l) 7536m
8 Contaminant concentration 4 g/l
9 Speed (varies) 400-550 rpm
10 Load (Fixed) 500N
11 Bearing weight 330 g approx.
12 Shaft sleeve weight (nominal) 178 g approx.
13 Out of roundness 2-10 μm
14 Surface finish required <1.0μm Ra
15 Particle counts Results vary
16 Additive type 5 different types
17 Radial clearance 80-90 μm
18 Probe scale factor 25 micron/V
19 ID bearing (tolerance) 016.0040+
− (mm)
20 OD shaft sleeve (tolerance) 0016.08.39 +
− (mm)
21 L/D ratio 1
58
Table 3.2 Hardness measurements on bearing and shaft sleeves
Bearing Part
Location 1 Hardness
Location 2 Hardness
Location 3 Hardness
Average Hardness
Steel Bush No.1 ASI 4143
27.5 HRC 27.5HRC 29.0HRC 28.0HRC
Steel Bush No.2 ASI 4143
27.0 HRC 28.5 HRC 28.5 HRC 28.0HRC
Steel Bush No.3 ASI 4131
29.5 HRC 27.0 HRC 27.0 HRC 28.15HRC
Bearing No.1 Bronze LG2 BS 1400
75 HB 78 HB 78 HB 75HRB
Bearing No.2 Bronze LG2 BS1400
81 HRB 74 HRB 74 HRB 76.3 HRB
Bearing No.3 Bronze LG2 BS1400
74 HRB 74 HRB 78 HRB 75.3 HRB
The bronze bearing and shaft sleeve materials were chosen as LG2 BS1400 and EN
ASI 4143 steel respectively. The hardness of the bronze varied from 74 to 81 Brinell
hardness (Rockwell 'C' scale). The shaft sleeve material chosen was low carbon steel
its hardness varied from 27 to 29.5 HRC. The hardness measurements for 3 bearings
randomly picked are shown in Table.3.2.
Initially the bearing was designed with 100 micron diametral clearance. When
preliminary tests were carried out, the wear marks due to misalignment occurred on
two opposite ends of the bearing and a larger clearance was chosen. It was decided
that the bearing would operate on low speeds and loads. A bearing with 180 micron
diametral clearance was found to be satisfactory. The bearing design called for a
lubricating oil of viscosity 0.042 Pa.s at 40 oC inlet temperature. Standard bearing
tolerances were maintained as desired for a bearing with chosen design parameters.
The calculations were also checked analytically. A sample of test conditions used in
59
this experimental study is given in Table 3.1. Similar analysis for different operating
conditions has been obtained using ESDU 84031 software program available on-line
and the results are included in the (Appendix-B).
3.4 Contaminant Selection and Characterization
Aluminium Oxide (Al2O3), which was used as the test contaminant, is one of the
most versatile refractory ceramic oxides and finds use in a wide range of
applications.
It is found in nature as corundum which exists as rhombohedral crystals with
hexagonal structure. The unit cell is an acute rhombohedron of side length 5.2Å and
plane angle ~55°. It is the close packing of the aluminium and oxygen atoms within
this structure that leads to its good mechanical and thermal properties.
This has very high melting point of 2015±15 °C and refractive index 1.765. The
product supplied by the manufacturer (Australian Norton Abrasive Pty Ltd) is
reported as 99.75% pure. An optical analysis on Scanning Electron Microscope is
shown in Figure 3.2. The grading of the powdered product is also given in Appendix-
A Table A3.2. The reported size of the particles is F 400 (17.3±1 microns mean
diameter). The bottle of sample used for the experimental study was labelled grade
F400 Micro-grit and 16 microns overall size.
The nominal size of the aluminium oxide powder was verified by Mastersizer 2000
supplied by Malvern Instruments. The powder was mixed with water to form slurry
for the analysis of size distribution. It can be seen that the actual particle size was
slightly larger than the 16 microns specified by the supplier. Since variation in the
measured and specified value was not greater the latter was accepted as the true
value for experimental work.. The reason for accepting the specified value was due
to doubts that powder may not have been well mixed before taking the sample and
the measurements could not be repeated. An Energy Dispersive X-ray Analysis
(EDAX) image of the Al2O3 shown in Figure 3.3 confirms the purity of the product.
60
Figure 3.2 SEM micrograph of Aluminium Oxide particles
3.5 Lubricant and Additive Selection
Selection of base oil is a part of hydrodynamic bearing design. As a first step of
bearing design, base oil viscosity and radial clearance are chosen for given operating
conditions by a rule of thumb. Then chosen oil viscosity is verified for design
criteria. If design criteria are not satisfied, either oil with different viscosity is chosen
or radial clearance is varied using iterative process till the design criteria are
satisfied. Viscosity is the most important lubricant characteristic directly responsible
for forming an optimum film thickness in the bearing for given operating conditions
However, a lubricant has to discharge many other duties apart from forming an
optimum oil film thickness in the bearing. Depending upon the application and
requirements oils are treated with additives. These additives are identified either by
their chemical names such as sulphur-phosphorous, MoS2, graphite etc. or by the
duty they discharge such as antiwear, friction modifiers, viscosity improvers, etc. In
this particular case, the base oil and antiwear additives were chosen by Fuchs
Lubricants Australia based on their long experience and they supplied the products
for the research.
61
3.5.1 Base oil
Fuchs supplied the base oil manufactured by Caltex Oil Company and the additives
by different additive manufacturers. The base oil belonged to paraffinic group,
ALOR 300 Solvent Neutral with a viscosity of 55.0 cSt at 40 oC, density 0.876 kg/L
at 15 oC and Viscosity Index 96. The lubricant contained 4g/L, powdered Al2O3 of
16-micron size as solid contaminants. Some experiments were also performed with
contaminant concentration 0.1 to 0.2 g/L concentration. The amount of wear was too
small and could not be measured reliably. Fuchs have found from experience that
solid contaminants in mining industry reaches more than 5 g/L. and hence 4g/L was
the chosen concentration.
3.5.2 Additives selection
There were a total of five additives supplied by the industrial research partner and
these were chosen from range which is widely used in the field. The reported
information is limited due to confidentiality policy of the manufacturers. Additive
details obtained from the supplier are given in Table 3.5. These antiwear additives
are commonly used in mining industry for lubricating sliding bearings.
Figure 3.3 EDAX elemental analysis of Al2 O3
62
Additive Treatment
The additives listed in Table 3.5 were mixed with base oil in recommended dosages
by following the safety instructions and procedures advised by the suppliers. The
standard procedure was as follows:
• Calculate the quantity of additive required for 4 litres of base oil either by
volume or by weight.
• Weigh the required quantity of contaminants
• Soak the contaminants in a petty dish with 10ml additive overnight.
• Take 4 litres of base oil and pour 3 litres of it in the sump of the test rig and
heat it to 60 0C and circulate it for 30 minutes in the test rig.
• Mix the quantity of additive with the sump oil.
• Mix the additive soaked contaminants in the sump of the test rig
• Rinse the additive container (beaker) and the contaminant container with the
remaining 1 litre oil to top up the sump to 4 litres.
• Let the oil circulate in the hot oil system for one hour before starting the test.
Figure 3.4 Test rig assembly
63
3.6 Test Rig and Instrumentation
An existing test rig (Figure 3.4) was used for this study. It required several
modifications to the loading and oil supply systems motor drive, housing and
proximity probe mounting system.
The rig has a provision to hold a stationary bearing of 40 mm diameter in a housing
floating at one end of a shaft supported on two rolling element bearings. The
designed test bearings were fixed in a 30-mm wall thick steel housing. The outer
surface of the bearing is slightly tapered with a bush in the housing. The bearing at
the non drive end of the shaft is guided through a pin so that its position aligns with
the housing holes for the proximity probes and the oil supply. This end of the shaft
forms a cantilever over which a test shaft sleeve is keyed. This sleeve fits inside the
test bearing which is mounted in the housing. Movement in the horizontal direction
is constrained by a check nut. The ID of the bearing was chosen in correspondence
with the shaft sleeve diameter to achieve the desired radial clearance. Test bearings
and shaft sleeves are replaceable.
Table 3.3 Antiwear additive properties
Additive** Lube Category Chemical Family Concentration*
A3 Hydraulic Oil additive Confidential 0.7% by volume @ sp.gr. 1.0
A4 Hydraulic fluid Aryl phosphate esters
1% by weight @ 0.89
A5 Hydraulic fluid ash less antiwear additive
proprietary blend, confidential
5% by Volume @ sp. Gr 0.95-1.055
A6 Gear and hydraulic oil antiwear, anti corrosion and friction modifier additive
Sulphur/ phosphorous based
2% by weight @ sp.gr. 1.69 (at 15.6 oC)
A7 Lubricant Additive Isopropyl Oleate; fatty acid, isopropyl ester
1% by weight @sp gr. 0.86
*Specific gravity measured at 15C. ** Product names were suppressed to maintain the manufacturer’s confidentiality.
64
The shaft is supported on two self-aligning rolling element bearings. Originally the
drive was via a constant speed a 1420 rpm constant speed 1200 Watt electric AC
motor, the speed of which could be changed with the help of three step pulleys. This
was changed to a continuous variable speed 3 phase AC speed control system giving
0 to 1420 rpm with an accuracy of ± 5 rpm. Further details of the test rig and some
preliminary results have been reported by Sharma and Hargreaves (2001) in
Appendix - A.
Loading system
The load was directly attached to a hanger attached to the bearing through an eye bolt
as shown in Figure 3.5.
Figure 3.5 Loading System
Loads
Journal
Hanger
65
Oil supply
The oil supply system consisted of a reservoir supplying oil to a peristaltic pump
below by a flexible tube. Flow rate and therefore oil pressure were controlled by pass
throttle valve in parallel with the main oil supply line. A constant temperature
heating system with a thermostat capable of ± 1 0C accuracy was fitted in the oil
reservoir. It was observed that the temperature recorded by the dial gauge of the
thermostat varied with the ambient temperature and so it was adjusted and set to give
40 0C at the closest possible location to the bearing oil inlet. The temperature at the
entry was monitored by a thermocouple connected to a handheld electronic
measuring unit. The oil entered through a flexible pipe connected to a nipple on the
bearing housing and the out coming oil was collected in a perspex container which
was connected to the main oil reservoir. The oil pressure was maintained at 100kPa
to assure full fluid film lubrication during the test. A type K chromel-alumel sensed
the temperature which was controlled by type PN-4BIC device. The controlled
temperature bath ensured oil supply to the bearing at 40 ± 1 0C. The oil circuit is
shown in Figure 3.6.
Figure 3.6 Oil Circuit
Journal bearing Main valve
Pump
Bypass valve Oil sump
66
Proximity probes for film thickness measurement
The bearing housing had a provision to mount two REBAM-3300 proximity probes
at right angles to each other for measuring the oil film thickness. The probes were
screwed into the housing and through the holes in the bearing at a minimum distance
from the target (shaft sleeve) in order to give a linear output. .This distance was 3
mm or 10volt output. The bearing also had a provision for mounting a thermocouple
at the bearing surface.
3.7 Multi-Wear Parameter Approach (MWPA)
Wear tests under lubricated conditions are very time consuming and the wear
measurements have very poor repeatability. During preliminary testing it was
observed that after running long duration tests the wear in bearing was small and
while repeating the test bearing failed without prior warning. There were two
potential strategies to improve the accuracy and saving time, first, to conduct long
duration tests and to repeat them till reliable results are obtained and second, to
conduct short duration tests and measure wear using different measurement methods.
Keeping in view the time factor and the resources available, the second strategy was
adopted. This required a methodology for selecting the wear measurement
parameters, which in this thesis is referred to as multi wear parameter approach
(MWPA). Using this approach the measurements were categorised in three main
groups, i.e.
a) weight loss
b) change in micro-geometry
c) change in particle count.
All performance parameters that could be measured confidently in the laboratory
were identified as discussed in section 3.2. In this research wear was measured by
different methods and test results were compared for their reliability and precision.
On the basis of consistency of results and principle merits, the best method was
chosen for characterisation of anti-wear additives. The MPWA is shown in Figure
3.7 and the various methods are discussed below.
67
3.7.1 Weight loss
Weight loss in the bearings and shaft sleeves was recorded by weighing the bearing
and shaft sleeve before and after the tests. The accuracy of the weighing scale was to
three decimal places in grams. All the bearing and shaft sleeve test specimens were
degreased by dipping them in hexane for five minutes and subsequently cleaning by
ultrasonics for ten minutes before weighing. It was difficult to repeat the results to
third decimal place accuracy. However after extending the test duration, accuracy to
two decimal places was obtained. The results obtained were useful for observing
metal transfer and determining the mode of wear of the components.
3.7.2 Out-of-roundness
Out of roundness was successfully used by Zeilinski (Zeilinski, 1997) and Martin
[Martin 1991] as a wear measuring parameter. The change in the out of roundness of
a bearing after the test correlated well with the amount of wear. The out of
roundness measurement process is time consuming and so it was decided to measure
it only for the test bearings since this was stationary part. Under loaded conditions,
wear occurs at the contact zone of the test bearing whereas; wear of the shaft sleeve
is virtually uniform around its circumference. Although the shaft sleeve out of
roundness may also change it is not as distinct as that of the bearing.
The Taylor Hobson Talyrond 100 was used for out-of-roundness measurements
Figure 3.8. These were taken at three positions along the bearing length, i.e., top,
middle and end. The measurement of out of roundness at the middle position was
difficult due to the oil feed groove and the holes for the proximity probes.
The standard probe of the Talyrond was modified to avoid the sudden jerk due to
holes in the bearing. These jerks caused displacement of the specimen and hence an
aluminium plate with a shallow groove of the size of the bearing OD was fabricated.
This plate could hold the bearing firmly in position. A closer study of the out of
roundness method revealed the superiority of this technique for measuring wear of
the bearings. It also raised several serious issues with regard to current metrological
practices in tribology. In fact, it raises the fundamental issue of accurately measuring
radial clearance.
68
Wear measures
Change in weight
Bearing wt. loss
Sleeve weight
loss
Wear debris Change in geometry
Debris weight
Particle count
Out of roundness Radial
clearance
Surface roughness
Max. wear depth
Sleeve Bearing
Circumferential Transverse
Transverse Circumferential
3.7.3 Radial clearance measurements
Radial clearance constitutes the micro-geometry of a bearing and is a key design
parameter for determining the load carrying capability. As a guide the radial
clearance for a bearing is taken as one thousandth of its radius. Smaller, clearances
generate greater load carrying capacity of a bearing for the same operating
conditions.
For studying the effect of change in micro-geometry on bearing minimum oil film
thickness, it is important that radial clearance in a bearing is measured with a fair
degree of accuracy especially when the anticipated minimum oil film thickness in the
bearing is small as every one micron change in radial clearance reduces the film
thickness by approximately 1%. The measurement of radial clearance for the test
bearings varied from one method to another and from one position in the bearing to
another by up to 50%, thus emphasising the importance of metrology for successful
bearing performance.
Figure 3.7 Multi-Wear Parameter Approach (MWPA)
69
The measurements of radial clearance were obtained by measuring the ID and OD of
the bearing and shaft sleeve respectively. The specimen was fabricated in the
workshop according to the drawing shown in Figure 3.1.
To confirm the dimensions the ID and OD of each test bearing was measured
separately to cross check the actual radial clearance. The design guidelines indicate
that the nominal ID and OD were 40.00 mm and 39.82 mm respectively, giving
minimal radial clearance 90 microns. However with tolerances the maximum ID
could be 40.00 + 0.016 mm and minimum OD 39.82-0.016 mm and hence the radial
clearance can fall any where between 90 microns and 106 microns. Thus
theoretically the radial clearance could vary more than 16 microns. If each micron
radial clearance added to nominal radial clearance of 90 microns will reduce the film
thickness by 1% (Chiu and Kay, 1974), it may compound the problem. Thus the
maximum variation in minimum oil film thickness can be reduced by 16% (2.5
microns) of the required minimum oil film thickness (16 microns) to 13.5 microns
which is approximately 84% of the required value.
Figure 3.8 Talyrond 100
70
When ID and OD of the test specimen were measured the measurements were taken
at 10 locations along the bearing circumference (5 at one end and 5 on the other) at
an interval of 36 degrees on either side covering the whole circumference. The main
focus was on the bearing ID measurements using four different instruments i.e.
vernier calliper, Sigmascope, Hole-test-gauge and Metroscope. OD of the shaft
sleeve was measured with the help of HP Laser system only.
The variations in ID measurements at different locations were observed more than
the tolerance limits. Even though the precision of each measuring device is known
the accuracy in measurement is not assured due to human errors in handling or
observing the readings. When diameters were measured on the same bearing sample
at different locations the variations in a bearing’s radial clearance were found to be
780, 760 and 25 microns respectively for these methods. These measurements
recorded at different locations along the bearing circumference of both ends are
shown in the Table 3.4. The highest and lowest values for each method have been
highlighted and an average value has been calculated for each measuring device. A
statistical analysis shown in Table 3.5 clearly states the superiority of ID
measurement with Metroscope (Figure 3.9) which works on the optical comparator
principle.
Figure 3.9 Metroscope for bearing ID measurements
71
The results show that the vernier measurements are the poorest in terms of accuracy
and Sigmascope results are the poorest for repeatability. With the latter there is
significant subjectivity involved in focusing the lamp and positioning the cursor.
The gauge test proved to be a better instrument but the Metroscope gave the best
results with an accuracy of up to 5 microns. The results of bearing ID measurements
have been discussed as an example but a similar situation occurred with the
measurements of shaft sleeve OD also. The use of all the above devices was
unsatisfactory in measuring the OD of the shaft sleeve and the best results were
obtained with the HP Laser System. The errors and problems associated with the
metrology of radial clearance measurements have been discussed in a separate
publication (Sharma , 2004).
Exaggerated end-view of these measurements has been plotted in the form of radar
graphs (Figure 3.10a-3.10e). The graphs 3.10e is the exaggerated radar plot of ID and
OD of concentric bearing and shaft sleeve together and highlight the metrological
problem associated with the radial clearance measurements.
Table 3.4 Bearing ID measurements
Observation No.
Vernier calliper
Sigmascope Hole-test- gauge
Metroscope
1 39.90 39.948 40.015 40.005 2 40.01 39.974 40.000 40.000 3 40.00 39.951 40.005 40.002 4 39.95 39.956 40.005 40.005 5 39.96 40.000 39.990 40.000 6 40.02 39.985 39.990 40.005 7 40.03 40.005 40.010 40.002 8 40.02 40.008 40.010 40.002 9 40.66 39.282 39.966 40.004 10 39.88 39.246 39.962 40.000 Average 39.986 39.9783 40.0031 40.0026
Figure 3.10e shows that there may be instances during the complete rotation of the
shaft when there is almost zero radial clearance in the bearing. As the film thickness
is the distance between the average lines of the roughness of the two mating surfaces,
72
the actual film thickness may frequently result in mixed lubrication, or boundary
lubrication regimes during each revolution of the shaft sleeve.
Table 3.5 Bearing ID measurements: statistical analysis
Instrument Standard Deviation
Median Mean
Vernier 0.222962 40.005 40.043 Sigmascope 0.302116 39.965 39.8355 Hole-test-guage 0.018421 40.0025 39.9953 Metroscope 0.002121 40.002 40.0025
3.7.3.1 Metrological issues
This study raises some serious issues pertaining to the radial clearance of the bearing.
It can be concluded from this study that the accuracy in radial clearance
measurement can be achieved by choosing the right instrument and precision
manufacturing process. No matter how accurate the measuring device is the diameter
of the bearing and shaft sleeve varies from one location to the other along the
circumference as well as along the bearing length. The out-of -roundness and
roughness varies from one location to the other and so too the diameters. This is
mainly due to limited control over the manufacturing process. The poor measuring
practice can cause a serious deficiency, if the error in measurement is of the same
order as operating minimum oil film thickness.
The current standard practice of choosing the average value for the diameter does not
guarantee that a desired oil film thickness is obtained at every instant through out the
bearing length. Designers should allow for these problems and researchers should be
pedantic in ensuring that they are dealing with actual dimensions (not specifications)
and thereby producing reliable research results.
73
Figure 3.10a Bearing ID profile by Vernier measurements
Figure 3.10b Bearing ID profile by Sigmascope
End view Vernier profile
39.739.839.9
40
12
34
5
6
7
89
1011
1213
14
15
16
17
18 19
20
End view Sigmascope profile
38.5
39
39.5
40
40.51
23
4
5
6
7
8
910
1112
13 14
15
16
17
18 19
20
Series1
74
End view profile hole test guage
39.9239.9439.9639.98
40
12
3
4
5
6
7
8
910
1112
13
14
15
16
17
18
1920
Series1
Figure 3.10c Bearing ID profile by Hole-test-gauge
End view Metroscope
39.99639.998
4040.00240.004
12
3
4
5
6
7
8
910
1112
13
14
15
16
17
18
1920
Series1
Figure 3.10d Bearing ID profile by Metroscope
75
Sleeve OD and bearing ID plot
3838.5
3939.5
4040.5
1
2
3
4
5
6
7
8
9
10
HP-laser data
metroscope data
Figure 3.10e Concentric bearing and shaft sleeve diameter graphs
Given the foregoing discussions it was difficult to select a single value for diameter.
The diameters were measured with Metroscope for the bearing ID and the HP laser
system for the shaft sleeve OD. Diameter was measured at different positions along
the bearing and shaft sleeve length and at equally spaced angular positions. Average
values were taken as the mean diameter. Both methods gave precision up to five
microns. A combined plot of Bearing ID data obtained from Metroscope and shaft
sleeve OD data obtained from HP laser system has been plotted and is shown in
Figure 3.10e. The plot indicates that under normal circumstances there are instances
where bearing radial clearance is almost zero.
The purpose of plotting these exaggerated profiles is to highlight the metrological
problems associated with bearings and the resulting impact on oil film thickness at
the contact.
76
3.7.3.2 Some observations about radial clearance measurements
Measurement of radial clearance is a key factor in this research, therefore, it is
imperative to measure the bearing ID and shaft sleeve OD with confidence.
However, the following aspects of radial clearance measurement need further
attention:
• The radius of the bearing varies along the circumference of the bearing,
leading to variations in the radial clearance.
• Accuracy in measuring the radial clearance depends upon the metrological
procedure and the precision of the measuring instrument.
• Specifying the radial clearance as the average of various measurements taken
along the circumference is not adequate.
• Due to out of roundness, the radial clearance of a bearing varies along the
circumference, as well as along the bearing length.
• Just as cut-off length is specified in measuring surface roughness, similarly,
out-of-roundness must be specified when radial clearance is reported.
• The rule of thumb that specific film thickness or lambda ratio 10 is adequate
for designing hydrodynamic journal bearings, needs to be reviewed in view
of change in bearing ID along its circumference
3.7.4 Gamma ratio for film thickness measurement
The out-of-roundness values recorded before the wear tests show that these are
higher as compared to surface roughness values. This indicates the bearing design
parameter called film parameter or lambda ratio need to be reviewed to ensure
adequate separation of the bearing surfaces during lubrication. Pursuing this matter
further a new parameter has been proposed.
3.7.4.1 Significance of film parameter
The film parameter popularly known as lambda ratio or specific film thickness is the
77
ratio between the oil film thickness and composite roughness of the two lubricated
surfaces, as shown in figures 3.11a and 3.11 b.
Where h is the oil film thickness and σ1 and σ2 are the rms values of roughness of
surfaces 1 and 2 respectively. It is important to note here that the film thickness ‘h’ is
the distance between the mean lines. Figure 311a shows an exaggerated view of this
concept. Alternatively, Figure 13.11b shows a simple sketch of rms values.
Figure 3.11a Oil film thickness between the surfaces
Figure 3.11b Oil film thickness based on composite roughness
The ratio of film thickness h, and the composite roughness of the two surfaces is
known as the film parameter, lambda ratio or specific film thickness. It can be
represented through Equation 3.1, where σ1 and σ2 are the rms roughness values of
surfaces 1 and 2 respectively.
h
σ1
σ2
h
78
⎟⎠⎞⎜
⎝⎛ +
=22
21 σσ
λ h (3.1)
Where: composite roughness ( )22
21 σσσ +=comp
The lambda ratio is a direct indicator of lubrication regime in which a bearing is
operating. The lowest value of lambda occurs in boundary lubrication (λ<1) and
maximum value occurs in hydrodynamic lubrication (λ >10). Bearing designers aim
to achieve lambda ratio close to 10 for hydrodynamic bearings. This ensures
adequate separation of the two surfaces, and thus, minimum frictional losses.
In this experimental study a dichotomy has arisen in specifying the micro-geometry
tolerance values. The variation in bearing ID along the circumference raised doubts,
if the surfaces in a lubricated contact will be adequately separated. While a
roughness based lambda ratio is commonly used, it seems that an out-of-roundness
based ratio would be more appropriate.
3.7.4.2 Film Shape Factor (FSF)
If out-of-roundness value of the bearing element is higher than the roughness values,
the chosen value of lambda ratio does not ensure the separation of the surfaces as
assumed. Therefore, a new parameter –called Film Shape Factor (FSF), or gamma (γ)
ratio– has been proposed, to make the bearing micro-geometry measurements more
robust.
Before elaborating on the FSF, it is important to understand the reason for choosing
the lambda ratio equal to or greater than 10, for hydrodynamic bearings. This is done
to ensure that the surfaces are separated by at least 10 times the composite roughness
value. However, this fails to ensure adequate separation if the out-of-roundness value
is higher than the roughness value.
This study shows that the ‘localised radius’ changes along the circumference of a
bearing. The change in localised radius may be due to two reasons: firstly, due to the
local roughness excursions, and secondly due to the local out-of roundness –called
79
localised roughness (σi) and localised out-of-roundness (Ori) respectively.
The current method of defining out-of-roundness is based on the ASME Y14.5 M-
1994 standard (Cho and Tu, 2001). This method specifies circularity tolerance based
on two extreme circle boundaries to confine the highest peak and lowest valley of a
roundness profile, as shown in Figure 3.12.
Figure 3.12 Film thickness based on composite out-of-roundness concept
Cho and Tu (2001) also analysed the profile variations within these two boundaries,
which are otherwise not being accounted far. In this research, these profile variations
have been called ‘localised out-of-roundness (Ori)’, which result in change of radius
from one location to another, as explained earlier.
In Figure 3.12 bearing and shaft sleeve have been referred by surfaces 1 and 2
respectively. The geometry of these two surfaces has been shown by two types of
circles i.e. outer extreme circles and inner extreme circles; where the mean circles
represent mean radii, similar to the mean lines for surface roughness. The
exaggerated out-of-roundness shape is drawn in between these two circles for both
the bearing elements with subscripts 1 and 2 respectively.
Figure 3.12 shows that areas where roughness values are higher than the out-of-
Extreme outer
/inner circle 1
σ2<Or2
Out-of-roundness profile 1
Out-of-roundness profile 1
Extreme inner/ outer circle 2
σ1>Or1
Mean circle 2
Mean circle 1
Film thickness
80
roundness values (i.e. σ2> Or2), keeping lambda ratio = 10 ensures adequate
separation of the two surfaces. However, when out-of-roundness dominates the
surface roughness value (i.e. σ2<Or2), even though the lambda ratio is 10 the
separation of two surfaces is not adequate. The figure shows, if out-of-roundness
value is too high, then metal to metal contact can occur even with lambda = 10.
The gamma ratio (γ), or the FSF, based on the out-of-roundness values is defined as
in Equation 3.2.
( )22
21 OrOr
h
+=γ (3.2)
Where γ is the film shape factor (FSF), Or1 and Or2 the out-of-roundness of
two surfaces, and h is the film thickness.
Figure 3.13 shows simplified diagram for defining γ ratio, where film thickness h is
the radial distance between the mean radius of bearing and mean radius of the shaft
sleeve.
Figure 3.13 Film thickness based on out-of-roundness concept
Inner circle extreme surface 1
Mean circle, 1
Outer extreme circle,1
Outer extreme circle 2
mean circle 2
Film thickness
Inner extreme circle 2
Inner extreme circle 1
81
Table 3.6 Roughness and out-of-roundness data of test bearings
Tests σb σs σcomp
hhdl = λ (10) x σcomp
Orb Hhdl = γ (10)
x Orb Hhdl / hhdl
A1 1.62 0.525 1.70 17.0 3 30 1.76A2 1.29 0.525 1.39 13.9 4.7 47 3.38A3 0.91 1.1375 1.46 14.6 4.6 46 3.15A4 0.39 0.725 0.82 8.2 14.3 143 17.43A5 1.19 0.625 1.34 13.4 6.8 68 5.07A6 1.36 0.4875 1.44 14.4 3.6 36 2.5A7 1.43 0.4375 1.50 15.0 10.06 100.6 6.7
Notes 1: All film thickness and rms roughness measurement are in microns 2: Shaft sleeve is presumed to be perfectly round (σs=0)
Table 3.6 shows the roughness and out-of-roundness data for the test bearings. In this
table, the values of bearing roughness as well as shaft sleeve roughness have been
converted to rms values (σb) and (σs), and from these values, the composite
roughness (σcomp) was calculated. Multiplying composite roughness by 10 gave the
required oil film thickness (hhdl) value, which is based on lambda ratio = 10.
The measured out-of-roundness values (Orb) for all the bearing elements are higher
than those for the respective composite roughness values (σcomp ). The Hhdl film
thickness values are calculated as 10 times the out-of-roundness values (Orb),
ensuring that the surfaces are adequately separated. To make the process simple the
shaft sleeves are presumed to be perfectly circular i.e. σs = 0.
The last column shows that Hhdl (the film thickness calculated from the out-of-
roundness) is a significant multiple of hhdl (the film thickness calculated from
roughness). Thus, if a bearing was designed for lambda ratio (λ) = 10, then the
resultant film thickness could be 17 times smaller than that required, as shown for
Test A4. Thus, a bearing designed with gamma (γ) = 10 is much safer than that
designed for lambda (λ) =10. This effect is exaggerated in dusty environments, and
therefore, the need for considering this newly proposed gamma ratio is even more
important for such applications.
The application of Film Shape Factor (or gamma ratio) can also be applied to flat
surface hydrodynamic thrust bearings, where surface waviness is higher than the
82
roughness values. In these bearings, waviness of the flat surfaces at a cross section
can be treated as out-of-roundness.
3.7.5 Bearing Component Roughness
Surface roughness is not a direct measurement of wear but is good indicator of the
mode of wear. The pattern of wear and surface topography can reveal useful
information when the effect of surface finish on lubrication is considered. The
surface roughness of bearing elements changes with operation. The roughness and
their orientation have a direct influence on bearing performance, because it affects
the oil film thickness in the contact zone. Studies by Tonder [1986], Patir and Cheng
[1978], revealed that roughness in the transverse or axial direction helps to increase
the minimum oil film thickness. Since thicker oil film indicates the higher load
carrying capacity, increase in the roughness value in transverse direction is useful in
carrying the higher loads for the same bearing design parameters. Similarly there is
evidence to show that circumferential roughness promotes the flow of fluid, thus
reducing the film thickness. However, it has been found that the frictional losses in
such bearings are lower compared to the transverse roughness case. Thus the average
roughness values of the bearing and shaft sleeve surfaces were measured in the
circumferential and transverse directions before and after the tests. The
measurements were taken on a Taylor Hobson’s Surtronic 3+ profilometer. The
equipment recorded several other roughness related parameters. The results obtained
will be discussed in the following chapter. The major problem associated with these
measurements was that the roughness varied drastically from one location to the
other within the contact zone. The areas of interest for roughness measurements were
identified either visually or through a microscope. The reported surface roughness
values are an average of three or more measurements within the wear zone.
3.7.6 Maximum Wear Depth
Every test performed in the presence of solid contaminants demonstrated that there is
discernable wear in the bearing and there was a shaft imprint on the bearing at the
contact zone. This wear area was visible with the naked eye. The maximum depth of
wear varied from one test condition to another depending upon the antiwear additive.
Direct measurement of the depth of these worn patterns was not possible and so
83
the out of roundness profiles for each bearing were obtained. Wear depth was
determined by recording the change in geometry. Although this is not a standard
method used by researchers for wear measurements, a comparison with other
measured wear parameters showed that the method is as reliable as many others.
3.7.7 Particle Counts in Oil Sample
Wear debris analysis and particle counting are well known condition monitoring
techniques. The solid contaminants ingress in the lubricating oils is unavoidable.
Starting from the stage of manufacturing, packaging and handling and usage in the
machines the particles from the environment or generated within the bearing harm
the bearing. The size of these particles may vary depending upon the sources and
this may change due to entrapment within the bearing surfaces and by further
crushing action. A Quant-Alert system shown in Figure 3.14 was used to measure the
number of particles of different size ranges present in the oil sample
The Quant Alert measures the number of particles present in the oil sample of 10 ml.
It works on pressure drop/flow principle. The equipment categorises particles in
eight different groups. The size groups selected for this study were: particles >5, >10,
>15, >20, >40, > 50, >75 and >100 micron. The change in particle number was
recorded by counting the particles at these sizes in the sample of pure base oil, oil
containing 4g/l Al2O3 before the test and then after the test. The number of particles
generated within the bearing depends upon the influence of the respective additives,
the change in motion of the particles within the contact and crushing of the particles
under the influence of each antiwear additive.
3.7.8 Minimum Oil Film Thickness Measurement
An optimum oil film thickness in the bearing contact is the key to successful bearing
performance. Film thickness is a very sensitive parameter especially when the
bearing is operating with small minimum oil film thickness and the accuracy in
measurements is of prime importance. An error of a few microns in calculations or
measurements can cause the bearing to run in boundary or mixed lubrication regime.
There are several conventional methods for measurement of oil film thickness such
as; sensors/transducers based on capacitance or eddy current, X-Ray, shock pulse,
84
optical and voltage discharge methods. The influence of wear due to contaminants
treated with antiwear additives resulted in change of micro-geometry of the bearing
which reduced the minimum oil film thickness in conjunction with the physical
obstruction of the oil flow at the bearing inlet. This was examined by recording the
change in minimum oil film thickness in the bearing contact at three stages during
the test i.e. at the beginning at the middle and at the end of the test with the help of
eddy current type proximity probes.
Figure 3.14 Quant Alert
3.7.8.1 Film thickness Measurement by proximity probes
Proximity probes were chosen for measuring the film thickness. They work on eddy
current principle where intensity of the current between the tip of the probe and the
target material relates to the distance between them. These are recommended to be
used in pairs at a time mounted at right angle to each other on the bearing housing.
85
These probes are said to be capable of measuring film thickness with an accuracy of
half a micron.
At no load the shaft and bearing are theoretically concentric. The probes measure the
distance between their tip and the nearest surface of the shaft in two perpendicular
directions. The eccentricity in the bearing was measured in two steps. First the
distances between the tip of the probe and the shaft sleeve were measured when
operating and theoretical concentricity. Later the bearing was run with a known load
and speed, and the distances between the tip and the shaft sleeve surface were
measured again. The distances measured by these two probes were treated as
coordinates of an imaginary point in space and measurements at no load and full load
conditions were treated as the change in coordinates of the point in space. The
displacement of the points in space can be calculated by knowing the change in
coordinates of these points. In a bearing, whose shaft is fixed and the bearing is
floating, if (X1, Y1) are the coordinates of the centre measured from the origin of a
fixed reference frame at no load and (X2,Y2 ) are the coordinates of the centre of the
bearing after the load is applied, the bearing eccentricity ‘e’ can be expressed as:
( )212
221 )()( YYXXe −+−= (3.3)
The minimum oil film thickness is the difference between the radial clearance and
the eccentricity, and can be expressed as:
eCh −=min (3.4)
It should be noted that in this experimental set-up the proximity probes were fixed on
the bearing, which is floating with respect to shaft, and hence the reference frame is
not fixed. Thus, the eccentricity cannot be measured directly by using Equation 3.3.
Ideally, at no load the gap between the shaft and bearing surface is supposed to be
equal to radial clearance ‘C’. But in reality the bearing was not concentric due to its
own weight (25 N).
86
Calculate, program constants
Set P(i,j) = 0
Calculate h(i,j) add WD(i)
Solve Reynolds Equation (finite difference form)
Set boundary conditions
Calculate P (i,j)’s & SPN
Set SPO = SPN
Assume ε and Φ
2
Calculate in finite difference
Is SPN -SPO .LT. 0.001
YES
1
3
Continued on next page
Input data
87
Figure 3.15 Flow chart
Calculate WX, ,WY & Φ
Is Wcal < W app.
YES
Print P(i,j), h(i,j), Φ,ψ
2
Calculate ψ
END
Is ψ < φ
YES NO
3
Continued from the previous page
88
Efforts were made to estimate this eccentricity ‘eo’ at no load and corresponding H0
using ESDU charts, but due to small bearing load it could not be determined with
sufficient accuracy. To improve the accuracy, value of eo was determined by
developing a program for this research, based on algorithm used by Pai and
Mazumdar (1992). In this case an error in measuring the eccentricity eo due to
change in viscosity was ignored, because the load is too small (25N). The flow chart
of the program developed in Fortran language is shown in Figure 3.15, and the
Fortran code in Appendix B. The film thickness predicted by this program and the
experimental results were also compared with the standard ESDU software. This
software package is developed by the ESDU, and is used for predicting journal
bearing performance for known input parameters for normal lubricated conditions.
The results of this program, for selected test cases are shown in Appendix C.
The actual load in bearings at so called “no load” condition is not zero because
bearings have their own weight, and hence the minimum oil film thickness hmin at no
load is also not equal to the radial clearance 'C'. The value of the film thickness at no
load condition reduces by an amount 'eo' which is the eccentricity created by the self
weight of the bearing. Thus amount 'eo' needs to be subtracted from all hmin values
measured experimentally. The steps followed for the measurements of film thickness
by proximity probes are given as below:
• Calibrate the probes
• Measure the radial clearance as accurately as possible.
• Calculate theoretically the value of hmin using the Fortran Program (will be
discussed later) for given operating parameters at self-weight of bearing (25
N) and speed 1420rpm (the reference experimental hmin for no load
condition).
• Measure the displacement from both the probes for no load conditions and
treat them like coordinates of an imaginary point in space. Since hmin is the
difference between radial clearance and eccentricity eo, the value needs to be
subtracted from each hmin value measured for operating conditions other than
no load conditions.
89
• Measure the coordinates for operating load and speed condition and calculate
the displacement between the imaginary points and calculate the film
thickness. The actual oil film thickness is smaller by an amount ‘eo’ as
determined in the previous step.
After completing the experiments using the above methodology, it was realised that
the calculations of film thickness were not straight forward. The proximity probes
were mounted on the floating bearing housing. Thus, the reference coordinates at no
load cannot be used as the reference point for measuring the shaft sleeve
displacement. Though the measurement error was not large and would have been
acceptable if measuring thick oil films; further improvement was considered
desirable, as film thickness under study is small. Investigations revealed that a key
phaser device could have been used to accurately map the change in coordinates of
the bearing with change in operating conditions, but that was too late. In order to
avoid repetition of the experiments a trigonometric solution was developed to solve
this problem which is explained in the following sections.
3.7.8.2 Calibration of Proximity Probes
Proximity probes REBAM 300 were chosen for film thickness measurements. These
probes with scale factor 40V/mm were found to be suitable for measuring thin films.
The probes were recalibrated to confirm their efficacy in the actual environment.
Change in output was measured by moving the target material (shaft sleeve) against
the probe tip with controlled displacements in micron steps. The change in mV
output was recorded against the displacement. In experimental set-up the probes
were mounted in a steel housing and tip passing through a hole in the bronze bearing
and there was oil in the gap between the probe tip and shaft sleeve. This required a
fixture where probe passed through a hole in the bronze bearing and the space
between the tip of the probe and shaft sleeve surface was filled with oil.
The mV out put was recorded for 180 micron gap equivalent to diametral clearance
of the bearing which could not be achieved successfully because the shaft sleeve
inside the bearing could not be held perfectly square. In Figure 3.16a the calibration
fixture of a half cut bearing is shown which was used for calibration to achieve
perfect squareness. A full calibration set-up is shown for the proximity probes in
90
Figure 3.16b; the probes were mounted on a stationary frame and a half piece of
bronze bearing was mounted on another stationary plate such that the probe passes
freely through the hole in the bearing.
The shaft sleeve was mounted on a micro-displacement table and was moved in steps
with the micro displacement controller against the probe tip. As shown in the Figure
3.16b the probes are connected to a proximeter with an extension lead of two metre
lengths which gives eddy current output in mV. It requires input power voltage 18-24
V. In this experimental study 24 V input was used throughout to get a stronger output
signal. The probes were calibrated under conditions as close as possible to the actual
experiments. To simulate the performance of the probe in the oil medium, an oil drop
was placed in between the probe tip and the shaft sleeve surface to simulate the probe
tip and the shaft sleeve surface covered with the oil film.
It was noticed that the calibration did not change due to the presence of lubricant.
However the bearing surrounding material affects the output depending upon
whether the bearing material is bronze or steel. It was also observed that the
calibration is linear only when the output voltage is more than 10V as recommended
by the manufacture.
Figure 3.16a Probe calibration fixture
91
Figure 3.16b Calibration setup
Figure 3.17a Calibration chart of probe 1
This output can be achieved with the minimum 3mm gap between the probe tip and
the target. Hence before mounting the probes on the bearing housing for actual
measurements, it was kept in mind that the minimum gap between them is more
Calibration of REBAM 300 Proximity Probe1
0
20
40
60
80
100
120
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15
Output (mV)
disp
alce
men
t (m
icro
ns)
Series1
92
than 3 mm or the output is more than 10V. The calibration graphs are shown in the
Figures 3.17 (a) and (b).
Figure 3.17b Calibration chart of probe -2
3.8 Trigonometric Solution of Film Thickness Measurement
The journal bearing test rig accommodates a fixed bearing and a rotating shaft
sleeve. The probes are mounted on the bearing housing, which floats with respect to
the shaft. Thus, the probes do not measure the displacements from a fixed reference
and hence, either an electronic device called key phaser be used else the Equation 3.3
requires geometrical corrections. A geometrical solution has been developed for
measuring the eccentricity of the bearing from the displacements recorded by two
proximity probes, as explained in the following paragraphs.
In Figure 3.18 two circles with centres S and B represent the shaft sleeve and bearing
with Rs and RB radii respectively – for a bearing operating at full speed and load.
The positions of the proximity probe tips mounted on the bearing housing are 90o
apart; these are represented by arrows ‘X’ and ‘Y’ in the figure. The probe ‘X’ is
located at 900 from the load direction of the bearing. Line of centre (LC) makes an
attitude angle ψ from load direction ‘W’. Bearing eccentricity is the distance between
Calibration graph REBAM 300 probe2
0
20
40
60
80
100
120
1 2 3 4 5 6 7 8 9 10 11 12 13 14
Output (mV)
disp
lace
men
t (m
icro
ns)
Series1
93
the two centres, i.e S and B, and is represented by ‘e’.
The distance measured by probe X is XM, and that measured by probe Y is YN,
denoted as PX and PY respectively. The relationship between ‘e’ and displacements
measured by probes X and Y can be derived as follows:
RB = Bearing radius
Rs = Shaft sleeve radius
PX = Displacement measured with probe X
PY = Displacement measured with probe Y
Thus:
yb PRBN −=
xb PRMB −=
sRSMSN ==
Since probes are at right angle, Δ MNB is a right angle triangle and hence;
22 BNMBMN +=
Or
( ) 22 )( ybxb PRPRMN −+−=
Consider Δ MNB, Δ SMN, and Δ SNB where:
β=∠SNB , θ=∠MNB , φ=∠MNS
Thus: φθβ ∠−∠=∠
and: φθφθφθβ SinSinCosCosCosCos ..)( −=−=
94
Figure 3.18 Geometrical representation of film thickness measurement
In Δ SNB, SN and BN are known. Thus:
MNPR
MNNBCos yb )( −
==θ
MNPR
MNMBSin xb )( −
==θ
Similarly considering Δ SMN;
sRMNCos.2
=φ
and
SNSPSin =φ or
s
s
R
MNRSin
22
2⎟⎠⎞
⎜⎝⎛−
=φ
C
S
X
PY
Px
β
Load
ψ
ø
θ
L
eB
M
N
Y
P
95
By substituting the values of MN and cos β in the following equation the value of ‘e’
can be calculated as follows:
βCosBNSNBNSNe ...222 −+= (3.5)
After substituting all the values:
(3.6)
For calculating the value of hmin , Value of eccentricity ‘e’ can be substituted in
Equation 3.4, which was:
eCh −=min
Thus eccentricity can be calculated (in microns) by substituting the values of shaft
sleeve radius, bearing radius, and displacement recorded by probes X and Y in
equation 3.6. The displacement measured by probes is recorded in mV output which
can be converted to Px and Py respectively in microns by multiplying it by scale
factor of the probe, where 1 mV corresponds to a 25 micron displacement.
3.9 Test Procedure and Experiment Design
The purpose of the test was to run the bearing at a ‘K’ ratio ≤ 1 for a fixed sliding
distance. A total of seven tests were performed. The first test was performed with
base oil alone. After running the test for fixed sliding distance, the change in wear
parameters as well as oil film thickness was recorded. The second test was
performed with the base oil containing 4g/l contaminants. The changes in wear
parameters were recorded for the base oil and base oil mixed with contaminants tests,
and results were compared to find out the amount of wear caused due to
contaminants. The remaining tests were performed with five different antiwear
additives. The film thickness and wear parameters were recorded for each test
condition. Thus the influence of contaminants over the base oil was studied from the
first two sets of experiments and the performance results obtained from the
remaining five tests were used to determine the effect of an individual antiwear
( )( ) ⎥
⎥
⎦
⎤
⎢⎢
⎣
⎡ −+−−
⎪⎭
⎪⎬⎫
⎪⎩
⎪⎨⎧
−+−
−−
⎭⎬⎫
⎩⎨⎧ −
−−−+=s
ybxbs
ybxb
xb
s
ybybsYbS R
PRPRR
PRPRPR
RPR
PRRPRRe.4
)()((.
)()(2).(
).(.222
22
22
96
additive. The wear results were compared to determine the efficacy of each antiwear
additive used in the tests.
Table 3.7 Operating parameters and experiment design
Lubricant/ Additive*
Radial Clearance C (μm)
eo(H0) (μm) No load1
Speed N (rpm)
Load W(N)
Duration t (s)
hmin (μm)
‘K’ Ratio
A1 (pure base oil)
80 2.8(77.2) 400
500 150 15.84 0.99
A2 (Al2O3) 82 3.0(79) 420 500 143 16.08 1.005
A3 89 3.8(85.2) 500 500 120 17.09 1.09
A4 80 2.8(77.2) 400 500 150 15.84 0.99
A5 89 3.8(85.2) 500 500 120 17.09 1.09
A6 89 3.8(85.2) 500 500 120 17.09 1.09
A7 80 2.8(77.2) 400 500 150 15.84 0.99
*The product names have been suppressed by randomised double blind trial testing to maintain the manufacturer’s confidentiality. 1. H0 is the film thickness measured at no load, but including the housing weight
Film thickness was recorded during the above seven test conditions. In most of the
cases the film thickness was recorded in the beginning, middle, and at the end of the
test. Minimum oil film thickness represented the tribological performance of the
bearing and revealed the lubrication status as well as its impact on the bearing life.
Table 3.7 gives an over view of the experiment design, it shows all important
operating and design parameters, such as radial clearance, actual ‘K’ ratio, load,
speed and duration of the tests.
3.10 Conclusion
An extensive literature search provided the required background for designing
experiments to meet the research aims and objectives. The experimental design and
micro-geometry metrological investigations reported in this chapter led to the
following conclusions:
• It was decided that the effect of 5 antiwear additives will be studied
on the wear of journal bearings lubricated with oil containing solid
97
contaminants.
• It was also decided that the effect of change in micro-geometry on the
tribological performance of the bearing will be studied by measuring
the change in minimum oil film thickness as a combined effect of a)
change in micro-geometry due to wear, b) antiwear additives and c)
presence of contaminants in and around the bearing contact.
• Modifications to the exiting test rig were required to conduct the
required experiments.
• It was necessary to design and produce a bronze bearing with shaft
sleeve using the ESDU 84031 method.
• A Fortarn program was required for predicting the minimum oil film
thickness and bearing performance, to validate the experimental
results.
• The need to minimise the test duration and selecting most suitable
method for measuring wear in test bearings resulted in development
of a multi-wear parameter approach (MWPA). Whereby, wear
measurements methods or techniques were categorised in three main
groups: 1) weight loss, 2) change in micro-geometry and 3) change in
particle counts.
• Experimental design comprised of testing seven sets of bearings for
three lubricating conditions: 1) with pure base oil; 2) with oil
containing Al 2O3 contaminants; and 3) with contaminated oil treated
with five different antiwear additives. All test conditions were kept
the same ensuring that the ‘K’ ratio ≈ 1, and the sliding distance =
7536 m.
• The graphical representation of ID and OD data of the bearing and
shaft sleeve respectively indicated that the radial clearance in the
bearing was not constant when measured at different locations on the
circumference. Therefore, it is recommended that the out-of-
98
roundness should always be specified along with the radial clearance
of a bearing.
• It was necessary to develop a geometrical method for measuring the
minimum oil film thickness with higher precision. Proximity probes
mounted on a floating bearing housing required the proposed
geometric correction.
• Designing hydrodynamic bearings with lambda ratio 10 was found to
be inadequate for bearings operating in hydrodynamic lubrication
regime, specially with out-of-roundness value more than the
roughness value, the proposed Film Shape Factor, (FSF or gamma
ratio) can be used as a more reliable design parameter.
• The concept of defining the Film Shape Factor (FSF) can also be
applied to flat surface hydrodynamic bearings, especially when the
waviness values are higher than the surface roughness values; this is
analogous to treating surface waviness similar to out-of-roundness.
99
CHAPTER-4
4. EXPERIMENTAL RESULTS AND ANALYSIS
4.1 Overview
The purpose of the experimental analysis was three fold, 1) to find out the effect of
antiwear additives on the wear of the test bearings, 2) to compare the wear
measurement technique for their accuracy leading to selection of the best technique
to be used for characterisation of anti-wear additives and 3) to examine the effect of
change in micro-geometry on the tribological performance of the test bearings.
The experiments were conducted on seven bronze bearing and shaft sleeve sets. All
tests were performed for a fixed sliding distance 7536 m and fixed “K” ratio of 1. As
stated in the previous chapter, the bearings varied in radial clearance hence for the
same operating conditions the thickness of the oil film formed in the contact zone
would be different. A combination of load and speed was determined such that the
bearings operated at a minimum oil film thickness close to the size of the
contaminants these giving a ‘K’ ratio close to 1. Since all bearings had operated with
a small oil film thickness, the influence of self weight (including housing) of the
bearing had to be taken into account. To cite one of the test examples, a bearing
having radial clearance 80 microns, self weight of 50N and speed 1420 rpm, at no
load, gives eccentricity (eo) of 2.8 micron or correspondingly 77.2 microns minimum
oil film thickness. For a required minimum oil film thickness of 16 microns a hit and
miss method was used to choose a speed that gives desired ‘K’ ratio.
A combination of fixed 500N load and a speed 400 rpm gave a minimum oil film
thickness of 15.89 microns and hence a ‘K’ value equal to 0.99. It was decided to run
the test for a duration that would cause a measurable amount of wear in the bearing.
150 minute duration (equivalent to 7536 m sliding distance) gave adequate wear in
the bearing. The ‘K’ ratio and the sliding distance were kept the same for all seven
tests. The tests were labelled A1 to A7; A1: pure base oil and A2: base oil mixed
100
with 4% (16-micron) Al2O3. The remaining five tests (A3 to A7) were performed
with contaminated oil (4% Al2O3), treated with five different antiwear additives.
Details of different tests (Test A1 to A7) are as follows:
A1- Base oil only
A2- Base oil mixed with contaminants (Al2O3 particles)
A3- Contaminants treated with commercial antiwear additive mixed with oil
A4- Contaminants treated with aryl phosphate group of additive mixed with oil
A5- Contaminants treated with Fuch’s proprietary additive mixed with oil
A6- Contaminants treated with sulphur/phosphorous group of additive mixed with oil
A7- Contaminants treated with isopropyl oleate, fatty acid and isopropyl ester group additive mixed with oil
The other conditions such as; base oil Solvent Neutral 300 with dynamic viscosity
0.042 Pa.s at 40 0C, solid contaminant; Al2O3 of size of 16-μm with concentration
4g/l, inlet temperature 40 0C and oil feed pressure of one bar were maintained the
same for all the test conditions. The sliding distance (7536 m) was also kept the
same for all the seven tests. However, to maintain the “K” ratio equal to one for
bearings with varied radial clearance the shaft speed requirements were different for
different sets of test bearings and hence the duration of the test to achieve the same
sliding distance changed from one test to another.
There were three main categories of wear tests in this experimental research viz.
weight loss, change in micro-geometry and change in particle count. Different wear
parameters and minimum oil film thickness in the bearing contact zone were
measured before and after the tests to examine the effect of the above additives on
the bearing performance. The need for measurements of different wear parameters
has been explained in the previous chapter through MWPA. These parameters were
recorded at the beginning and end of each test and change in their value gave a
measure of wear. Total 14 wear parameters were recorded for each set of test bearing
using MWPA, and these are listed as grouped below:
1. Weight loss in bearing (ΔWb)
2. Weight loss in shaft sleeve shaft (ΔWs )
101
3. Change in bearing out-of-roundness (ΔORb)
4. Change in bearing radial geometry (ΔODs, ΔIDs, ΔC )
5. Change in bearing and shaft sleeve roughness in transverse as well as in
circumferential directions (Δ Rb, ΔRbt, ΔRs, ΔRst )
6. Change in wear particle counts and change in weight of wear debris
generated (ΔPC, ΔPcg)
7. Maximum wear depth in bearing contact (ΔWD)
8. Change in minimum oil film thickness (Δhmin)
The above parameters were recorded before and after the tests and the change in their
values gave the wear in the test bearings. The initial values or values before the tests
of relevant MWPA parameters are given in Table 4.1.
Table 4.1 Initial measurements before the tests
Tests Contaminants treated with additives
Wear Parameters
Base oil only A1
Al2O3 in Oil A2
A3 A4 A5 A6 A7
Wb (gm) 333.72 333.71 330.14 333.25 341.01 340.22 340.41Ws (gm) 178.52 178.51 178.60 178.80 179.71 181.43 181.30ORb (μm) 3.0 4.7 4.6 14.3 6.8 3.6 10.6 ODs (μm) 39.799 39.801 39.796 39.792 39.847 39.842 39.842IDb (μm) 39.959 39.965 39.969 39.953 40.025 40.020 40.001C (μm) 80.0 82.0 86.0 80.0 89.0 89.0 80.0 Rb (μm) 1.22 0.965 0.69 0.30 0.90 1.03 1.08 Rbt (μm) 1.25 0.94 0.53 0.27 1.03 0.99 0.64 Rs (μm) 0.42 0.42 0.91 0.58 0.50 0.39 0.35 Rst (μm) 0.36 0.3 0.525 0.38 0.47 0.28 0.40 PC*(count/ml) 9 340 227 226 305 255 287 Pcg (g/l) 0.036 5.513 5.38 5.59 5.22 5.8 5.4 Hxps (μm) 17.86 17.24 18.28 14.76 15.62 15.60 14.90 WD (μm) 0.0 0.0 0.0 0.0 0.0 0.0 0.0
*After mixing Al2 O3 powder of 15 micron size in all the oil samples except A1
102
Amongst these parameters the minimum oil film thickness was considered to be the
key bearing performance parameter and was recorded at three intervals during the
test; at the start, midway and at the end of each test. The oil film thickness was
measured using proximity probes and compared for three test conditions i.e. before
adding any contaminants or additives, after adding contaminants and after treatment
with different antiwear additives. The oil film thickness measured before adding the
contaminants represents the performance under standard operating conditions;
measurements after adding the contaminants indicate the adverse effect of
contaminants; measurements after adding an antiwear additive represents the
influence of individual additives on performance. Thus an additive was evaluated in
terms of longevity of the bearing and performance as related to oil film forming
capacity. Changes in oil film thickness were observed during the test by recording
measurements at the start of the test (Hxps), in the middle (Hxpm) and at the end of the
test (Hxpe).
Surface micrographs of some worn bearings and shaft sleeves were also prepared.
These micrographs helped in analysing the wear mode, wear severity and metal
transfer between bearing and shaft sleeve surfaces.
Particle counts were obtained using QuantAlert apparatus to examine the changes in
the number of particles present in the oil. The source of these particles could be
different such as; generated wear debris, added contaminants or broken particles
during the test. The results were obtained by subtracting initial values from the final
values (after the tests values) of different wear parameters obtained from various
measurement techniques. These results, as a change in wear parameters are reported
Table 4.2 and discussed further for their relevance to this research.
The multi-wear parameter method (MWPA) was useful in saving considerable amount of
testing time, and also in comparing the wear measuring methods for their precision as well as
the reliability of results.
The MWPA methodology will be expounded in the following sections.
103
Table 4.2 Experimental results
Test Wear Parameters
A1 (Base oil)
A2 (Al2O3)
A3 A4 A5 A6 A7
ΔWb (g) 0.01 0.25 0.15 0.17 0.15 0.08 -0.09 ΔWs (g) 0.01 0.125 0.03 0.07 0.07 -0.02 0.05 ΔORb (μm) 1.7 31.3 14.3 25.00 16.4 6.6 11.8 ΔODs (μm) -1 15 7 3 4 1 -1 ΔIDb (μm) 2 37 28 39 32 10 16 ΔC (μm) 1 26 17.5 21 18 5.5 7.5 Δ Rb (μm) -0.25 0.2 0 0.63 -0.15 -0.42 -0.48 ΔRbt (μm) -0.290 0.005 0.16 0.41 -0.37 -0.21 -0.26 ΔRs (μm) -0.09 0.11 -0.11 0.13 -0.09 0.05 -0.01 ΔRst (μm) -0.06 0.05 0.08 -0.03 -0.05 0.1 -0.04 ΔPC*
(count/ml) 21 1008 181 318 337 174 257 ΔPcg (g/l) 0.3 1.35 0.72 1.2 0.95 0.3 0.6 ΔWDmax (μm) 1.66 42 29 42 36 12.5 20 Δhmin (μm) 0.0 10.9 4.1 5.8 5.15 3.8 3.8
*Change in 15 micron particle size counts only
4.2 Weight Loss
Weight loss is the directly measured wear parameter. The bearing and sleeve both
were worn during the tests and hence, the weight loss in bearing (ΔWb) and shaft
sleeve (ΔWs) were measured with an electronic balance, which gave an accuracy of
up to three decimal places of a gram. The weight loss results shown in Table 4.2 are
illustrated graphically in Figures 4.1.
Bearing weight loss
-0.2-0.1
00.10.20.3
A1 A2 A3 A4 A5 A6 A7
Tests
Wt.
loss
(mg)
w eight loss
Figure 4.1 Weight loss in bearings
Figure 4.1 shows that, the weight loss in the bearing lubricated with base oil only
(Test A1) was very low and almost insignificant in comparison to others. It would
104
be expected that a bearing operating under full fluid film conditions would not incur
wear except during the starting and stopping. The bearing weight loss was a
maximum at 0.25g for Test A2 which was expected due to presence of solid
contaminants and no antiwear additive.
Figure 4.2a a shows a micrograph of test bearing used in Test A2 after the test. In
this micrograph embedding of particles is clearly visible. The weight loss and
embedding occurring simultaneously in the bearing indicate that the rate of wear in
the bearing might have been changed after the embedding.
Figure 4.2a After Test A2 bearing surface (X50)
The effect of additive used in test A7 showed a weight gain of 0.09g. This weight
gain may be attributed to embedded particles in the bearing surface, as evident from
micrographs of the test bearing A7 shown in Figure 4.2a and 4.2b. Figure 4.2b shows
Al2O3, and steel particles embedded in the bearing surface. As evident from the bar
chart, this is the only test where weight gain was recorded. Minimum weight loss of
0.08 g was observed in Test A6. The weight loss results of Test A6 show that the
additive with Sulphur and Phosphorus (S and P) technology was more effective than
others.
105
Figure 4.2b Micrograph of bearing surface after Test A7 (X100)
Sleeve weight loss
-0.050
0.050.1
0.15
A1 A2 A3 A4 A5 A6 A7
Tests
wei
ght l
oss
(mg)
w eight loss
Figure 4.3 Weight loss in shaft sleeves
The weight loss in shaft sleeves is shown in Figure 4.3. The loss varied from 0.01 g
in the case of pure base oil (A1) without any contaminant to maximum 0.12 g in the
case of Al2O3 (A2). In Test A7 the ester based additive caused weight loss of the
shaft sleeve of 0.05g. Minimum shaft sleeve wear of 0.03g was observed in Test
106
A3. Additive A4 and A5 showed similar antiwear performance with weight loss of
0.07g and proved to be less effective in the case of shaft sleeve wear. Shaft sleeve
wear loss in Test A1 is too small and seems to be an observational error, whereas in
case of Test A6 the trend shows the weight gain of 0.02g which is very close to
experimental error. It concludes that the wear in this case was minimal and S-P
technology based additive proved to be the best performer. It may also be possible
that the shaft sleeve in Test A7 was worn more compared to other tests because the
particles embedded in the bearing acted as cutting tools for the matching shaft sleeve.
Micrographs (figures 4.2a and 4.2b) clearly show an evidence of embedding of small
wear particles on the bearing surface which excludes the possibility of cutting action.
Similar micrographs have also been obtained (Appendix D) for selected test
bearings. Micrographs of other tests bearings indicate that there was a severe
abrasive wear and some signs of cutting wear in case of Test A2 and Test A6.
Cutting action can also be seen in sleeves of Test A5 and A6 and bearings of Test A5
and A6.
The weight gain in Test A7 needs more support to prove the claim and hence the
surface micrographs of these baring surfaces were examined as shown in Figures
4.2a and 4.2b and were compared with the micrographs of other test bearings. The
Other micrographs clearly show the marks of abrasive wear –indicating weight loss,
and also the signs of metal transfer due to embedding of particles –leading to weight
gain. However, none of these micrographs give any evidence that weight gain due to
embedding of particles (weight gain) is more than the weight loss due to wear. Hence
there is a need for further support to claim the weight gain in the test bearing.
4.3. Out-of-roundness
Out-of-roundness (OR) was considered to be a sensitive micro-geometry parameter
and a superior wear indicator than any other wear parameter. To save the time, it was
decided that the out-of-roundness will be measured for test bearings only. The shaft
sleeves are rotating elements of the test bearings, and hence the change in their out-
of-roundness due to wear is not significant as compared to bearings.
The results of change in out-of-roundness (ΔORb) as recorded in Table 4.2 have been
107
represented graphically in Figure 4.4. To illustrate the process of determining the
change in out-of-roundness (ΔORb) various graphs obtained from the Talyrond for
Test A2 and A6 are shown in Figures 4.5a to 4.5f. The out-of-roundness of the test
bearings was recorded at three fixed locations: at the top end, middle and bottom end
of the test bearing. The out-of-roundness of a bearing is the average of these three
readings.
Change in out-of-roundness of bearing
010203040
A1 A2 A3 A4 A5 A6 A7
Tests
Cha
nge
in o
ut o
f ro
undn
ess
(mic
rons
)
out-of-roundness
Figure 4.4 Change in out-of-roundness of bearings
It should be noted that the shape of the out-of-roundness graph is not the true
representation of the bearing geometry. The graph represents the amplified
excursions from the nominal circumference. In other words two points diagonally
opposite to each other on the out-of-roundness trace do not represent the diameter of
the circle. It depends upon choosing the position of the trace point on the graph paper
from the centre of the chart. The shape of out-of-roundness graph at one cross-
section may look entirely different by choosing different tracer positions and
magnifications. Magnification of the graph can be adjusted and it also depends upon
the type of transducer arm length.
The actual out-of–roundness is calculated with the help of a monograph. One should
not conclude from these traces that their out-of-roundness has changed with change
in magnification or tracer position. However, the foot print of the wear area in the
bearing can be clearly identified from the graph as the excursions in the region of the
108
wear. These excursions are in general much larger than the normal out-of-roundness
trace of the rest of the periphery. Similarly position of a transducer can also alter the
shape of the out-of-roundness graph. Normally wear in a bearing appears concave in
the trace. A convex or inverted out-of roundness graph of a bearing was also
obtained (Figure 4.5h).
The reverse trace can be obtained by mounting the transducer in the reverse
direction. To demonstrate the procedure adopted for measuring the out-of-roundness
and the results obtained for the best and the worst case of wear have been discussed
through Talyrond graphs shown in Figures 4.5a to Figure 4.5h. The worst case of
wear is Test A2 where contaminants were used without any additive and the value of
out-of-roundness was 31 microns. The best results were obtained with S-P additive
used in Test A6, where out-of-roundness changed by 6.6 microns only. The change
in out-of-roundness of the bearing used in Test A4 was close to Test A2 with a
change of 25 microns and hence the additive performance was of the lowest order in
comparison to other additives.
Minimum wear occurred in Test A1, where base oil was used and no contaminant or
additive was present in the oil. As expected, there was hardly any change in
roundness of the test bearing lubricated with base oil only. Additives, A3 and A5
results were comparable at 14.3 microns and 16.54 microns respectively.
Performance of ester based additive used in A7 resulted in a change of 11.8 microns,
which is almost two times higher than the Test A6 bearing –10.2 microns.
The micrographs showed that embedding as well as normal wear took place
simultaneously in almost all the test bearings and hence the weight gain in Test A7
cannot be justified with the higher value of change in out-of–roundness. Thus the
out-of-roundness test results do not support the weight gain in Test A7, and hence the
weight gain appears to be an experimental error.
109
Figure 4.5a Bottom end out-of-roundness before Test A2
Figure 4.5b Bottom end out-of-roundness (convex graph) after Test A2
110
Figure 4.5c Top end out-of-roundness before Test A3
Figure 4.5d Middle position out-of-roundness before Test A6
111
Figure 4.5e Bottom end out-of-roundness before Test A6
Figure 4.5f Top end out-of-roundness of bearing after Test A6
112
Figure 4.5g Middle position out-of-roundness of bearing after Test A6
Figure 4.5h Bottom end out-of-roundness Test A6
113
4.4 Radial Clearance
In the existing journal bearing test rig a 40 mm nominal bore journal bearing with
radial clearance of 90 microns was used. As discussed in the previous chapter, to
avoid the misalignment of the shaft the radial clearance (C) was chosen in the higher
range. However due to manufacturing limitations the bearings with exact dimensions
cannot be produced and hence the radial clearance varied from one test bearing to
another. Keeping in view this fact as well as the constraints associated with the
metrology of measurements, the problem was further investigated.
Figure 4.6 Shaft sleeve trace inside the bearing out-of-roundness trace
Shown in Figure 4.6 the out-of-roundness trace of shaft sleeve and bearing were
recorded on the same graph paper with the same magnification but tracer position
slightly shifted from one another. It should be noted that individually the out-of-
roundness of these elements is less than 2 microns which is well within the
acceptable limits. However, the graph clearly shows that the clearance between the
two surfaces is not uniform through their periphery when the bearing is stationery,
and it varies from one location to another. If this bearing is in motion the transient
values of radial clearance within the bearing may vary with time. This also highlights
114
the problem of assigning a single value of radial clearance apart from the metrology
issues discussed in the previous chapter. The problem exaggerates when out-of-
roundness is larger. Traces in the figure clearly indicate that no matter how precise
the bearing and sleeve samples are, they are never perfect and assigning a single
value of radial clearance in a bearing with a high precision is not as simple as
thought to be.
Radial clearance has a direct influence on the load carrying capacity of bearings; i.e.
the film thickness in the bearing changes significantly with minor change in radial
clearance (Chu and Kay, 19740). The common practice of finding the radial
clearance is to measure the difference in nominal ID of the bearing and the nominal
OD of the shaft. It is presumed that this value of radial clearance holds well in a
running bearing along the line of centres which forms certain attitude angle from the
load line. However, the actual radial clearance which is the transient clearance is
time dependent and may depend upon the out-of-roundness at an instance at a fixed
location. The change in radial clearance (ΔC) was found by recording the radial
clearance before and after the test. The initial values were calculated from the
nominal ID and OD of the bearing and shaft sleeve respectively. The final values
were calculated from the maximum bearing ID measured in the worn surface area
(with hole-test-gauge) and shaft sleeve OD – measured with the HP laser system.
The results of change in radial clearance are shown in Figure 4.7. The figure shows
that the change in Test A6 was minimal with a value of 5.5 microns and was close to
the performance of additive used in Test A7 which had a an increase of 7.5 microns.
The performances of additives used in tests A3, A4 and A5 were of the same order of
magnitude with change in values 17.5 microns, 21microns and 18 microns
respectively. As anticipated with pure base oil in Test A1 a negligible change of 1
micron was recorded. In case of Test A2 with contaminants in absence of additive
treatment was recorded to the highest value of 26 microns. It may be noticed that in
Test A2 the change in radial clearance was not too high in comparison to Test A4
which performed poorly, and the change is only 21 microns. This may be due to
embedding of wear particles in the bearing surface of Test A2, which might have
reduced the successive rate of wear in the bearing.
115
Change in radial clearance
0
10
20
30
A1 A2 A3 A4 A5 A6 A7
Tests
Cha
nge
in ra
dial
cl
eran
ce (m
icro
ns)
radial clerance
Figure 4.7 Change in radial clearance of bearings
The results also show that the change in radial clearance varied from one test bearing
to another depending upon the oil condition and the influence of their respective
antiwear additives. The change in radial clearance may occur due to dimensional
change in either bearing ID or shaft sleeve OD or both. In case of Test A5 the
increase in bearing ID was recorded 32 microns and a reduction of 4 microns in the
shaft sleeve ID, resulting in 18 micron change in radial clearance. Whereas, in Test
A2 bearing ID increased by 37 microns and the shaft sleeve OD reduced by 15
microns, correspondingly the change in radial clearance was 26 microns. Thus, in
this test, the contribution of change in bearing ID is significant in comparison to
other tests. It is also important to note that an additive influences the micro-geometry
of the bearing as well as of the shaft sleeve and its degree of influence differs from
one another.
Though, it is hard to determine the effect of an additive individually on both the
bearing elements, the data collected is represented graphically in Figure 4.8. It shows
the changes in ID and OD of test bearings and the shaft sleeve respectively. The
graph shows the absolute change in diameter of the bearing and the shaft sleeve,
where change in bearing ID means increase in its diameter and change in shaft sleeve
diameter corresponds to reduction in OD.
116
Figure 4.8 shows that the bearings were worn more than the shaft sleeves. It also
shows that in Test A2 the shaft sleeve and the bearing both were worn significantly
and even in the absence of an antiwear additive the bearing was not worn as severely
as in case of Test A4, where antiwear additive was supposed to be effective. This
indicates that the metal transfer in the Test A2 bearing could have been effective at
the early stage, which retarded the wear rate after embedding of the wear particles in
the bearing surface. Phosphorus based additive used in Test A4 shows an increase in
bearing ID (39 microns) higher than Test A2, where no additive was used. The
reason for higher bearing wear in Test A4 could be opposite to Test A2, where
cutting action from the particles embedded on the shaft sleeve resulted in gauging of
the bearing.
Change in bearing geometry
-100
1020304050
A1 A2 A3 A4 A5 A6 A7
Tets
Cha
nge
in d
iam
eter
(m
icro
ns)
reduction in sleeve OD increase in bearing ID
Figure 4.8 Changes in bearing element geometry
This can be supported with the weight loss results of the shaft sleeve in Test A2,
which was higher (0.125 g) as compared to 0.07 g in Test A4. Unlike bearing, the
change in shaft sleeve OD of Test A4 was lower (by 4 microns) in comparison to 15
microns in Test A2. In Test A3, wear in bearing as well as in the shaft was higher in
comparison to Test A6 and A7 which is unexplained. In case of Test A6, the
117
additive was most effective with the least change in ID and OD of the respective
bearing elements i.e. 1 micron and 10 microns respectively. A change of 1 micron in
shaft sleeve OD of Test A7 is so small that it can be treated as a variation in
measurement. Moreover, negative change or no change in shaft sleeve OD proves
that the weight gain in Test A7 bearing is not possible.
4.5 Change in Roughness
The literature survey revealed that the roughness of bearing elements affects the
formation of minimum oil film thickness at the contact zone. Minimum oil film
thickness in the bearing contact is the best indicator of its load carrying capacity. It is
common to think that ‘smooth is best’ but the orientation of roughness in a preferred
direction can help to develop thicker films. Transverse roughness helps to increase
the oil film thickness at the contact whereas rougher surfaces in the circumferential
direction increase the fluid flow through the contact and hence give thinner films, but
an easy passage for contaminants flow. Keeping in view this phenomenon,
roughness of bearings and shaft sleeves were recorded in both orientations. The
roughness was measured in both directions before and after the tests and respective
changes were recorded. The change was determined by subtracting the initial
roughness value from the final value (after the test). The negative change in surface
roughness means smoothening effect after the test and positive change shows the
opposite.
A number of surface topography parameters were used including Ra (arithmetic mean
height) Rq, (root mean square roughness) Rt, (maximum valley to peak excursion),
Bearing ratio (i.e. the percentage of the assessment at a level below the highest peak)
was also measured. The cut-off length of 0.8 mm was chosen for all the surface
roughness measurements and the surface roughness were measured as average
roughness or Ra values. The roughness of the bearing and sleeve surface varied from
one location to another and so the average values were recorded. Surtronic-3
instrument was used for the measurements, which averaged 5 scanned cut-off lengths
of 0.8 mm each (total 4 mm linear stroke) and gave Ra values in microns.
118
Roughness measurements were taken at different locations on the bearing’s inner
surface and on the shaft sleeve’s outer surface –before the tests by choosing locations
randomly. It was often found difficult to select the location as a true representative of
the worn surface. Because, in test samples, some times the roughness within the
worn area varied drastically from one location to another. It is common to find a burr
or scuff mark in the worn area which is lot rougher than the normal worn area. Thus
abnormal roughness values did not make sense, and hence discarded to make a useful
analysis of the results. The measurement locations were chosen by visual judgment,
some times using a magnifying glass and the average of minimum three
measurements was considered to be the representative roughness of the bearing
elements in the worn area. It was not possible to record traces for all the
measurements, these were recorded randomly and hence Ra values of the traces
included in figures do not match with the reported average Ra values in the tables 4.1
and 4.2 and so in the bar charts. The roughness values recorded in the roughness
traces are as close to the reported average values as possible any value matching with
the average value is a coincidence.
4.5.1 Bearing roughness
Although the bearings were produced in lots by using the same manufacturing
process, their roughness varied from one test specimen to another. Table 4.1 shows
the initial roughness values of the entire lot of test specimen before the test. The
change in roughness value (Ra) was recorded by subtracting the before test
roughness value from the after test value. The negative values indicate that the
surface has smoothened. The changes in surface roughness of all the test bearings in
circumferential and transverse direction are presented in Table 4.2 and shown
graphically in Figure 4.9 (a) and (b).
Figure 4.9a shows that the bearing surface becomes smoother after operating with
base oil only (Test A1). This indicates that the start and stop conditions were helpful
for the running-in process in the bearings.
119
Change in bearing roughness
-1
-0.5
0
0.5
A1 A2 A3 A4 A5 A6 A7
Tests
Cha
nge
in R
a va
lue
(mic
rons
)roughness
Figure 4.9a Change in bearing circumferential roughness
The change in the bearing Ra value -0.25 microns states that the surface has
smoothened from 1.22 microns to 0.97 microns. However in Test A2 with oil
containing 4g/litre of Al2O3, the roughness increased by 0.2 micron i.e. from 0.97
microns to 1.17 microns. In Test A4, the phosphate ester based additive gave the
worst results, where Ra value increased maximum from 0.27 to 0.4 microns in
comparison to all other cases including Test A2 where no additive was used. The
reason for increase in roughness is not well understood, however, it is anticipated
that the initial roughness of this bearing was very low (0.3 microns) and results in
Table 4.1 and 4.2 show that there is a trend that the bearings with lower initial
roughness were worn more. It is obvious that Test A4 additive had shown even
worse performance than the Test A2. The increase in roughness of test bearing A4 is
three times (0.63) higher than the Test A2 (0.2 microns Ra value). Test A7 showed
the best performance as the smoothening effect is maximum (-0.48 microns). Thus
Test A7 showed better smoothening than Test A6 which gave -0.42 micron Ra value.
120
.
Change in bearing transverse roughness
-0.4-0.2
00.20.4
A1 A2 A3 A4 A5 A6 A7
Tests
Cha
nge
in b
earin
g tr
ansv
erse
Ra
(mic
rons
)
roughness
Figure 4.9b Change in bearing transverse roughness
Change in transverse roughness of the bearings is shown in Figure 4.9(b). In this case
also the results are some what similar to that of circumferential roughness. In this
case Test A2 bearing surface was even smoother in transverse direction than
circumferential direction that is from 0.2 to 0.005 micron Ra value. Test A4 results
again showed that the additive was not effective in reducing the roughness in either
direction. The influence of phosphate ester additive in Test A4 was adverse,
transverse surface roughness of the bearing changed from 0.27 to 0.68 micron Ra
value which is 150% higher than the initial Ra value, and these values are very high
as compared to Test A2. Test A6 did not show the best smoothening effect as
compare to other antiwear additives, but the additive was effective in reducing the
roughness in both the directions. Another surprising roughness result was obtained
for Test A3 where roughness has increased from 0.53 to 0.16. This shows that the
additive was more effective in reducing the roughness in circumferential direction
than the transverse direction and this additive was also not as effective in reducing
the roughness as pure contaminants in absence of any additive (Test A2). Test A1
results were consistent and showed slightly better smoothening effect. The
commercial additive used in Test A5 has maximum smoothening effect, reducing Ra
value by 36% from 1.03 to 0.66 microns. In case of transverse roughness
performance of additives used in Tests A6 and A7, their performance was close to
each other with change in Ra values to -0.21 and -0.26 respectively –approximately
121
40% smoothening effect.
4.5.2 Shaft sleeve roughness
Changes in shaft sleeve circumferential and transverse roughness values are shown in
Figures 4.10a and 4.10b respectively.
Change in sleeve roughness
-0.15-0.1
-0.050
0.050.1
0.15
A1 A2 A3 A4 A5 A6 A7T est s
Cha
nge
in
roug
hnes
s (m
icro
ns)
roughness
Figure 4.10a Change in shaft sleeve circumferential roughness
In Test A1 (for base oil only) the circumferential roughness (Ra) of the shaft sleeve
changed from 0.42 to 0.33 microns, which is smoother by 0.09 microns. The
smoothening effect can also be seen in Tests A3, A5 and A7, with reduction in Ra
values to 0.11, 0.09 and 0.01 microns respectively. In case of Test A3 the bearing
was rougher than sleeves after the test. The shaft sleeve was smoother in transverse
direction; the effect was very prominent in comparison to other tests as change was -
0.0.11 micron Ra value. In Tests A5 again the shaft sleeve was highly smooth by -0.9
micron Ra value which of the same order as Test A5. In Test A7 the effect on the
shaft sleeve was smoothening but lowers than the bearing roughness this was close to
.0.01 micron which is with the measurement error limit. Results of Test A6 were
unexpected when compared with bearing roughness because roughness has increased
by 0.05 microns Ra value.
The shaft sleeve used in Tests A2, with contaminants without additive showed an
increase in roughness by 26% from its initial value of 0.42 microns Ra value. Test A4
showed poor results in this case also; where roughness of the shaft sleeve
122
increased by 0.13 microns Ra value (22% increase). The most smoothening effect of
about 9% was observed with the commercial additive used in Test A3. In Test A2;
with the base oil contaminants but no additives, increased the shaft sleeve roughness
by the same order of magnitude as for the Test A4 i.e. 0.13 and 0.11 microns
respectively.
Change in sleeve transverse roughness
-0.1-0.05
00.050.1
A1 A2 A3 A4 A5 A6 A7
Tests
Cha
nge
in
tran
seve
rse
Ra
(mic
rons
)
roughness
Figure 4.10 b Change in shaft sleeve transverse roughness
Additives influenced transverse roughness of the shaft sleeve in a random fashion as
shown in Figure 4.10b. Test A1, where only base oil was used showed that the
surfaces were smoother after the test by 0.06 microns Ra value. The shaft sleeve was
rougher in transverse direction by 17% to 0.35 micron Ra value in Test A2. Thus
Test A2 showed that the surfaces were rough in both the elements in either direction.
This effect in transverse direction was an increase of only 0.05 microns Ra value.
The shaft sleeve surface was rougher in transverse direction for Test A3 by 0.08
microns Ra value (15%) contrary to the smoothening effect in the circumferential
direction of about 12%. Though the smoothening effect was very mild, unlike
previous increases in roughness of both the bearing and the sleeve, the shaft sleeve
were smoother in transverse direction was smoother in Test A4 by -0.03 microns Ra
value. Similar to Test A4 the effect of S-P based additive in Test A6 was also
unexpected where roughness of the sleeve increased in the transverse direction by 0.1
micron Ra value (from 0.28 to 0.38 microns Ra value). This effect of phosphate ester
additive is unexplained.
123
The smoothening effect in Test A5 was significant, where shaft sleeve surface was
smoother by approximately 10% and Ra values changed from 0.47 to 0.42 microns.
Smoothening effect in the Test A7 was not as strong as in the case of Test A5, where
Ra value changed from 0.4 to 0.38 microns – a 5% change only.
4.5.3 Individual additive effects
Effects of individual additives were also analysed on each bearing element’s
roughness in both the directions and are represented graphically in Figure 4.11a to
Figure 4.11g. Pure base oil in Test A1 had smoothening effect in all the bearing
elements on both type of roughness (circumferential and transverse). The effect was
more prominent on bearings than shaft sleeves. The maximum smoothening effect
occurred on the bearing in transverse direction. A minimal smoothening effect was
observed on the shaft sleeve in transverse direction.
In Test A2 with oil containing contaminants Al2O3 but no additive, both the elements
were rougher and the influence was more prominent on the transverse roughness in
comparison to circumferential direction. The difference in increase in their values
was not in the same proportion because very small rise in roughness took place in the
transverse direction of the bearing when compared with all other elements. Similarly
the maximum rise in roughness was recorded in the bearing circumferential
roughness. The rise in roughness between the shaft sleeve directions was more
prominent in the circumferential direction than the transverse direction.
Roughness changes in Test A1
-0.4
-0.3
-0.2
-0.1
0RB RBT RS RST
Bearing and sleeve roughness patterns
Cha
nge
in R
a va
lue
(mic
rons
)
roughness
Figure 4.11a Roughness effects after Test A1
124
Contaminants in the absence of antiwear additives are likely to cause more wear than
the cases where antiwear additives are used or no contaminants are used. As shown
in Figure 4.11(b) the maximum effect of about 20% occurred with circumferential
roughness of bearing whereas the effect on transverse roughness of bearing is
negligible.
R o ughness changes in T est A 2
0
0.1
0.2
0.3
RB RBT RS RST
B earing and sleeve roughness pat t erns
roughness
Figure 4.11b Roughness effects after Test A2
The influence on circumferential roughness of the shaft sleeve is almost half the
bearings circumferential roughness whereas in comparison to transverse roughness of
the shaft sleeve about 100% higher. However the overall influence is less than some
of the additives. This may be due to running-in effect or lapping effects during the
test.
Figure 4.11c shows the effect of commercial additive used in Test A3. The pattern of
influence was irregular. Though effect on the bearing was more prominent than the
shaft sleeve, no regular pattern could be established.
125
Roughness changes in Test A3
-0.2-0.1
00.10.2
RB RBT RS RST
Bearing and sleeve roughness patterns
Cha
nge
in R
a va
lue
(mic
rons
)roughness
Figure 4.11c Roughness effects after Test A3
The influence of the additive was more favourable in case of circumferential
direction when compared with the transverse roughness and as in previous cases
increase in roughness was more in case of bearings than the shaft sleeves. The rise in
transverse roughness was higher for both the elements as compared to
circumferential roughness.
In Figure 4.11d influence of additive use in Test A4 shows that the rise in roughness
was higher in the bearing as compared to shaft sleeves. Similarly in both the
elements effect on the circumferential direction was more prominent as compared to
transverse direction. Though negligible, the only smoothening effect was observed
on the shaft sleeve in transverse direction which was about 7%.
126
Roughness change in Test A4
-0.20
0.20.40.60.8
RB RBT RS RST
Bearning and sleeve roughness patterns
Cha
nge
in R
a va
lue
(mic
rons
)
roughness
Figure 4.11d Roughness effects after Test A4
Shown in Figure 4.11 (e) the commercial additive used in Test A5 improved the
roughness in both the elements in either direction. Bearing transverse roughness was
maximum smoothened maximum by 35%, whereas the influence on the shaft sleeve
transverse roughness in the transverse direction was approximately 10%.
Roughness changes in Test A5
-0.4-0.3-0.2-0.1
0RB RBT RS RST
Bearing and sleeve roughness patternsCha
nge
in R
a va
lue
(mic
rons
)
roughness
Figure 4.11e Roughness effects after Test A5
As shown in figure 4.11f the influence of the phosphorous-sulphur based additive in
Test A6 was quite significant in all cases except transverse roughness of the shaft
sleeve where surface rougher by 3.5%. The influence on circumferential roughness
of the bearing was most dominant with a reduction of roughness by approximately
127
40%. Similarly its effect on shaft sleeve smoothening was observed in
circumferential direction whereas there was arise in roughness value in the transverse
direction. The effect was almost 100% when compared between circumferential and
transverse direction.
Roughness changes in Test A6
-0.6-0.4-0.2
00.2
RB RBT RS RST
Bearing and sleeve roughness pattern
Cha
nge
in R
a va
lue(
mic
rons
0
roughness
Figure 4.11f Roughness effects after Test A6
The isopropyl oleate additive used in Test A7 proved to be the dominant antiwear
additive influencing the circumferential roughness of the bearing (maximum 44%).
Figure 4.11g shows the minimum smoothening effect observed was the
circumferential roughness of the shaft sleeve (a negligible by 2.8%).
Roughness changes in Test A7
-0.6
-0.4
-0.2
0RB RBT RS RST
Bearing and sleeve roughness pattern
Cha
nge
in R
a va
lue
(mic
rons
)
roughness
Figure 4.11g Roughness effects after Test A7
128
4.5.4 Roughness traces of bearing elements
To cite a few examples, the traces of roughness measurements for Test A5 have been
shown in Figure 4.12(a) to (h).
Figure 4.12a Bearing circumferential roughness before Test A5
Figure 4.12b Bearing circumferential roughness after Test A5
Shown in Figure 4.12a and Figure 4.12b the traces of surface roughness where Ra
values shown are 0.9 micron before the test and 0.71 microns after the test. As
explained earlier, these are the traces acquired randomly, and hence different but
close to the reported average values (0.9and 0.75 microns respectively) as shown in
129
the Table 4.1 and 4.2. In this case the roughness has improved from 0.9 to 0.75
microns Ra value, which is a smoothening effect of -0.15 microns Ra value. These
traces show that the Ra values are very close to the average of several values
measured in the worn area. It could be observed from the test that the peaks have
been knocked off after the test.
Figure 4.12c Bearing transverse roughness before Test A5
The peak to valley height Rt has increased after the test from 5.5 to 9.8 microns. The
only cause of rise in Rt value can be attributed to the cutting action by the edge of the
solid contaminants.
Figure 4.12d Bearing transverse roughness after Test A5
The change in transverse roughness of the bearing is evident from Figures 4.12c and
4.12d, where the Ra value has decreased from 1.03 microns to 0.66 microns. For
130
example, Ra = 0.68 microns in the sample trace shown in figure 4.12d, but the Rt
value remained unchanged at 6.3 microns. Nonetheless, the smoothening of the
surface is obvious due to the additive used in this test.
Figure 4.12e and Figure 4.12f show that the Ra value of the shaft sleeve in the
circumferential direction has improved from 0.52 to 0.44 microns, but the maximum
peak to valley height (Rt) has only insignificant increase i.e. from 6.7 to 6.9 microns
only.
Figure 4.12e Shaft sleeve roughness before Test A5
There is no standard explanation for cutting action by the edge of the solid
contaminants; however the Ra values vary, being statistical numbers that are close to
each other, making it difficult to highlight the effect of one parameter over the other
with existing equipment. Moreover, the roughness parameters vary drastically within
the wear zone where measurement cut-off lengths were chosen at different locations
randomly. In case of transverse roughness of the shaft sleeve the Ra value appears to
decreased from 0.47 to 0.42 (0.40 in the trace) as shown in Figure 4.12g and Figure
4.12h. The change in Rt value has occurred from 4.5 to 3.4 microns Ra value, which
indicates that the cutting action does not dominate the surface roughness change.
Similar traces can be referred in Appendix-F for selective test conditions. These
traces show that the roughness often varies significantly within the wear zone. It is
difficult to represent the surface finish by a single trace or a single roughness
131
value. Therefore, the average of a significant number of readings must be used for
characterising the surface. The results analysis shows that it is hard to explain the
phenomenon behind the change in Rt value after the tests.
Figure 4.12f Shaft sleeve roughness after Test A5
Figure 4.12 g Shaft sleeve transverse roughness before TestA5
132
Figure 4.12 h Shaft sleeve transverse roughness after Test A5
4. 6 Particle Counts (PC)
Particle counts technique for used oils is a well known tool for monitoring the wear
in machines. Lubricating oils contain solid particles even when they are fresh
supplied. The sources of these particles in fresh oil could be several, starting from the
environment in which they are manufactured, handled or stored. Whereas in used oil
the main source is the wear debris generated during the operation in addition to
particles that oil inherit from different sources prior to actual use.
The morphology of wear debris changes with time as particles may break into
smaller sizes and new particles are generated due to wear and so the actual particle
counts change with the usage of oil. Particle counts were measured with the
QuantAlert system for the fresh oil and after mixing the contaminants (Al2O3).
The samples were collected following the standard collection method –directly
tapped from the oil line. The oil was thoroughly mixed before and soon after the
tests. The counts were measured using eight different channels of different particle
sizes. These sensors gave the counts of particles greater than 5, 10, 15, 20, 25, 40, 75
and 100 microns respectively.
The counts of different sizes recorded by QuantAlert from different channels as well
as the change in the weight of total particles present in the oil are given in Table 4.3.
Initial counts and change in counts of particles larger than 15 micron but smaller
133
than 20 microns were the main focus of this study and are listed in Table 4.1 and 4.2
respectively. Table 4.3 presents the change in particle counts of other sizes for
different tests, and in addition the change in weight of total solid contaminants.
Combining particle count and their weight gives a more reliable comparison for the
overall change in wear particles.
Table 4.3 Rise in particle counts of different sizes
Particle Size (μm) Test
>5 >10 >15 >20 >2 >40 >75 <100 Gravimetric change (mg/l)
A1 141 39 21 7 3 0 1 0 0.3
A2 9430 2627 1008 465 24 53 5 10 1.35 A3 1684 469 181 83 43 10 0 0 0.72
A4 1609 449 318 79 42 9 1 10 1.2
A5 3315 926 337 154 80 17 1 5 0.95
A6 1628 453 174 80 41 9 1 0 0.3
A7 4822 1343 257 238 12 28 25 7 0.6
Initially, particles of different sizes as mentioned in the Table 4.3 were measured
from the fresh oil sample. Subsequently the total weight of all types of solid particles
suspended in this oil sample was measured.
The particle count results showed that the supplied oil was well within the acceptable
cleanliness limits of ISO4406 (1987) standards 16/14/11. Some of the used oil
samples were very dark and opaque and required dilution with hexane up to 8 times.
However the dilution varied with oil condition. Efforts were made to keep the
sampling and testing procedures the same during the measurements. However the
variations in counts were large with poor repeatability in many cases.
To make the analysis meaningful only relevant data was considered. Main emphasis
was given to the counts of particles larger than 15 microns, keeping in view the ‘K’
ratio of the operating bearing which makes it mandatory that the minimum oil film
thickness is of the size of the averaged size of the contaminant particles (16 microns).
134
Change in particle count >15 microns
0
500
1000
1500
A1 A2 A3 A4 A5 A6 A7
Tests
Part
icle
cou
nt
total counts
Figure 4.13 Comparison of change in counts for different tests
The analysis of particle counts of each size for different oil samples could have been
more time consuming but less useful keeping the objectives in view, data from all the
channels was collected to observe the pattern of change in counts. Initial values of
particle counts per millilitre of oil sample have been recorded in Table 4.1 and
change in number of particles of 15 microns or larger but smaller than 20 microns in
one millilitre of oil sample has been recorded in Table 4.2. Table 4.2 also shows the
change in weight of total solid particles suspended in the oil sample in g/l. The
results of particles larger than 15 microns are the main concern in this study and have
been presented graphically in Figure 4.13. The performance of individual additives
in generating particles of different sizes has also been illustrated by their respective
graphs in Figures 4.13 (a) to (g).
Figure 4.13 shows that the change in the count of particles larger than 15 microns in
Test A1 was low and increased by 15 counts (from 9 to 30 per ml). Ideally there
should have been no change but experimental errors and ingestion of contaminants
from the surroundings or generation of wear debris during the starting and stopping
could be the contributing factors. The high count rise of 1008 after Test A2 was
expected due to contamination of oil without any additive. The oil sample initially
contained 340 particles per millilitre (after mixing 4g/l Al2O3) before the test and it
increased to 1348 after the test. The counts after the Test A4 commensurate with the
other wear measure parameters with an increase of 318. However the additive used
in Test A5 resulted in an increase in particle counts of 337 which is high in
135
comparison to Test A4. Performance of additive A3 and A6 was also of the same
order with a rise of 181 and 174 respectively. But the effect of additive A7 was
moderate with a rise of 257. Additive A6 proved to be the best performer. However
the performance of the remaining additives did not give a clear indication of
superiority of one over the other. As discussed the repeatability of the QuantAlert
was poor and after repeating the tests only the data deemed relevant was considered
for the analysis.
Test A1 results for sizes of particles chosen from the QuantAlert are given in Table
4.3 and shown in Figure 4.13(a). It shows the change in different sizes of particles
after the test graphically. Results for 100 microns and 40 microns are absent.
However, >75 micron particles were detected in the sample which may be an
experimental error or an inclusion of particles from some external source. Normally
the number of larger particles present in the sample is small compared with the
smaller size particles but there can be exceptions. The largest change for the 5
microns size was 141 which is too small and indicates that the oil was clean. Ideally
there should be no change in counts under hydrodynamic lubrication condition; this
small rise in counts was expected mainly due to small wear during start and stop
conditions or ingestion from surroundings. The rise in count for other sizes of
particles such as >10,>15, >20 and >25 were small and with in acceptable limits.
Change in particle count Test A1
020406080
100120140160
>5 >10 >15 >20 >25 >40 >75 <100
Particle size
Cha
nge
in p
artic
le c
ount
/ m
g
particle counts
Figure 4.13a Change in counts after Test A1
136
In Test A2 the contaminants in the absence of additive caused a large change in
particle counts which indicates severe wear in the bearing. The change in the number
of particles larger than 15 microns rose to 1008 counts. The particles larger than 40
microns were very few in this case but the number of particles larger than 15 microns
was very large in comparison to other test conditions. The results were anticipated in
presence of solid Al2O3 particles without any kind of antiwear additive treatment.
Change in particle count Test A2
0
20004000
60008000
10000
>5 >10 >15 >20 >25 >40 >75 <100
Particle size
Cha
nge
in p
artic
le
coun
t/ m
g
particle counts
Figure 4.13b Changes in counts after Test A2
Results of other tests shown in Figures 4.13c to 4.13g follow the same trend as test
A1 and A2. The results obtained for all the tests for eight different sizes have been
plotted for a comparison and shown in Figure 4.14. The graph shows the dominance
of one additive over another with superiority of additive A6 and poorest performance
of additive A5. The repeatability of the results is poor because generation of particle
of a particular size is not fixed and also the grinding and crushing of bigger particles
results into smaller particles cannot be predicted. To improve the reliability of the
results the total count of all the eight different sizes of particles were recorded for
each test case and presented in Figure 4.13h.
137
Change in particle count Test A3
0
500
1000
1500
2000
>5 >10 >15 >20 >25 >40 >75 <100
Particle size
Cha
nge
in p
artic
le
coun
t/ m
g
particle counts
Figure 4.13c Changes in counts after Test A3
Change in particle count Test A4
0
500
1000
1500
2000
>5 >10 >15 >20 >25 >40 >75 <100
Particle size
Cha
nge
in p
artic
le
coun
t/ m
g
particle counts
Figure 4.13d Changes in Counts After Test A4
138
Change in particle count Test A5
0
1000
2000
3000
4000
>5 >10 >15 >20 >25 >40 >75 <100
Particle size
Cha
nge
in p
artic
le
coun
t/ m
g
particle counts
Figure 4.13e Changes in counts after Test A5
Change in particle count Test A6
0
500
1000
1500
2000
>5 >10 >15 >20 >25 >40 >75 <100
Particle size
Cha
nge
in p
artic
le
coun
t/ m
g
particle counts
Figure 4.13f Change in counts after Test A6
139
Change in particle count Test A7
0
2000
4000
6000
>5 >10 >15 >20 >25 >40 >75 <100
Particle size
Cha
nge
in p
artic
le
coun
t/ m
l
particle counts
Figure 4.13g Changes in counts after Test A7
The figure shows that the trend in change is similar to change of >15 micron
particles shown in Figure 4.13. However, keeping the poor repeatability of the
equipment in mind the similarity in the trend of results makes analysis worth
considering. The only limitation of the results shown in Figure 4.13(h) is that the
total number of particles counted for only eight different sizes and there may be
many more other sizes smaller than 5 microns or bigger than 100 microns may be
unaccounted.
Change in total counts
02000400060008000
10000120001400016000
A1 A2 A3 A4 A5 A6 A7
Tests
Cha
ngei
n co
unts
/ml
Series1
Figure 4.13h Change in total particle count
140
comparison of particle generation
0
5000
10000
15000
20000
25000
30000
35000
>5 >10 >15 >20 >40 >50 >75 >100
particle sizes
Cou
nts/
ml
A1A2A3A4A5A6A7Gravimetric
Figure 4.14 Comparison of wear particles changes
The generation of particles in a bearing is a complex phenomenon. As particles
during the motion break and their number may change drastically. The particle size
less than 5 microns are very large in number because they are formed by breaking
and rubbing of large particles. Hence, this analysis is merely a guideline and not a
guarantee that particles of any other size were not generated during the wear process.
Keeping in view the problem, the results of weight of total particles present in the oil
sample gives better trend of wear than the total number from different channels of
Quant Alert which is discussed further in the next section.
4.6.1 Gravimetric change (PCg)
The morphology of particles and so their sizes change as the wear progresses in the
bearing. This makes it difficult to analyse the effect of the contaminant particle size
on lubrication and further wear in the system. It is even more difficult to assess the
accurate number of particles generated in the system below five microns. However,
the weight of the total particles generated can be assessed by recording the change in
the total weight of the solid particles present in the oil. This can prove to be a good
technique but errors due to inclusions from the environment, accuracy in
measurements could be the limitations of this technique. The gravimetric
141
measurements taken by Quantalert are presented graphically in Figure 4.15. The
initial values and the corresponding change in their counts are presented in Table 4.1
and 4.2 respectively.
The weight of total particles was measured with the QuantAlert in 10ml sample
which was later converted to g/l as shown in Figure 4.15. The initial weight of the
contaminants in Test A1 could not be detected before the test; however the weight
change of 0.3 mg was recorded after the test. In ideal conditions there was no wear
and hence there should not be wear debris generation but the change may be due to
environment or there may be errors in accuracy.
The change in weight of total debris generated after Test 2 was maximum and
commensurate with the change in particle counts. However, Table 4.2 shows that in
the case of Test A4 even though the particle counts is lower as compared to Test A5,
the change in total weight of the solid particles is higher due to the presence of
particles >100 microns. The particles of >100 microns are lot heavier than the
smaller size particles hence the influence of the number of smaller size particles is
not so severe on the total weight. Change in weight after Test 6 is the lowest, and
similar to the change in particle count. Change in weight after the Test A7 is
comparable to Test A3. These results show that the gravimetric change is also
significant in case of Test A1 compared with Test A6.
There is no good explanation as to why the change in particle count is not similar to
the change in total weight of the particles. The only explanation could be that the
particle morphology changes with time and there is no guarantee that particles larger
than 15 micron changes in the same order as other particle. Hence a comparison of
total weight change, with the change in particle count of size larger than 15 microns,
is not a valid comparison. Though gravimetric results are expected to give more
reliable results, the source of particles or the material being weighed is not
necessarily the wear particles or suspended solid contaminants. There is likelihood
that polymer material which is not solid may also be agglomerated and weighed
during the gravimetric measurements. Overall the particle count technique did not
prove to be a reliable method as claimed by the equipment manufacturers.
142
Particle weight change
0
5
10
15
20
A1 A2 A3 A4 A5 A6 A7
Tests
Cha
nge
in w
eigh
t (g/
l)
total w t.
Figure 4.15 Changes in total weight of contaminants
4.7 Maximum Wear Depth (WDmax)
The worn area in the bearing varies from one test to another. The maximum wear
depth is independent of the projected area and that decides the wear profile for each
worn bearing. Wear profile of each bearing is unique for different bearings worn
under the same operating conditions and hence they create their own signatures.
Wear depth is the maximum departure of the worn surface from the original surface
in the radial direction within the wear zone.
There were a number of options to record the maximum wear depth (WDmax). In fact
the maximum change in bearing ID is the maximum wear depth. Data shown in
Table 4.1 and Table 4.2 are the average value of changes in ID and OD of the
bearings and shafts respectively. These measurements were taken with sophisticated
measuring apparatus such as Metroscope and HP Laser system. Identifying
maximum wear depth as an easy tool for wear measurements, another method was
developed where magnified out-of-roundness trace was used to achieve higher
accuracy. This method will be discussed further in the following chapter. However,
using sophisticated measuring devices was found to be time consuming and using
out-of-roundness trace involved subjectivity. Hence a simple method was used where
bearing ID was measured with a bore gauge at several locations in the wear zone of
the bearing and the highest value of change in ID was recorded as maximum wear
143
depth (WDmax).
The results showed that the WDmax parameter can be used successfully as a wear
measurement parameter similar to that of wear scar diameter in four ball test
machine. The results have been presented in Table 4.2 and were compared with the
maximum change in bearing ID measured with a Metroscope and hole-gauge-test.
Figure 4.16 shows that the trend of measured WDmax values are similar to change in
maximum ID though the values are slightly lower than the change in maximum
bearing ID measured with a Metroscope. The maximum variation is close to 25% for
Test A7, where change in IDmax is 16 microns and change in WDmax is 20 microns.
Minimum variation was recorded for Test A3 where the measured WDmax is 29
microns whereas the change in maximum ID is of the same order i.e. 28 microns.
Figure 4.16 shows the results graphically where maximum wear depth in case of Test
A1 could be ignored as it is too low and could be due to human error in
measurements. The results of Test A2 and A4 are the same with WDmax values 42
microns. These results are susceptive as Test A4 results indicate that the phosphate
ester additive was ineffective. The sources of variations may be several such as;
human error and lack of user controlled environment. However the trend is same as
change in IDmax.
Comparison of maximum wear depth and change in bearing ID
01020304050
A1 A2 A3 A4 A5 A6 A7
Tests
WD
max
/ ID
ch
ange
(mic
rons
)
WD max ID max change
Figure 4.16 Changes in maximum wear depth
144
4.8 Changes in Minimum Oil Film Thickness (Hmin)
The change in micro-geometry of the bearing due to successive wear influences the
tribological performance of a bearing. This influence is reflected through the change
in minimum oil film thickness in the bearing contact. The change in minimum oil
film thickness is a compound effect of various factors; the main ones are: change in
radial clearance, out-of roundness, roughness, and wear depth. The other most
important factors are the presence of solid contaminants in and around the bearing
contact which creates restriction to the free flow of the lubricant and lastly the
antiwear additives, which influences the motion of the solid contaminants.
As wear progresses in the bearing the radial clearance increases and the minimum oil
film thickness decreases. Reduction in minimum oil film thickness during the test
causes a rise in oil temperature and/or progression of wear. Though, this parameter is
not directly related to the wear, however, as a result of wear, film thickness reduces
and decreases the load capacity of the bearing. If this phenomenon continues, a
situation arises when film thickness reduces to an extent that metal to metal contact
takes place, which brings hydrodynamic regime to mixed and boundary lubrication
regime, and finally scuffing takes place in the bearing contact.
The operating parameters were chosen in such a way that the oil film thickness and
sliding distance remained the same for all the test conditions. Since radial clearance
of each test bearing was different the speed required to achieve 16 micron minimum
oil thickness was also different. Thus test duration for each case varied such that the
total sliding distance remains the same. Minimum oil film thickness was recorded at
three intervals of time during the tests i.e. at the start (Hxps), middle (Hxpm) and end
(Hxpe). While recording the film thickness measurements it was observed that the
readings were fluctuating in a random manner. Several probe output readings were
recorded out of which only relevant data was considered. The output of the proximity
probes could not be directly converted to film thickness as they were giving the
relationship with the minimum distances between the tip of the proximity probe and
the shaft sleeve surface. Using a geometrical method developed for this study
(explained in Chapter 3) the minimum oil film thickness was calculated from the
output recorded in millivolts. Minimum oil film thickness at the start of the each of
145
test to achieve ‘K’ ratio close to one have been recorded in Table 4.1
The changes in minimum oil film thickness values calculated with the help of the
proposed geometrical method are presented in Table 4.2. The table shows that the
film thickness reduced with time as contaminants caused wear in the bearing. In a
worn bearing the minimum oil film thickness was based on the eccentricity of the
shaft sleeve centre which occurs in the direction of line of centres.
In case of a worn bearing the minimum oil film thickness does not necessarily occur
along the line of centres because the wear depth along the line is normally a deep
trough. However, in this analysis minimum oil film thickness was calculated based
on the eccentricity in the bearing and wear depth was not considered.
Minimum oil film thickness was recorded for each test bearing of known radial
clearance ‘C’ and operating speed ‘N’ for a constant load of 500N. The values of
film thickness recorded at the start (Hxps), in the middle (Hxpm) and at the end (Hxpe)
of the experiments are shown in Table 4.4. Figure 4.17 shows the change in
minimum oil film thickness results graphically. The change in minimum oil film
thickness was calculated by subtracting Hxpe readings from Hxps readings. The values
show that as wear progresses the value of film thickness gets reduced.
Table 4.4 Changes in measured minimum oil film thickness
Tests C (μm)
N (rpm)
Hxps (μm)
Hxpm (μm)
Hxpe (μm)
ΔHmin (μm)
% change
A1 82 400 17.86 17.82 17.82 0.04 0.2
A2 82 420 17.24 11.01 6.43 10.9 63.2
A3 86 470 18.28 15.107 14.18 4.1 22.4
A4 80 400 17.46 14.41 8.96 5.8 33.2
A5 89 500 15.62 8.0 10.47 5.15 32.9
A6 89 500 15.6 13.06 11.80 3.8 24.4
A7 80 400 14.9 11.13 11.10 3.8 25.5
The following equation derived by geometrical method in Chapter 3 (Equation 3.3)
was used to calculate the eccentricity in the bearing and subsequently by substituting
the value in Equation 3.6 the minimum oil film thickness was calculated for different
146
operating conditions:
Change in minimum oil film thickness
0
2
4
6
8
10
12
A1 A2 A3 A4 A5 A6 A7
Tests
Cha
nge
in m
in. o
il fil
m
htic
knes
s (m
icro
ns)
Figure 4.17 Reduction in minimum oil film thickness
Figure 4.17 shows that the reduction in minimum oil film thickness was negligible
in Test A1. This was mainly because the bearing was operating on a full
hydrodynamic oil film throughout the test. In the absence of solid contaminants,
wear occurred only during starting and stopping of the tests. In Test A2, where 4g/l
Al2O3 was used as a solid contaminant, the wear resulted in reducing the minimum
oil film thickness drastically i.e. from 17.24 microns to 6.43 microns; almost 63%.
Reduction in minimum oil film thickness was also recorded in other test conditions.
For Test A3 the value changed from 18.28 microns to 14.18 microns, a 22%
reduction. For Test A4, with phosphate ester additive present, the minimum oil film
thickness reduced from 14.76 to 8.96 microns, a 40% reduction. Wear results
obtained from other methods confirm the poor performance of the phosphorous
based additive used in Test. Test A5 shows that the change was greater than
expected. In Test A3 the reduction in minimum oil film thickness was 5.15 microns,
33% lower than the initial value. Changes in Test A7, A6 and A3 are of similar
magnitude: 25%, 24% and 22.4% respectively. The measurements recorded for Test
A5 show that there is an error in recording Hxpe (10.47 microns) value, which is
( )( ) ⎥
⎥
⎦
⎤
⎢⎢
⎣
⎡ −+−−
⎪⎭
⎪⎬⎫
⎪⎩
⎪⎨⎧
−+−
−−
⎭⎬⎫
⎩⎨⎧ −
−−−+=s
ybxbs
ybxb
xb
s
ybybsYbS R
PRPRR
PRPRPR
RPR
PRRPRRe.4
)()((.
)()(2).(
).(.222
22
22
147
higher than the Hxpm value of 8 microns, and it cannot be so. Like other wear test
methods, this method also proved to be reliable only for cases where changes due to
performance of the additive are discernable.
The overall results do not show a clarity and high confidence level for predicting
wear in the bearing contact. Though the required minimum oil film thickness was 16
microns but measured values vary from 14.9 microns to 18.28 microns. This may be
due to error in measurement or human error in maintaining constant test conditions
such as temperature, pressure and speed. The tests were noisy after adding the
contaminants in the oil. Vibrations in the bearing also caused difficulty in recording
the measurements fluctuations in reading were excessive hence data was analysed
thoroughly and redundant or irrelevant data was discarded.
The rate of film thickness reduction was more pronounced in some tests; for
example, in Test A2 the rate of film thickness reduction up until middle of the test
was comparable that in the other tests, except Test A5. However, in Test A2 the loss
in film thickness from middle to end of the test was rapid, followed by that in Test
A4. The film thickness reduction rate was minimal in Test A3. In tests A5, A6 and
A7 these were rather similar. In case of Test A5 the minimum oil film thickness
reading is unexpected and there seems to be some error as the value observed in the
middle of the test was 8.0 microns, whereas at the end of the test it was 10.47
microns. The rate at which the film thickness reduced and its mechanism may be
linked to the additive chemistry which is beyond the scope of this study.
The above results show that a bearing operating at ‘K ≈ 1’ gives rise to more wear
and low oil film thickness, which leads to higher friction. Even though the minimum
oil film thickness of half a micron can be recorded with the proximity probes,
excessive vibrations caused unacceptable fluctuations in readings. Thus the norm of
operating a hydrodynamic bearing at a lambda ratio 10 is not universally applicable.
4.9 Comparative Analysis of Techniques and Results
This section compares the methodologies adopted in this study as a part of the multi-
wear parameter approach (MWPA). The outcome of this comparative study helped in
choosing the most suitable wear measurement technique that helped in developing a
148
tool for characterising antiwear additives. Furthermore, based on the selected wear
measuring technique, the performance of various antiwear additives was evaluated.
4.9.1 Methodologies
Weight loss is one of the most commonly used parameter to quantify wear, however
it can give false results if the wear is too small, which can be further exacerbated due
to human or equipment error. For example, Test A7 showed a weight gain of 0.09 g
contrary to what is expected; while all other tests showed a weight loss. The above
analysis showed that it is difficult to conclude as to which parameter was dominating
i.e. wear or metal transfer due to embedding. Since, micrographs show that the trend
of wear and metal transfer in Test A7 is similar to that in the other tests (Appendix-
D), the reason for this contradiction (i.e. weight gain in Test 7) can be an error in
measurement.
Out-of-roundness results show the wear reducing behaviour of additives clearly,
because the graphs are highly magnified. These traces give a unique signature of
wear reducing behaviour of the additives, and bearing wear can be measured with
higher precision. The special feature of this method is that it gives highly magnified
view of the departure of circumference from the ideal shape, without magnifying the
overall size of the test piece. Also, a small geometrical change which may not be
possible to detect from other techniques, can often be observed visually by
comparing the out-of-roundness traces taken before and after the tests. Thus the
possibility of error in measurements is lower as compared to that in the other
methods.
The methodology based on change in radial clearance gives results with poor
repeatability. More over the results show that the radial clearance varies along the
circumference of the bearing. However, this methodology can give better design
guidelines if its metrology can be improved as discussed in chapter 3. The method
highlights the need of the correct metrological practices. The method itself needs
more improvements for predicting results with desired confidence.
Roughness measurements raised several issues, such as, selection of the
measurement points within the wear zone, as roughness varied significantly from one
149
point to another; hence, it is hard to choose a representative sample area. The same is
true with the shaft sleeve as well. Averaging the readings taken at several locations
does not give a true picture, and there was no repeatable pattern in the wear reducing
behaviour. These problems were encountered with both bearing elements, i.e.
bearing and shaft sleeve, in transverse as well as circumferential directions. Though,
efforts were made to consider the most representative readings, these did not depict
any relationship of preferential path, as proposed by Martin (1991). However, the
measurement of surface roughness helped in identifying the severity of abrasive
wear. As stated earlier the roughness values are statistically calculated numbers, in
this case these are not controlled roughness, hence random pattern of roughness
cannot be compared with a controlled roughness case.
The methodology based on measured maximum wear depth gives reasonably reliable
results. Experimental data demonstrated that the pattern of wear reducing behaviour
of additives predicted by this method is commensurate with the other accepted
techniques. However, locating the point of maximum wear depth within the wear
zone remains a challenge and hence the method may not be highly reliable.
In the particle counting methodology, even a 4 % concentration of Al2O3 in oil
caused opacity; consequently, the oil was diluted several times to minimise this
effect. However, the repeatability of the results was still poor. The results showed
that the variations could be as much as 50% to 100%. In some cases where results
were repeated after few days the repeatability was even poorer and hence an error
analysis was not thought to be a useful exercise. The total weight of the suspended
particles (in the oil samples) was recorded using the QuantAlert particle counter.
This methodology showed similar wear reducing trend as other techniques. However,
keeping in view of the poor repeatability of the results obtained from this equipment,
this technique cannot be relied upon.
Minimum oil film thickness measurements showed that proximity probe method is
reliable and gives results with desired accuracy when there are no contaminants
suspended in the oil. Minimum oil film thickness was recorded at the start, in the
middle, and at the end of the test. However, for high concentration of solid
contaminants the results were less reliable, due to rapid drop in oil film thickness
150
leading to excessive vibrations.
4.9.2 Additive performance
Analysis of the results showed that it is difficult to predict the performance of an
individual antiwear additive based on just one of the wear measurement techniques,
due to very little difference between the performance results and the precision of the
existing wear measuring methodologies. Ranking of antiwear additives based on
their wear measurements is shown in shown in Table 4.5.
Table 4.5 Comparison of performance of antiwear additives
No Technique Test A3 Test A4 Test A5 Test A6 Test A7 1 ΔWb (gm) 3 5 4 1 2*
2 ΔORb (μm) 3 5 4 1 2 3 ΔC (μm) 3 5 4 1 2 4 ΔID (μm) 3 5 4 1 2 5 Δhmin 3 5 4 1 2 6 ΔWD max 3 5 4 1 2 7 ΔPCg 3 5 4 1 2 8 ΔWs (gm) 2 5 4 1 3 9 ΔOD (μm) 5 3 4 2 1 10 ΔPC (count/ml) 2 4 5 1 3 11 Rb (microns) 4 5 3 2 1 12 Rbt (microns) 4 5 1 3 2 13 Rs (microns) 1 5 2 4 3 14 Rst (microns) 4 3 1 5 2
(* Presumed weight gain was an experimental error)
As shown in the table, techniques 1 to 7 depict same results, whereas, the remaining
techniques show much variance in additive performance. Table 4.5 shows that the
additive used in Test A6 proved to be the best and that of the additive used in Test
A4 was the worst. However, technique number 2 (out-of-roundness method) gave the
most distinct results for the various additives, and techniques 1 and 3-7 confirm these
results. Therefore, the out-of-roundness method has been selected for the
characterisation of the additives.
151
4.10 Conclusion
The analysis of results helped in understanding the behaviour of different antiwear
additives on the bearing element’s wear. It helped in selecting the most appropriate
wear measurement method for characterising the antiwear additives and also in
determining the effect of change in micro-geometry on the tribological performance
of the bearing. The following conclusions were drawn from the analysis:
• This chapter compared six methods including fourteen wear measurement
parameters that comprise the MWPA (multi-wear parameter approach) suite.
The aim was to identify the most efficacious testing methodology for
measuring wear under dusty environments.
• The analysis of data showed that the antiwear additive used in Test A6
performed the best and additive used in Test A4 had almost no antiwear
properties. This finding was supported consistently by seven measured wear
parameters. There is a clear distinction between the wear caused by pure base
oil, and that by oils containing solid contaminants. However, there is only a
marginal difference in the performances of some additives, which is rather
difficult to quantify.
• Weight loss is the parameter which quantifies wear loss in a tribological
component most directly; however, significant errors are likely if the amount
of wear is too low. The out-of-roundness emerged as the most reliable and
suitable methodology for quantifying wear.
• This research revealed a high concentration of solid contaminants in and
around the bearing contact. Accumulation of solid contaminants at the contact
entry and within the contact zone restricts the lubricant flow, and hence,
increasing level of mixed or boundary lubrication conditions occur as the test
progresses.
• The reduction in minimum oil film thickness was considered as the
cumulative effect of all the micro-geometry parameters, as well as that of the
concentration of the contaminants in the contact zone. This led to rapid
reduction in minimum oil film thickness, and excessive wear;
152
consequently, the film thickness measurements did not produce repeatable
results. However, it clearly depicts the adverse effect of changed micro-
geometry and contaminants on the bearing’s tribological performance. In
other methodologies, either due to lack of precision in measurements, or
variation in measurement location resulted in poor data repeatability.
• Surface roughness measurements were random and inconsistent within the
wear zone, hence could not be used for further analysis. The concept of
preferential path could also not be established because actual surface
roughness patterns were random.
• Particle count method did not prove to be reliable because of heavy
concentration of solid particles, which could not be counted reliably.
Furthermore, break up of polymeric additives into particle-like structures
caused interference with particle count.
• In conclusion, the out-of-roundness technique gave the most reliable results
and has been chosen for the characterisation of the additives.
153
CHAPTER 5
5. CHARACTERISATION OF ANTIWEAR ADDITIVES
5.1 Overview
Use of antiwear additives is common in situations where machines are prone to
excessive wear. Machinery working in mining industry, highway construction or any
dusty environment is subject to substantial solid contamination. Antiwear additives
enhance the life of tribological components under worn conditions, and, to some
extent, reduce energy consumption due to reduction in frictional forces between the
mating surfaces.
The question has often been asked if the added cost of the additives is justified under
such conditions. Several additive manufacturers claim their product is better than the
others. In general, industry needs some universally agreed guidelines to judge the
efficacy of a product. But, there is no reliable measure by which two products can be
compared, except by evaluating their performance in the machine. It may be possible
to prove that one product is better than another under certain laboratory conditions;
however this may not be replicated in real machine operation.
The life expectancy of a machine depends on several factors, and improvement in a
component’s life may be so marginal that it is hard to isolate that factor which has
maximum influence. Similarly, the difference in the performance of antiwear
additives is so small that it may be hard to distinguish between the effect of antiwear
additives and the other factors influencing the performance of a bearing.
Journal bearings normally operate in the hydrodynamic regime, where the two
mating surfaces are separated by more than 10 times the composite roughness.
However, very small amount of wear is expected in such bearings, especially when
bearing starts or stops. The amount of wear is so small that it may be hard to record.
However, a small change in coefficient of friction, and consequential reduction in
wear can result in significant savings when considering industry-wide application,
154
where thousands of bearings are used. Reduced downtime costs due to lower failure
rates and energy saving are the key factors for targeting an efficiently operating
machine.
The interaction between solid contaminant particles and the bearing surfaces, as well
as the interaction amongst the solid contaminant particles causes additional wear and
friction. The tribology of such bearings is not well understood. There is a need to
develop a standard methodology by which the wear reducing performance of an
additive can be judged. This will help in selecting the proper lubricant for a specific
machine and its operating environment.
In this research different additives have been compared for journal bearing wear
performance. Test bearings were lubricated with oil containing Al2O3 as a solid
contaminant. Various tests were performed after mixing different types of antiwear
additives in prescribed dosage. Tests were run for the same operating conditions and
the same sliding distance. The effect of these additives on journal bearing wear was
observed using Multi-Wear Parameter Approach (MWPA), using various wear
measurement techniques. Analysis of the data obtained from different measurement
techniques used under MWPA showed that weight loss and out-of–roundness are
more reliable wear measure techniques. Using these two techniques, a methodology
has been developed to characterise the antiwear additives for their wear reducing
behaviour.
5.2 Wear Measurement by Weight Loss
Weight loss is one of the quantitative measures of wear in bearings and gives direct
evidence of wear quantity. However, human or equipment errors in measuring the
minuscule weight loss can reduce the reliability of the results. Therefore our aim is
to develop a method that does not rely solely on the measurement of weight loss. It
was shown in Chapter-4 that the out-of-roundness trace of a worn test bearing can
provide a magnified view of wear in the bearing. By measuring the change in the
shape of a trace, wear in a bearing can be quantified reliably and easily. It is also
important to note that the shape of the out-of-roundness trace for each additive is
different and hence each additive gives a unique signature of its anti-wear
performance. Literature review revealed that the out-of-roundness measurement has
155
not been used effectively, thus far. We postulated that out-of-roundness measurement
can be an efficacious method for wear characterisation. Its features such as: unique
wear signatures and high magnification lead to better accuracy and consistency in
calculating weight loss. In this research a methodology has been developed to
calculate weight loss in worn bearings from their out-of-roundness traces for known
bearing length and material density.
Out-of-roundness trace magnifies the deviation of the circumference from the true
circular shape at a fixed bearing cross section without magnifying the overall
diameter of the circular work piece. The region of the trace, where worn surface
departs from the original is identified as the wear zone. By comparing the traces of a
bearing before and after the test, the zone can be identified by visual inspection, in
the form of a wear crescent, as shown in Figure 5.1. A process has been developed,
by which the area of this wear crescent can be measured with desired precision, and
subsequently wear volume and weight loss can be computed for known bearing
length and material density.
Figure 5.1 Wear profile of a worn bearing
5.2.1 Wear Computation from Out-of-roundness Trace
The out-of-roundness trace of a bearing gives more precise distance between the
centre about which the work piece is rotated and the circumference at any angular
Unworn surface
Maximum wear depth
Wear zone width
Wear zone
Worn surface
156
location. The trace of a worn bearing is similar to the end view of the bearing shown
in Figure 5.1. The departure from normal of the bearing surface in the wear zone is
characterised by the wear depth, which varies along the wear crescent. The area of
the wear crescent is the wear area at a fixed cross section of the bearing and is termed
as cross sectional wear area (CSWA). In order to find the weight loss, the CSWA
value needs to be determined with high precision, subsequently wear volume and
weight loss can be computed, for known bearing length and material density.
Figure 5.2 Roundness measurements locations
The out-of roundness traces were recorded at three fixed locations along the bearing
length, i.e. at H1= 5mm, H2 =15mm and H3 = 35 mm as shown in Figure 5.2. Ideally
the CSWA should be the same all along the bearing length. However, it changes
from point to point along the length due to two reasons: firstly, the changes in out-of-
roundness from one cross section to another, and secondly, wear due to misalignment
of the shaft when bearing is loaded. Figure 5.3 shows the effect of misalignment
where shape of the worn area along the bearing length changes with the degree of
misalignment. Under ideal conditions the worn area should look like a rectangle, as
shown in this (Figure 5.3) by solid lines. Similarly, under extreme misalignment
conditions the wear surface may look like a triangle, as shown by dotted lines.
However, the size of the triangle will depend upon the radial clearance as well as the
load applied. If the amount of misalignment is small, as is the case in this study, the
worn area may look like the shaded area shown in Figure 5.3.
H3 =35mm L =40mm
H1 = 5 mm
H2 =15 mm
157
Figure 5.3 Wear zone shape
The width of the worn area at any point along the bearing length is termed as wear
zone width (WZW) at that point. In other words WZW varies along the bearing
length depending upon the degree of misalignment. The WZW may vary from point
to point all along the bearing length, either due to change in out-of-roundness, due to
misalignment, or both. Thus the CSWA may vary from one cross section to another.
Initially, three out-of-roundness traces were acquired to find the average value of the
CSWA. The CSWA computed on the two ends varied around 5% only; hence the
variation can be ignored. Under such circumstances, only one trace that gave largest
WZW was considered and it was presumed that the CSWA is the same for all
bearing cross sections. As stated the CSWA needs to be estimated with high
precision, therefore a systematic procedure has been developed to minimise errors in
measurements.
5.2.2 Computation of Cross Sectional Wear Area (CSWA)
Computation of weight loss from the out-of-roundness method requires calculation
of area of the wear crescent with high precision. The technique for measuring this
Cross Sectional Wear Area (CSWA)
Bearing length
wear zone width WZW
Extreme misalignment wear shape
Ideal rectangular wear shape
158
area is based on the measurement of out-of-roundness of the circular work pieces,
using Talyrond equipment.
Ideally, an out-of-roundness trace of a perfectly circular bearing should be a perfect
circle. An enlarged view of a trace obtained for a bearing before the test is shown
Figure 5.4. This trace shows that the shape is not a perfect circle, implying that it has
some degree of out-of-roundness. In order to understand the method of computing
the CSWA, it is necessary to understand the principle used by the Talyrond
equipment. An out-of-roundness trace of a circular work piece is acquired by
rotating it against a transducer which records the circumference, showing the
departure of the periphery from the true circle.
The out-of-roundness from this trace is measured with the help of a monograph. This
monograph is a transparent plastic sheet on which concentric circles of different
diameters are marked. The monograph is placed on the trace in such a way that one
of the concentric circles sits inside the trace and touches the trace at minimum two
points. The farthest point of the trace touching the largest diameter concentric circle
gives the maximum departure of the traced circumference from the touching circle.
This maximum departure can be measured in terms of the number of radially
graduated divisions marked on the trace graph paper (Figure 5.4).
Each division marked on the graph paper corresponds to a calibrated value in
microns, depending upon the magnification chosen to acquire the trace. Multiplying
the divisions of departure with the magnification factor (in microns per division)
gives the out-of-roundness of the bearing in microns.
Measurement of the CSWA with precision is the most important step for the weight
loss computation process. The following step-by-step process has been developed
for the computation of CSWA with precision.
159
Graduations marked
Figure 5.4 Out-of-roundness ‘before test trace’
Step1: Prepare out-of-roundness traces
Select a high magnification (say few hundred times) on a photocopier depending
upon the paper size available and the accuracy required. Photocopy the ‘before test
trace,’ on an overhead transparency (Figure 5.4). Using the same magnification
photocopy the ‘after test trace’ for the same bearing location – preferably on an
overhead transparency (Figure 5.5).
160
Figure 5.5 Out-of-roundness ‘after test trace’
Step2: Identify Wear Crescent
Place the ‘before test trace’ transparency over the ‘after test trace’ and try to match
the shape, identifying the wear zone. By topography complete the wear zone part of
the circumference of the after test trace by drawing the crescent arc with dotted line.
Thus, the wear crescent can be identified by locating the inner crescent arc and the
outer crescent arc (Figure 5.6). These crescent arcs are also shown in Figure 5.7
where an inner crescent arc has been drawn on an actual out-of-roundness graph
paper. Thus on the ‘after test trace’, the worn part of the ‘before test trace’ can be
identified clearly.
161
Figure 5.6 Computed out-of-roundness shape of a worn bearing
Step3: Mark computed maximum wear depth (WD)
Figure 5.8 shows the process used for finding the computed wear depth (WD).
According to this process the centre of the ‘before test trace’ component called trace
centre ‘O’ can be marked by visual inspection; else by inscribing a concentric circle
marked on the monograph touching the ‘before test trace’ part at minimum three
points.
Mark mid-point D on the outer crescent arc in the wear zone in such a way that ‘OD‘
is the largest distance between the trace centre (O) and the outer crescent arc. Now
draw the line OD intersecting the inner crescent arc at W (Figure 5.8). Distance WD
in this figure is the computed maximum wear depth.
Step4: Mark wear depths (wn dn)
It is observed that the angular extent of WZW is less than 1200. Keeping in view this
assumption, mark nodes w1d1 and w17 d17 on 60o angle either sides of the mid-point D
respectively. These are the two extreme ends of the wear crescent. Divide the outer
crescent arc in ‘n’ number of parts. In this study the inner crescent arc has been
Wear zone start
Wear zone ends
Crescent Arc (inner)
Crescent Arc (outer)
Trace centre ‘O’
162
divided in 16 equal parts. There is no specific reason for choosing this number except
that it may be helpful in developing a simple theoretical model in future. Mark
nodes from w1 to w17 total n+1 nodes (from w1 to wn+1). Now draw lines Od1 Od2
Od3 etc. up to Od17, intersecting inner crescent arc at w1, w2, w3 etc corresponding to
d1, d2, d3, etc up to d17. Thus giving w1d1,.w2d2,. w3d3,. w4d4,.up to w17d17,.as wear
depths at 17 equally spaced nodes within the wear zone (Figure 5.8). Wear depths
can also be identified by generic term wndn where ‘n’ represents the ‘nth’ node. It
should be noted that on some extreme end nodes the wear depth can be even zero
depending upon the angular extent of the WZW.
Step5: Measure nodal distance
Measure the distance between the two nodes called nodal distances in millimetres for
a 120 0 angle arc.
Figure 5.7 Actual trace of a test bearing with redrawn shape
Computed shape before test
Graduations
Maximum wear depth
CSWA
Inner Crescent Arc
163
Step 6: Find out Scale Factor (SF)
Measure the distance between the two grid points of the graduations marked radially
on the enlarged (photocopied) trace. Compare this value with the equipment
manufacturer’s calibrated value for each division (microns/division) for the chosen
magnification and convert the distance between the two grids into scale factor (SF) in
microns/mm.
Figure 5.8 Wear depth measurement at different nodes (wndn)
Inner Crescent Arc
Nodes
w1d1
w17 d17
WD
Out-of-roundness trace
d2
d6
O
d3
D
W d 16
Trace Centre
Outer Crescent Arc
164
Table 5.1 Computed maximum wear depth data
Nodes/ wear depth (wndn)
Grid distance mm
(wndn) (µm) A2
(wndn) (µm) A3
(wndn) (µm) A4
(wndn) (µm) A5
(wndn) (µm) A6
(wndn) (µm) A7
1 2.616 0 0 0 0 0 0 2 5.233 2.4 0 2 0 1.5 1 3 7.85 8.8 2 5 3 2.5 5 4 10.466 15.2 6 13 11 5 10 5 13.083 22.4 14 16 16 8 14 6 15.7 30.4 18 21 21 9 16 7 18.316 42.4 23 28 27 10 20 8 20.933 43.2 24 34 29 10 21 9 23.55 41.6 29 35 31 9 22 10 26.166 38.4 28 33 31 7.5 21 11 28.783 30.4 24 36 30 5.5 18 12 31.4 24.8 21 34 29 4 12 13 34.016 18.4 15 29 24 2 5 14 36.633 11.2 7 26 20 1 0 15 39.25 7.2 2 16 11 0 0 16 41.866 3.2 0 10 0 0 0 17 44.483 0 0 0 0 0 0
Step7: Calculate wear depths (wn dn)
Measure wear depths wndn for all ‘n’ number of nodes from the enlarged trace and
multiply them with the scale factor (SF) to convert them into microns. These values
have been recorded in Table 5.1 for all the tests. The values highlighted with bold are
the computed maximum wear depths (WD) recorded from this method.
Step 8: Develop graphs and equations
The distance between the nodes and the corresponding wear depths (wndn) can now
be used for plotting wear profile using Microsoft Excel program. This profile looks
different from the actual out-of-roundness profile because in this case the X-axis is a
straight line representing the inner crescent arc. Furthermore, equations can also be
165
developed to represent the outer crescent arc. These equations are discussed in
Section 5.2.3.
Step 9: Calculate CSWA
Develop a suitable numerical method for calculating the area of the wear crescent or
use equations of these curves and find the area between the two curves. Equations
derived for this purpose are discussed in the following section.
5.2.3 Wear Characteristic Equation
The wear depth data obtained from the out-of-roundness traces was used for
representing the outer crescent arc as polynomial equations. The angular
displacement on the inner crescent arc was converted to linear distance in millimetres
as explained in step-8 of the previous subsection.
These wear depths were treated as independent variable and a graph between the
linear distance and the wear depths were plotted with the help of Microsoft Excel
program. Add Trend Line Function of the Excel program gave the best fit curve as
well as the equation of the curve. The equations developed for these best fit curves
are called “Wear Characteristic Equations (WCE)” and were developed for all the six
test conditions using wear depth data shown in Table 5.1.
Best fit curves and their corresponding WCE for Test A2 and A3 are given in Figure
5.9 and 5.10 respectively. In developing these equations and best fit curve the
distance between the nodes is taken in millimetres whereas wear depth is microns.
Polynomial equations of higher order derived to represent the curves the error in
wear depth calculations is minimised.
The magnification of the trace is also very high; the combination of these two factors
minimises the errors in CSWA calculations. These equations also serve as unique
characterisation of an individual antiwear additive, and can be used for deriving the
area under the curve: i.e. CSWA.
166
The WCE’s developed for the tests A2 to A7 are given as Equation 5.1 to 5.6
respectively, where y represents the wear depth in microns and X represents the
angular displacement along the bearing circumference, in millimetres.
3774.1178.2499.134.3015.00088.00002.0 23456 ++−+−+−= XXXXXXy (5.1)
64.414.776.2226.00044.0.106.106 2345566 +−+−+−= −− XXXXXxXxy (5.2)
212.013.1123.4822.3627.8903.429.5 23456 +−++−−−= XXXXXXy (5.3)
55.091.628.127.8692.12139.5403.8 23456 +−++−−−= XXXXXXy (5.4)
04.091.1009.7037.4808.1180.1513.3 23456 +−+−−++−= XXXXXXy (5.5)
94.145.32174.4364.285637.9627.13 23456 ++−+−= XXXXXy (5.6)
167
Test A2- Wear profiley = -0.0002x6 + 0.0088x5 - 0.1501x4 +
0.9686x3 - 1.4993x2 + 2.1718x - 1.3774
-10
0
10
20
30
40
50
G1 G3 G5 G7 G9G11 G13 G15 G17
Node no
wea
r dep
th m
icro
ns
WDmaxPoly. (WDmax)
Figure 5.9 Wear Characteristic Equation for Test A2
Wear Profile Test A3y = 6E-06x6 - 6E-05x5 + 0.0044x4 -
0.2266x3 + 2.7687x2 - 7.1473x + 4.6425
-505
101520253035
1 3 5 7 9 11 13 15 17Node number
wea
r de
pth
(mic
rons
)
wear profilePoly. (wear profile)
Figure 5.10 Wear Characteristic Equation for Test A3
5.2.4 Area measurement by Newton Cotes method
Several methods can be used to find out the area of an irregular shape. Area of the
wear crescent CSWA can be calculated by any numerical method such as: Simpson’s
Rule, Trapezoidal Rule, etc. for known equally space ordinates. The CSWA can also
be calculated as area between the two curves, i.e the ‘inner crescent trace’ and the
‘outer crescent trace’. The equation of the outer crescent is the WCE derived in the
previous section, and the inner trace arc can be treated as a straight line due its very
168
low curvature.
A planimeter can be used to measure the area of an irregular shape; therefore it could
have been used to measure CSWA directly. However, CSWA was calculated by
using a numerical method for known node spacing and wear depths (wndn).
The Newton-Cotes Weisstein (2004) numerical method (an enhanced algorithm of
Simpson’s rule) was used due to simplicity and accuracy. The CSWA was computed
by approximating the integral of a function f(x) using quadratic polynomials (i.e.,
parabolic arcs instead of the straight line segments in the trapezoidal rule). Simpson’s
rule can be derived by integrating the third-order Lagrange interpolating polynomial
fit, Trott (2004), Weisstein (2004) and was used by Allen and Kline (1971) for
micro-fluid analysis. The procedure for deriving the function is explained as below:
Let the function f (x) be tabulated at X0, X1, and X2 equally spaced by distance h.
Then Simpson’s Allen (1971) Rule states that :
∫ ∫+
=2 2
)()(X
Xo
hXo
Xo
dXXfdxxf (5.7)
Τ ( )21431 fffh o ++≈ (5.8)
Since it is derived using quadratic polynomials to approximate functions, Simpson’s
rule actually gives exact results when approximating integrals of polynomials up to
cubic order.
In Equation 5.8
Τ = Cross Sectional Wear Area (CSWA) (mm2)
Δh = Δx or distance between the two nodes (mm)
169
f0 = wear depth at the first node (microns)
f1= sum of all the wd’s at odd nodes (i.e. 3,5, 7,……..17)
f2 = sum of all the wd’s at even nodes (i.e. 2, 4, 6…….16)
Thus the wear depth data was substituted in the Equation (5.8) and CSWA values for
all the test cases were calculated. The results of CSWA are reported in Table 5.2.
5.3 Wear Assessment from Out-of-roundness Traces
The wear depth data has been used for computing various wear parameters, such as:
the CSWA, wear depth, wear characteristics equations (WCE), wear volume, and
finally, the weight loss.
Though, each of the above parameters can be used as a wear measure in its own
right, their reliability and practicability need to be adjudged before choosing the best
one for characterising the antiwear additives. Thus each method was compared with
the MWPA technique. However, only two parameters derived from the out-of-
roundness method are comparable to the measured values: computed maximum wear
depth (WD) and computed weight loss (W). Though wear volume is widely used, in
this study it could not be relied upon due to the small dimensional changes caused by
wear. The measured value of maximum wear depth (WDmax) is, similarly less than
ideal, as the wear depth changes abruptly at some points. Thus, computed weight loss
was considered to be the best wear measure parameter, to characterise additives.
Further discussion on these high potential wear parameters is necessary to understand
their utility as well as their limitations.
5.3.1 Maximum wear depth
The ‘measured maximum wear depth’ (WDmax) was determined by using hole-test
gauge as explained in Chapter 3. The ‘calculated maximum wear depth’ (WD) has
been computed from the out-of-roundness traces, which are nothing but the
maximum departure of the surface arch from the original surface arch. In Chapter-4
the maximum change in radial clearance (ΔC) was derived from the change in ID and
OD of the bearing and shaft sleeve respectively, this is theoretically nothing but the
170
maximum wear depth (WDmax).
The values of measured maximum wear depth WDmax, maximum wear depth
computed ‘(WD)’ and maximum change in radial clearance (ΔC) for various tests are
given in Table 5.2. These values have also been compared graphically in the bar
chart shown in Figure 5.11.
Figure 5.11 shows that the trend of wear reducing behaviour of different additives is,
by-and-large, similar to that indicated by most of the other techniques. The computed
values are comparable to the change in radial clearance (ΔC) in most of the cases,
though they are lower by as much as 50% in case of Test A6. By and large the
measured maximum wear depth (WDmax) values are higher than 16% the computed
wear depth (WD) values, e.g. Test A4.
For Test A4, the computed and the measured values are quite comparable, as the
maximum variation in case of Test A4 is only 6 microns (Figure 5.11). All the
techniques used for measuring the maximum wear depth (WD) show that the wear
reducing behaviour of antiwear additives is rather similar, the worst performing case
is when there is no additive present in the oil i.e. Test A2 and the best case is Test
A6.
Measurements of the change in radial clearance (ΔC) parameter are not highly
reliable; however, the measurements of the maximum wear depth parameters
(WDmax, and WD) can be relied upon but for some errors if there is high roughness in
the wear zone.
Table 5.2 Comparison of maximum wear depth*
Tests Change in radial clearance (ΔC)
Max. Wear depth measured(WDmax)
Max. Wear Depth computed (WD)
A2 26 42 43.2 A3 17.5 29 29 A4 21 42 36 A5 18 36 31 A6 5.5 12.5 10 A7 7.5 20 22 *All dimensions in microns
171
Figure 5.11 Comparison of maximum wear depth
5.3.2 Computed wear volume (V)
The change in wear area is not a widely used wear parameter and hence the CSWA was not
chosen as the wear measure parameter in this research. Wear volume is a very commonly
used wear parameter which can be derived from the CSWA.
Measured values of CSWA are the values of the wear crescent at a given point along
the bearing length (L). If test bearings are perfectly circular and the shaft is perfectly
aligned, the value of CSWA will be the same at all points along the bearing length,
and hence the wear volume can be calculated by multiplying CSWA by the bearing
length. For greater precision the volume should be calculated by taking the aveaged
of the various CSWA values measured along the bearing length (see Section 5.2.1).
However, the wear volume (V) was calculated using Equation 5.9.
LTV .= (5.9)
Where L is the bearing length and T is the higher CSWA value computed at one of
the bearing ends. This gives a more conservative valus for the wear volume, and is
desirable for more robust bearing design.
Though it may be complex to measure the wear volume of the wear crescent form
Maximum wear depth comparison
0 10 20 30 40 50
A1 A2 A3 A4 A5 A6 A7Tests
Wea
r de
pth
(mic
rons
)
WDmax measured WD IDmax
172
through out the bearing length, It can be possible to develop a measurement
technique similar to Plint and Partner’s pin-on-disc machine used by Sharma (1994)
for wear measurements where a change in diameter of a taper groove formed on the
test bearing can give the wear volume loss in the bearing. However, in current
situation this parameter cannot be compared with any other MWPA technique.
5.3.3 Computed weight loss (W)
Weight loss is considered to be the most straight forward approach of quantifying the
wear in tribological components. However, its measurement with precision is
difficult when wear in the component is too small. The out-of-roundness trace
method is an alternate method of computing the weight loss with minimal errors and
higher precision. Weight loss in a test bearing can be calculated from CSWA by
multiplying it by bearing length and material density or from wear volume by simply
multiplying it by material density.
The out-of-roundness is another wear measure parameter that can be used for
comparing the wear in machine components; however, it cannot quantify wear in
absolute terms and hence cannot be compared with other methods. The wear volume
measured by out-of-roundness method gives better quantification of wear in
bearings. However, the amount of wear cannot be compared with any other MWPA
technique used in this study. Therefore, computation of weight loss from out-of-
roundness trace can be more useful and reliable in comparing or characterising
different antiwear additives.
Weight loss in the bearing can be calculated for the known wear volume and material
density. The density of the bearing material reported in standard handbooks/
literature for bearing material BS LG2500 is 8g/cm3 (Tzeng and Saibel (1967))
(0.008g/mm3). Thus, weight loss (W) is calculated using Equation 5.10.:
ρ..LTW = (5.10)
Where:
173
W = computed wear
ρ = material density
L = bearing length
Τ = CSWA
The computed weight loss (W), wear volume (V) and CSWA (Τ) is compared with
the measured weight loss (ΔW) in Table 5.3. The computed weight loss and
measured weight loss results are shown graphically in Figure 5.12, where,
discrepancy in results can be seen for Test A7. This discrepancy clearly indicates the
error in weight loss measurement. The result analysis in Chapter-4 showed that
except roughness, particle count and film thickness all other wear parameters showed
the similar trends of additive wear performing behaviour. However, the only other
parameter that showed discrepancy in results was the Test A7 where only weight loss
was the odd reading and it was suspected that the weight gain may be due to metal
transfer. However, the out-of-roundness method verifies and proves that there was an
error in weight loss measurement.
Table 5.3 Wear Volume
Tests (Additives)
CSWA (T) mm2
Computed Wear Volume ‘V’ (mm3)
Computed weight loss (W), (g)
Measured weight loss Δ Wb , (g)
% change
Test-A2 0.738 29.52 0.236 0.25 -0.06 Test-A3 0.542 21.68 0.173 0.15 o.13 Test-A4 0.747 29.88 0.238 0.17 0.28 Test –A5 0.671 26.84 0.214 0.15 0.3 Test-A6 0.164 6.56 0.052 0.08 -.53 Test-A7 0.471 18.84 0.150 -0.09 1.6
174
Figure 5.12 Comparison of computed weight loss and measured weight loss
Results plotted in Figure 5.12 show that the computed values are either higher than
the measured values, or are close to each other, except for Test A7. The behaviour is
similar to that shown by other techniques, where best performing additive is A6 and
the worst one is A4. In case of Test A2 and A6 the computed test results are slightly
higher than the measured results.
Since weight loss is one of the universally accepted wear measure parameters, and
furthermore, using out-of-roundness for calculating the weight loss is the most
reliable method, the calculated weight loss is the most suitable parameter for
characterising the antiwear additives.
5.4 Characterisation of Additives
The results obtained from different wear measuring techniques showed that in some
cases the results for different additives are very close to each other. Thus, it is
difficult to distinguish the wear behaviour of one additive from another. Hence, there
is a need to develop a method by which an additive can be identified for its wear
reducing performance with confidence. This goal can be achieved by using a reliable
wear measure parameter correlating it with other operating conditions. After a
Weight loss estimated vs measured
-0.15 -0.1
-0.05 0
0.05 0.1
0.15 0.2
0.25 0.3
A2 A3 A4 A5 A6 A7
Tests
wei
ght l
oss
(mg)
estimated
measured
175
thorough review, so far the computed weight loss (W) has been identified as one of
the most reliable wear parameter and was thus used for deriving a number called
wear characteristic number (WCN). This number can be used for identifying an
antiwear additive for its wear reducing performance. The method for deriving WCN
is explained in the following section.
5.4.1 Wear Characteristic Number (N)
Wear coefficient is a well known parameter as explained by Rowe (1986) for
expressing the wear characteristics of a material under lubricated conditions. The
standard relationship amongst material parameters is expressed as:
HWK
lV
= (5.11)
Where:
V =Wear Volume (mm3)
I = sliding distance (m)
W = Normal load (N) or W
H = hardness (N/m2)
K = wear coefficient
A similar approach has been used in this research for characterising the antiwear
additives. For this purpose data has been generated using test conditions by adding
Al2O3 as solid contaminant. Though the tests were run for short duration, the
authenticity of the data has been verified using MWPA, and the out-of-roundness
method. Of the various MWPA techniques, investigated, out-of–roundness method,
prove to be most reliable and precise technique. Though, weight loss is used widely,
the out-of-roundness method for computing the weight loss gave higher precision.
176
Wear volume has also been derived from the wear crescent (CSWA). Thus, making
use of these parameters, the Equation 5.11 can be rewritten as:
WlHVN
..
= (5.12)
Where N has been defined as wear characteristic number (WCN or N), V is the
computed volume in mm3, hardness H=1500-1650 kgf/mm2 for Al2O3 and hardness
of bronze is 75 using Rockwell scale B, l is sliding distance in meters and W is the
computed weight loss in mg.
Rowe (1986) attempted to find wear coefficients under lubricated conditions for
some lubricating oils with different viscosities. But these coefficients were calculated
only for clean oils, and no further work has been reported on their implementation.
Following Rowe’s guidelines the additives used in test A2 to A7 were characterised
by their wear coefficients for bronze bearing and steel shaft sleeve pair. The wear
characteristic number (N) obtained from Equation 5.12 for different additives are
given in Table 5.14. Their values vary from 1.1X10 -11 to 9.7X10-12. A higher value
for wear coefficient indicates poor antiwear behaviour of the additive. Additive A6
was found to be the best, with WCN (N) = 3.4XE-12. The highest value of ‘N’ was
obtained with no additive in Test A2, where ‘N’ = 1.5XE-11. The worst performing
additive is A4 with ‘N’ = 1.5XE-11, the same as with no additive. WCN (N) is a
unique number derived from the most reliable wear parameter i.e. CSWA. This
number highlights the wear reducing performance of antiwear additives.
Thus wear characteristic number can be used successfully in identifying the wear
reducing performance of an antiwear additive. The range of this number varies
between from E-11 and E-12; thus the performance can be characterised with
confidence.
177
Table 5.4 Wear Coefficients of antiwear additives
Test Rpm (N)
Sliding/distance (l)
Wear volume m3(V)
Wear Characteristic Number (N)
A1 400 7536 0 0 A2 420 7536 1.48E-08 1.52E-11 A3 470 7536 1.08E-08 1.1E-11 A4 400 7536 1.49E-08 1.5E-11 A5 500 7536 1.34E-08 1.4E-11 A6 500 7536 3.30E-09 3.4E-12 A7 500 7536 9.42E-09 9.7E-12
5.5 Discussion on Results
The results obtained for CSWA show that the trend of wear reducing behaviour of
different additives is similar to that shown by other methods except in case of Test
A7. In this test, weight loss (ΔW) measurement results indicate a weight gain of 0.09
mg; whereas; the results obtained from computed weight loss (W) show a weight loss
of 0.15 mg. The weight loss results also highlight the errors in measurements, which
vary from one test to another: -6% in case of test A2, and +28% for Test A4,. -53%
for Test A5. The results confirm the trend of wear reducing behaviour of additives
revealed by most of MWPA techniques such as: measured weight loss, maximum
wear depth measured and computed (WDmax and WD), change in radial clearance
(ΔC), out-of-roundness and particle weight (PCg). This proves that even though the
weight loss measurement methods is considered to be one of the most reliable and
authentic wear measure parameter, it is not suitable for measuring low wear volumes.
The results of computed weight loss (W) indicate that the computed values are in
general higher than the measured values. This is mainly due to reasons that the
CSWA value considered is not the average value which could be achieved by taking
an average of several CSWA values measured on different points along the bearing
length. However, it could be safely concluded that the measured weight loss results
for Test A7 incur an error. Because no other test method supports this phenomenon
including weigh loss computation. Thus reliability of out-of-roundness method is
proved. Similarly a discrepancy in weight loss for Test A5 also shows that this
178
may be due to error in measuring weight loss rather than computing weight loss. The
possibility of error in computation is low as discussed earlier, because the wear
reducing trend is similar to most of the MWPA techniques used in this research.
The WCN may still have scope for refinement but this has no bearing on comparing
additives performance, because this is the quantity derived from CSWA and bearing
length which is constant.
Methodology developed for characterising the antiwear additives gives a unique
number (N) as wear coefficient which can help distinguishing the wear reducing
performances of any two antiwear additives. The methodology can be further
modified by better average value of CSWA.
Measured maximum wear depth (WDmax) may have some discrepancies because in
case of maximum change in radial clearance method the measured values also take
into account the wear of shaft sleeve and it is expected that the over all change in
radial clearance is greater than expected. Similarly measurement of wear depth can
also be subjective because within the wear zone, some times the contact is rough in
the tested bearing and it is hard to locate a single location with maximum wear depth.
Which is similar to roughness measurement as discussed in the previous chapter. The
change in radial clearance results indicate that the measurements of ID and OD using
hole-test-gauge may not give desired accuracy because wear depth may be too small
as compared to the radius of the hole-test-gauge stylus.
5.6 Conclusion
Out-of-roundness technique was found to be the most suitable technique for
measuring the wear in test bearings. A methodology was developed where out-of-
roundness traces were used to estimate weight loss, which helped in deriving the
Wear Characteristic Number (N). The best antiwear performance was obtained for
Test A6 where N = 1.4 X 10-12; and the worst case was obtained for Test A2 (without
antiwar additive) where N = 1.52 X 10-11.
The results demonstrated that even though the weight loss measurement is
inherently one of the most reliable wear measure parameters, it may not be suitable
for low wear volumes. The out-of-roundness method for computing the wear
179
volume, and weight loss – using the CSWA (cross-sectional wear area) from the
trace – proved to be the most reliable technique.
Precision in computing weight loss (W) can be further improved by measuring the
density of the material as well as bearing length with higher accuracy. Wear
characteristic equations were developed, which can be used to calculate the wear
volume and maximum wear depth, and consequently, derive wear signatures for
individual additives.
Finally, a wear characteristic number (N) was derived; this number is useful for
comparing the efficacy of various antiwear additives. Users as well as manufacturers
of antiwear additives can successfully characterise additives by using the proposed
method. This wear characteristic number (N) can be derived at any independent test
laboratory, and thus, national or international standards organisations can assign
ratings to various antiwear additives to characterise their performance for dusty
applications.
180
BLANK PAGE
181
CHAPTER 6
6. CONCLUSIONS
The main conclusions of this research are presented in this chapter. These
conclusions cover the rational for this research project, knowledge gaps identified
from the literature review, experimental design approach, new knowledge derived
from experimental results and the main contributions of this research to the existing
body of knowledge.
6.1 Problem Statement
This research aimed to investigate the effect of antiwear additives on journal
bearings operating in hydrodynamic conditions. In this research the following three
main aspects have been studied:
a) The effect of contaminants –treated with antiwear additives– on journal bearing
wear.
b) The effect of change in micro-geometry on bearing’s tribological performance.
c) Characterisation of additives using the most suitable wear measuring technique.
The research methodology adopted for this research can be broken into the following
five phases, with matching deliverables:
1. Compare the tribological performance of: a) journal bearings lubricated with
pure base oil, b) base oil containing solid contaminants, and c) oils containing
solid contaminants treated with different antiwear additives.
2. Determine the effect of antiwear additives on wear and micro-geometry of a
journal bearing, operating with lubricants containing solid contaminants.
3. Evaluate different wear measurement techniques for their suitability to
identifying a methodology for characterising the antiwear additives.
182
4. Study the effect of micro-geometry on the tribological performance, by
measuring the change in minimum oil film thickness.
5. Characterise antiwear additives using a unique number, and rank them for
their efficacy.
6.2 Literature Review
This research required an in-depth knowledge and literature search in the following
main areas:
• Contaminants’ effect on wear
• Effect of solid contaminants on journal bearings
• Tribological performance and criteria for its measurement
• Micro-geometry parameters and their effect on bearing lubrication
• Antiwear additives and their characterisation
The literature review revealed the following information on different aspects of the
problem that was important for this research project:
Contaminants:
The literature revealed the information about various aspects of contaminants such
as: sources of contaminants, latest ISO 4406 cleanliness requirements, different
multi-body wear mechanisms and modes, advances in micro-polar effects in
lubrication, and information related failures due to contaminants.
‘K’ Ratio:
The literature highlighted the detrimental effects of ‘K’ ratio and its relationship with
wear, friction and embedding of the particles on the journal bearing surfaces.
Maximum wear occurs for K=1, and hence, the effect of antiwear additives has been
studied for this condition.
Hardness Ratio:
Literature on interaction of contaminants with the bearing surfaces gave the insight
183
that if the ratio of contaminant hardness to bearing surface hardness is close to three,
it leads to high wear.
Micro-geometry:
The literature research in the area of micro-geometry showed substantial amount of
existing research on roughness effects in both transverse and circumferential
directions; however, this research does not appear to be conclusive, consequently,
‘smooth-is-best’ concept is still widely accepted.
Only limited amount of research has been carried out to study the effect of out-of-
roundness and radial clearance of journal bearings. Out-of-roundness has been
accepted by some researchers as a valid wear measure parameter, however, it has not
been widely used for bearing specification. This research demonstrates that out-of-
roundness measurements are as important as roughness measurements.
Radial clearance has direct impact on the load bearing capacity of a journal bearing;
therefore, this research investigated its impact on oil film thickness.
Measurement techniques:
Meta-research exposed the state-of-art in measurement of oil film thickness and wear
measurement techniques. However, no study highlighted the preference of one
measurement method over the others. Therefore, this research has compared the
suitability of various wear measurement techniques, to select the most suitable one
for characterising antiwear additives.
Antiwear additives:
Literature on antiwear additives discussed mainly their chemistry and applications.
Little information is available on the performance of antiwear additives when used in
bearings containing solid contaminants. This research is thus focused on filling this
knowledge gap.
Characterisation:
Literature search on characterisation of antiwear additives revealed the limited
amount of past research on this topic. The literature found on this topic focused on
characterisation of additives with clean oils in Elastohydrodynamic or concentrated
184
contact regimes. A study on the journal bearing wear with oil containing antiwear
additive was also found; however, this study discussed the effect of antiwear
additives on the wear modes and mechanisms only. Therefore, this research included
the development of a model to characterise antiwear additives, particularly for dusty
journal bearings.
Other important areas:
Literature search in allied areas revealed useful information on topics such as:
theoretical modelling of wear performance in worn journal bearings, vibration
monitoring of contaminated bearings, ISO cleanliness and filtration requirements,
and micro-grooved two-component surface layers. Literature review in these areas
helped in identifying appropriate methodologies and techniques used in this research,
and ensuring that this project builds on existing knowledge.
Knowledge Gaps
The literature review revealed the following knowledge gaps in areas relevant to this
research project:
• The effect of solid contaminants treated with antiwear additives on journal
bearing wear has not been fully studied.
• Characterisation of antiwear additives based on their efficacy for dusty
applications under hydrodynamic lubrication has not been carried out.
• The effect of solid contaminants on the bearing micro-geometry and its effect on
the bearing’s tribological performance are not well understood.
• There is no standard numerical parameter for classifying the performance of
antiwear additives operating in dusty hydrodynamic lubrication conditions.
6.3 Experiment Design and Development
Experiment design comprised the road map for the experiment setup and procedures
used in this project, and this led to the development of new practical and theoretical
methodologies for improving bearing design and metrology.
185
Test procedures
The experiments were designed for best utilisation of the available resources, which
led to the flowing strategic decisions:
• Conduct short duration test without repeating them.
• Compare the wear measurements obtained from different methods to compare
their accuracy, and select the most reliable method to obtain reliable results.
• Keep testing environment and procedures consistent, because the tests are not
to be repeated.
• Keep K=1, sliding distance = 7536m, and all other operating and
environmental parameters the same.
• Develop a performance parameter selection process based on weight loss,
micro-geometry and particle counts. This process was called Multi Wear
Parameter Approach (MWPA).
Preliminary problems and solutions:
The aims and objectives of this study demanded that micro-geometry parameters are
carefully monitored. This required preliminary measurements of radial clearance,
roughness and roundness measurements prior to wear tests. Further analysis of these
measurements lead to the identification of the following problems, and their
solutions:
• Radial clearance varied from one location to another all along the
circumference of the bearing, mainly due to out-of-roundness; as a result, a
new heuristic was developed: that out-of-roundness must be specified along
with the radial clearance, just as cut-off length is specified with the surface
roughness.
• In this preliminary study, it was observed that roughness values are lower
than the out-of-roundness values, this predicated that the film parameter (λ=
10) does not ensure adequate separation of lubricated surfaces and need
186
another design parameter. A new design parameter called Film Shape Factor
(FSF, or gamma ratio) has been defined.
• Film thickness measurement technique using proximity probes require
verification based on theoretical models. This led to the development of a
FORTRAN based program for predicting the bearing performance.
6.4 Results and Analysis
The analysis of results was aimed to determine the effect of each antiwear additive
on the journal bearing. A set of 14 wear measure parameters were used as a part of
the MWPA methodology to identify the most suitable method for characterising the
additives. The following conclusions were drawn from this analysis:
Efficacy of antiwear additives:
The antiwear additive based on sulphur / phosphorous chemistry used in Test A6 has
maximum wear-reducing effect. Whereas, the additive used in Test A4 has almost no
antiwear effect. This fact was supported by seven wear measure parameters.
Weight loss:
Even though weight loss is a direct method for wear measurement, with microscopic
wear, it did not prove to be reliable.
Radial clearance:
There were two reasons for not accepting radial clearance as the best candidate for
measuring wear. First, the radial clearance may change due to change in bearing ID
as well as shaft sleeve OD, and it is difficult to isolate the contribution of each
bearing element. Second, the radial clearance changes within a single rotation from
one location to another. Hence, it is not accurate to assign a single value to this
parameter.
Surface roughness:
There were two problems with roughness measurements, and hence, these could not
187
be used for characterising the additives:
• The roughness on both the bearing elements varies randomly (in either
direction), hence, neither the theory of preferential path nor the theoretical
predictions made by other researchers could be confirmed.
• Roughness values vary drastically from one location to another within the
wear zone, hence there is a subjectivity involved in recording the data.
Out-of- roundness:
Out-of-roundness proved to the best micro-geometry parameter; it gave the most
reliable results, and hence was chosen for the characterisation of antiwear additives.
This methodology magnified the departure of circumference from the perfect circle
several thousand times, without magnifying the overall diameter of the bearing,
leading to highly accurate results.
Particle counts and debris weight:
Particle count is a widely used condition monitoring technique; however, due to
heavy contamination, repeatability of test results was poor, and hence was not
suitable. Change in total wear debris weight was not used as a wear measurement
parameter because of inconsistencies in results obtained for the same.
Minimum oil film thickness:
Minimum oil film thickness measurements were reliable when pure base oil results
were compared with the predicted values obtained from the FORTRAN program, or
the on-line ESDU program. However, as the tests proceeded, the contaminant
congestion caused severe fluctuations in readings, thus the results were not reliable.
However, observations demonstrated that K = 1 condition (i.e. contaminate size =
film thickness) causes sever congestion in the bearing contacts.
6.5 Characterisation of Antiwear Additives
The salient features of the method developed for characterising the antiwear
additives, using the out-of-roundness traces, are as follows:
188
• The out-of-roundness trace methodology provides a systematic and step-by-
step process for characterising antiwar additives.
• It calculates the area of wear on the out-of-roundness trace at a chosen
bearing cross-section (CSWA). The weight loss calculated from CSWA is
less error prone than physical measurement of these very small quantities.
• The wear depths measured at different angular location along the bearing
circumference give the true geometry of the worn area, and so is the case with
the maximum wear depth.
• Wear characteristic equations developed from the wear depth data give
unique wear performance signature for the antiwear additives. However,
these equations cannot quantify the wear.
• Estimated wear loss from the out-of-roundness traces was used for deriving a
wear characteristic number (N). This number can be used for selecting the
most appropriate antiwear additive for a given dusty application.
6.6 Research Contributions
This research has made the following main contributions to the existing body of
knowledge in this domain:
• This study is first of its own kind where effect of solid contaminates treated
with antiwear additives have been experimentally studied on a journal
bearing.
• A unique method has been developed for computing the microscopic weight
loss in bearings with high precision and reliability using the out-of-roundness
traces.
• A geometrical method has been developed by which oil film thickness can be
measured with higher precision in a bearing where proximity probes are
mounted on a floating bearing housing.
• It is demonstrated that the radius of the bearing changes with the location
189
along the circumference, which is mainly due to out-of-roundness, and hence
out-of-roundness should be specified along with the radial clearance –just as
the cut-of length is specified along with surface roughness.
• A seminal concept have been proposed in which the Film Shape Factor (FSF
or gamma ratio) is used for calculating minimum oil film thickness in journal
bearings whose out-of-roundness values are higher than roughness values.
• Wear characteristics equations have been developed with the help of out-of-
roundness traces. These act as signatures of antiwear additive’s wear
behaviour.
• A method for finding Wear Characteristic Number (WCN, or N) has been
developed, and different additives can be ranked for their efficacy based on
this number. Therefore, the users can select the most suitable additive for
their dusty applications by using this number.
190
BLANK PAGE
191
CHAPTER 7
7. SCOPE FOR FUTURE WORK
This research has investigated the effect of antiwear additives on bearings lubricated
with oils containing solid contaminants. Though the main objectives of this project
have been achieved, this research can be advanced in the following directions:
Metrology of the bearings: This research revealed that there is need to study the
effect of the metrology of the bearing under dusty environments. Effects of changes
in oil film thickness due to change in instantaneous radial clearance of a running
bearing need to be studied in more detail, so that its influence on bearing
performance can be better understood.
Film Shape Factor: The proposed Film Shape Factor (γ ratio) is a seminal model,
and needs further research to regiourously test its validity and apply it more widely.
Furthermore, this model can be extended to derived a similar factor for
hydrodynamic thrust bearings – by treating the surface waviness as out-of-
roundness.
Method for finding CSWA: A method for finding CSWA with higher precision
needs to be developed, such as, considering more traces along the bearing length.
Theoretical modelling of worn bearings: A theoretical model should be developed
for predicting minimum oil film thickness in a dynamic system with radial clearance
as a time variant. Such a model would be helpful in developing an expert system for
condition monitoring of machines operating in dusty environments.
Testing of more antiwear additives: A wider variety of antiwear additives should
be tested to characterise these for the benefit of industrial users.
A wider variety of antiwear additives should be tested to characterise them for the
benefit of industrial users.
192
Influence of some other parameters: The bearing operating parameters such as ‘K’
ratios, bearing clearances, temperature rise, types of contaminants and their
concentration need to be varied and their effect on bearing wear and tribological
performance be studied in more detail.
A study of reduction in friction due to antiwear additives needs to be pursued, with
regards to energy saving in dusty applications.
193
8.0 REFERENCES
1. Allen, S. J., Kline, K. A., (1971), "Lubrication Theory for Micro-polar Fluids", Transactions of the ASME Journal of Lubrication Technology, Vol. 9, pp. 646-650.
2. Archard, J. F., (1953), "Contact and Rubbing of Flat Surfaces."Journal of Applied Physics. 24: 981-988.
3. ASTM-D2266, (2004), “Standard Test Method for Wear Preventive Characteristics of Lubricating Grease (Four-ball method)”, Annual book of ASTM Standard 2004, V 05.01, pp. 287-289.
4. ASTM-D2596 (2004), “Standard Test Method for Measurement of Extreme Pressure Properties of Lubricating Grease (Four ball method)”, ASTM Standard 2004, Vol.05.01, pp. 966-970.
5. Bagnel, A., (1978), “Talyrond Operating Manual, Ed., T. Hobson”, Taylor Hobson Pty. Ltd., Leicester, England: pp. 1-85.
6. Bayer, Raymond George ( 2004), “Mechanical Wear’, CRC Press, ISBN 0824746201, pp 229-230.
7. Beghini, E., et. al., (1992), "Elasto/plastic Contact and Endurance Life Prediction.", Journal of Physics, vol.(D), pp. 379-383.
8. Bell, J. C., et al., (1992), “The Removal of Substrate Material Through Thick Zinc Dithiophosphate”, Wear of Metals, Elsevier Science Publishers, Leeds.
9. Bell, M. E. and Findlay, J. H., (1941), "Molybdenum as a New Lubricant", Physics Review, pp. 59: 922.
10. Berthier, Y. and Godet, M., (1989). "Velocity Accommodation in Friction.", Tribology Transaction,. 32(4), pp. 490-496.
11. Blau, P. J., et. al., (1997), "Birth and history of the International Conferences on Wear of Materials", Wear, issue : March 10, pp. 203-204.
12. Booser, E. R., (Ed.), (1983). “Handbook of Lubrication; Theory and Practice of Tribology”, CRC Press Inc., Florida, USA,
13. Bourn, J., The Comptroller and Auditor General, British Ministry of Defence, (2001), “Exercise Saif Sareea II, Report no. HC 1097 session 2001-2002, 1 August 2002. Viewed 20 October 2002. http://www.nao.org.uk/ publications/nao_reports/01-02/01021097.pdf
14. Broeder, J. J. and Heijnekamp, J. W., (1965), "Abrasive Wear of Journal Bearings by Particles in the Oil (Apparatus, Experiments, and Observations)", Proc. Of Institution of Mechanicam Enginers- 180, part 3K, pp. 21-31.
15. Chandrasekaran, R., et. al., (1985). "Effect of Abrasive Contaminants on Scuffing", Tribology International, 18(August), pp. 219-222.
16. Christensen, H., (1970), "Stochastic Models for Hydrodynamic Lubrication
194
of Rough Surfaces", Proceedings of the Institution of Mechanical Engineers-184, Part 1(55), pp. 1013-1022.
17. Chu, P. S. and Kay, E., (1974), "Optimum Clearance Fits for Journal Bearings in Relation to to lubrication theory", Wear, 27(3), pp. 329-343.
18. Czichos, H., (1978), “Tribology -A System Approach to the Science and Technology of Friction Lubrication and Wear”, Elsevier Scientific Publication company, New York.
19. Dai, F. and Khonsari, M. M., (1992), "Analytical Solution for Mixture of a Newtonian Fluid and Granules in Hydrodynamic Bearings", Wear, V. 156, pp. 327-344.
20. Das, S. et. al., (2002), "On the Steady State Performance of Misaligned Hydrodynamic Journal Bearings Lubricated with Micro-polar Fluids", Tribology International, 35(4), pp. 201-210.
21. Das, S., et. Al., (2004), "Linear Stability Analysis of Hydrodynamic Journal Bearings under Micro-polar Lubrication", Tribology International, 38(5), pp. 500-507.
22. Din, M. D. and Kassfeldt (1999), "Wear Characteristics with Mixed Lubrication Condition in Full Scale Journal Bearing", Wear, 232: 192-198.
23. Dong, J. X. and Z. S. Hu (1998). "A Study of the Anti-Wear and Friction-Reducing Properties of the Lubricant Additive: Zinc Borate", Tribology International, 31(5), pp. 219-223.
24. Dorinson, A. and Ludema, K.C., (1985), “Mechanics and Chemistry in Lubrication”, Elsevier science Publishers, Amsterdam.
25. Douglas, G., (1989), "Clean Dry Oil Prolongs Life of Lubricated Machines", Lubrication Engineering, (January):Vol. 4-8.
26. Duchowski, J. K., (1998). "Examination of Journal Bearing Filtration Requirements", Lubrication Engineering, (September), pp. 18-28.
27. Duckworth, W. E., et.al., (1957), “Wear of Lubricated journal bearing. Lubrication and Wear”, Proceedings of Institution of Mechanical Engineers, London, pp. 86-92.
28. Dufrane, K. F. and Kannel, J.W., (1989). "Thermally Induced Seizures of Journal Bearing." Transactions of the ASME, Journal of Tribology. 111(April): 288-292.
29. Dufrane, et.al., (1983), "Wear of Steam Turbine Journal Bearing at Low Operating Speeds", Journal of Lubrication Technology, 105, pp. 313-317.
30. Dwyer- Joyce, R. S., (1993), “The Effect of Lubricant Contamination on Rolling Bearing Performance”, Ph.D. Thesis, Mechanical Engineering, Imperial College of Science Technology and Medicine, London, pp. 238.
31. Dwyer-Joyce, R. S., et al., Ed., (1990), “Surface Damage Effects Caused by Debris in Rolling Bearing Lubricants with a Particular Emphasis on Friable Debris Materials, Rolling Element Bearings -Towards the 21 st Century”, Mechanical Engineering Publications for the I.Mech.E.
32. Dwyer-Joyce, Ed. (1992), “Lubricant Screening for Debris Effects to
195
Improve Fatigue and Wear Life”, Wear Particles, Elsevier Science Publishers, Amsterdam.
33. Elwell, R. C., (1977), "Foreign Object Damage in Journal Bearing", Lubrication Engineering, Vol. 34, pp.187-192.
34. Eringer, A., (1964), "Simple Micro-fluids", International Journal of Engineering Science, Vol. 2, pp. 205-217.
35. ESDU-84031, (1996) “Journal Bearing Calculations”, Vol.2, Issue6, Engineering Science Data Unit International Plc, Essex, U.K., pp. 1-46.
36. Fang, L., et al, (1993), "Movement Pattern of Abrasive Particles in Three-body Abrasion", Wear, Vol. 162-164, pp. 782-789.
37. Fang, L., et al, (1992), "A Wear Tester Capable of Monitoring and Evaluating the Movement Pattern of Abrasive Particles in Three-Body Abrasion", Wear, Vol. 159, pp. 115-120.
38. Fang, L., et al, (1991), "An Explanation of the Relation between Wear and Material Hardness in Three-Body Abrasion", Wear, Vol. 151, pp. 313-321.
39. Feng, L., et al, (1995), “Synthesis and Hydrolytic Stability of Aquec Antiwear Agents of Organic Phosphate and Organic Thiophosphate”, Journal of Synthetic Lubrication, Vol. 14, Issue 3,pp. 259-266.
40. Fillon, M. and. Boyer, J., (2004). "Thermodynamic Analysis of a Worn Plain Journal Bearing." Tribology International, 37(2), pp. 129-136.
41. Fitch, J. C, (1991), "Contaminant Monitoring Targets Root Causes of Machinery Problems", P/PM Technology, Issue March/April, pp. 34-35.
42. Fodor, J., (1980), “Energy Savings in I.C. Engines through Controlling the Contaminants”, Research Institute of Automotive Industry, Budapest, Hungary,pp. 101-118.
43. Fodor, J., (1987), “Wear and Friction of Hydrodynamically Lubricated Bearings in Abrasive Environment”, Tribotechnica-5th Conference on Friction Lubrication and Wear, Bucharest. pp. 121-124.
44. Fodor , J. and Ling, F.F. (1987). "Friction Reduction in I.C. Engine Through Improved Filtration and new Lubricant Additive?" Lubrication Engineering. 614-617.
45. Forbes, E. S. and Battersby, J., (1974) "The Effect of Chemical Structure on the Load Carrying and Adsorption Properties of Dialkyl Phosphites", ASLE Trans., Vol. 17, pp. 263-269.
46. Forrester, P. B., (1960), “Bearing and Journal Wear”, Symposium on Wear in the Gasoline Engine, Thornton Research Centre, U.K. pp. 75-91.
47. Frith, R. H. and W. Scott (1993), "Control of Solid Contamination in Hydraulic Systems- An Overview", Wear, Vo. 165, pp. 69-74.
48. Katzenmeier, G. T., (1972), “The Influence of Materials and Surface Quality on Wear Behaviour and Load Capacity of Journal Bearing”, Proc. of The Tribology Convention, Inst. Mech. Eng., U.K.
49. Glaeser, W. A., (1992), “Ball Mill Simulation of Wear Debris Attrition”,
196
Wear Particles, Elsevier Science Publishers B.V.
50. Godfrey, D., (1989), "Clean Dry , Oil Prolongs Life of Lubricated Machines", Lubrication Engineering, Issue: January, pp. 4-8.
51. Gradner, W. W., (1978), "Journal Bearing Operation at Low Sommerfeld numbers", ASL:E Transactions, 19(3), pp. 187-194.
52. Greenwood, J. A, (1970), "Re-examination of EHL film-thickness Measurements", Wear. 15(3): 281-289.
53. Groszek, A. J., (1962), "Heat of Preferential Adsorption of Surfactants on Porous Solid and its Relation to Wear of Sliding Steel Surfaces", ASLE Trans. 5, pp. 105-113.
54. Halling, J., (1975), Principles of Tribology, The Macmillian Press Limited.
55. Hamer, J. C., et al., (1989), "Particle Deformation and Counter face Damage When Relatively Soft Particles are Squashed Between Hard Anvils", Tribology Transactions, 32(3),pp. 281-288.
56. Hashimoto, H., et. al., (1986), "Performance Characteristics of Worn Journal Bearings in both Laminar and Turbulent Regimes Part-I: Steady State Characteristics", ASLE Transactions, 29(4), pp. 565-571.
57. Hirstch, P., Scott, W. and Kirkcaldie, R. K., (1980), “Wear of Journal Bearing with Abrasive Containing Lubricants”,. Proceedings of National Conference on Lubrication, Friction and Wear in Engineering, Melbourne, IEAust, 1-5 December, Publication No. 80/12, pp. 14-18.
58. ISO 4406 (1987), “ Hydraulic Fluid Power Fluids-Method for Coding Levels of Contamination by Solid Particles”, International Standards Organisation.
59. Jahanmir, S., (1986), "Wear Reduction and Surface Layer Formation by a ZDDP Additive", ASME. 86-Trib-46: 10.
60. Jahanmir, S. (1987). "Wear Reduction and Surface Layer Formation by ZDDP Additive." ASME Journal of Tribology. 109(October): 577-586.
61. Jiusheng, L., et.al., (2003), "Influence of Base oil Properties on the Efficiency of Action of ZDDP-Type Additives", Journal of Synthetic Lubricants, 20(2), pp. 151-158.
62. Kano, M. Y., Yoshiteru, Y and Jiping, E., (2003), "Friction Properties of Tribofilm Formed from Engine Oil Additives (Part 1) - The Effect of DLC Coated Material on Forming Ttribo-films", Toraibarojisuto/Journal of Japanese Society of Tribologists. 48(1),pp. 54-59.
63. Kapsa, P. M., et.al., (1981), "Antiwear Mechanism of ZDDP in the Presence of calcium Sulfonate Detergent", Journal of Lubrication Technology, 103(4),pp. 486-496.
64. Kawamura, M. and Fujita, K., (1983), "Antiwear Property of Lubricant Additives for High Silicon Alloy under Boundary Lubricating Conditions", Wear, Vol. 89, pp. 99-105.
65. Khonsari, M. M. and Brewe, D. E., (1989), "On the Performance of Finite Journal Bearing Lubricated with Micropolar Fluids", Tribology
197
Transactions, 32(2), pp. 155-160.
66. Khonsari, M. M. and Kim, H. J., (1989), "On Thermally Induced Seizure in Journal Bearings", Transaction of ASME, Journal of Tribology. 111, pp. 661-667.
67. Khonsari, M. M., et al., (1999), "On the Scuffing Failure of Hydrodynamic Bearings in the Presence of an Abrasive Contaminant", Trans. of ASME, Journal of Tribology, 121(January),pp. 90-96.
68. Khonsari, M. M. and Wang, S. H., (1990), "On the Role of Particulate Contamination in Scuffing Failure", Wear,.Vol. 137,pp. 51-62.
69. Khorramian, B. A., et al., (1993), "Review of Antiwear Additive for Crank Case Oil", Wear, Vol. 169, pp. 87-95.
70. Klamann, D. P., (1985), "Effect of Lubricating Oil Parameters on the Automobile Engines Fuel Consumption. Part 1: CEC-Testing Method, Viscosity and Additives Effects.)", Erdoel und Kohle, Erdgas, Petrochemie Vereinigt mit Brennstoff-Chemie, Vol. 38(9), pp 393-399.
71. Klaus, E. E. and Bieber, H.E., (1965), "Effect on P32 Impurities on the Behaviour of Tricresyl Phosphate-32 as an Antiwear Additive", ASLE Trans., Vol.,8, pp. 12-20.
72. Kragelsky, I. V., et.al., (1982), “Friction and Wear : Calculation Methods”; Translated from Russian by N. Standen., Pergamon Press, Oxford.
73. Kuhnell, B. T., (1992).,"Numerical Computer Image Analysis of Wear particles for Machine Condition Monitoring",."The Bulettin of CMCM Monash University, Vol.-4, No.1, pp.130-135.
74. Kulczycki, A. K., (1994), "Influence of Base Oil Properties on the Efficiency of ZDDP-Type Additive", Lubrication Science, 6(2), pp. 161-179.
75. Kumar, A., and Mishra, S.S. (1996). "Steady state analysis of noncircular worn journal bearings in nonlaminar lubrication regimes "Tribology International. 29(6): 439-498.
76. Lansdown, A. R., (1982), “Lubrication, a practical guide to lubricant selection”, Pergamon Press, Oxford.
77. Lansdown, A. R., (1999), “Molybdenum Disulphide Lubrication”,. Elsivier Science Publishing V.B., Amsterdam.
78. Larsen-Basse, .S., (1975), "Influence of Atmospheric Humidity on Abrasive Wear EM Dash 2.2-Body Abrasion", Wear, 31(1), pp. 373-379.
79. Liston, T. V., (1992), “Title of Paper”, Japanese Society of Lubrication Engineers, Issue May, pp. 389-395.
80. Mahanti, A. C., (1976), "A Theoretical Study of the Effect of Solid Particles in the Lubricant of a Partial Journal Bearing", Wear, Vol. 39, pp. 45-53.
81. Marina, F. Y., et. Al., (1997). "Chemical Characterisation of Tribochemical and Thermal Films Generated from Neutral and Basic ZDDPs Using X-ray Absorption Spectroscopy", Tribology International, Vol. 30(4), pp. 305-315.
198
82. Martin, N. J, (1991), “Wear of Journal Bearing with Solid Particle Contaminated Lubricants”, Masters thesis, Mechanical and Manufacturing Engineering Department, Queensland University of Technology, Brisbane, pp. 280.
83. Maru, M., et.al. (2007), “Study of Solid Contamination in Ball Bearings through Vibration and Wear Analysis”, Tribology International, Vol. 40, Issue 3, pp 433-440.
84. McKee, S. A., (1927), "Effect of Abrasive in Lubricant", SAE Transactions pp. 73-77.
85. Mehan, R. L., (1988), "The Wear of Selected Materials in Mineral Oil Containing a solid Contaminant", Wear, Vol. 124, pp. 65-85.
86. Misra, A. and Finnie, I., (1980), "A Classification of Three-body Wear and Design of a New Tester", Wear, Vol. 60(1), pp. 111-121.
87. Misra, A. and Finnie, I., (1981a), “Correlation between Two-body and Three-body Abrasion and Erosion of Metals”, Wear, Vol. 68, No. 1, pp. 33-39.
88. Misra, A. and Finnie, I., (1981b), "On the Size Effect in Abrasive and Erosive Wear", Wear, Vol. 65, pp. 359-373.
89. Misra, A. and Finnie, I., (1982), "A Review of the Abrasive Wear of Metals", Transaction of ASME. 94(104), pp. 94-101.
90. Misra, A. and Finnie, I., (1983), "An Experimental Study of Three-Body Abrasive Wear", Wear, Vol. 85(1), pp. 33-39.
91. Mitsuya, Y., T., et al., (1989), "Average Reynolds Equation Extended to Gas Lubrication Processing Surface Roughness in Slip Flow Regime: Approximate Method and Confirmation Experiments", Journal of Tribology, ASME Transactions, Vol. 111(3),pp. 495-503.
92. Mitsuya, Y. and Ota, H., (1991), "Stiffness and Damping of Compressible Lubricating Films Between Computer Flying Heads and Texture Media: Perturbation Analysis Using the Finitie Element Method", Journal of Tribology, ASME Transactions, Vol.196(4), pp. 554-560.
93. Mizuhara, K., M., et al., (2000), "Effect of Particle on Lubricated Friction", STLE Tribology Transactions, Vol. 43(1), pp. 51-56.
94. Mokhtar, M. O., et. al., (1978)), "Wear Characteristics of Plain Hydrodynamic Journal Bearing During Repeated Starting and Stopping", ASLE Trans., Vol. 20(3), pp. 191-194.
95. Moon, M., (2007), “How Clean are Your lubricants ?”, Trends in Food Science and Technology, Vol. 18, Suppliment 1, January 2007, pp. s74-79.
96. Narayanan, R. N., Nair, C. C. and Prabhakaran, K., (1995), "Analysis of Mass Transfer Effect on Performance of Journal Bearing Using Micro-polar Lubricants", Heat and Mass Transfer, Vol. 30(6), pp. 429-433.
97. Odi-Owei, S. and Roylance, B., J., (1987), "Lubricated Three Body Abrasive Wear- Contamination Condition Verseus Bounding Surface Material Hardness", Tribology International, Vol. 20(Feb.), pp. 32-40.
199
98. Offley, E., (2003), “Wrong Lubricant, Jammed Weapons, dead Soldiers”, 18 June, News.Defence Watch: The voice of the Grunt, 23 March 03: http://www.militec-1.com/articles/SFTT.html
99. Pai, R. and. Mazumdar., B.C., (1992),. "Stability of Submerged Four-lobe Oil Journal Bearings Under Dynamic Load", Wear, Vol. 154(1), pp. 95-108.
100. Patir, N. and Cheng , H. S., (1978), “An Average Flow Model for Determining Effects of Three- Dimensional Roughness on Partial Hydrodynamic Lubrication", Transaction of ASME, Vol. 100(January), pp. 12-18.
101. Peng, Z., Kessissoglou N.J., et.al. (2005), “ Study of the Effect of Contaminant Particles in Lubricants Using Wear Debris and Vibrationcondition Monitoring Techniques”, Wear, Volume 258, Issue 11-12, pp.1651-1662.
102. Prakash, B. and Kumar, B., (1987). “Influence of Contaminant Particle Size on Wear in Journal Bearing Under Boundary Lubrication Conditions”, International Tribology Conference, Melbourne, pp. 33-38.
103. Peterson, M. B., Winer, W.O., Ed., (1980), “Wear Control Hand Book”.,
American Society of Mechanical Engineers, New York. 104. Prakash, J. and Sinha, P., (1977), "Theoretical Effects of Solid Particles on
the Lubrication of Journal Bearings Considering Cavitation", Wear, Vol. 41, pp. 233-249.
105. Prakash, J. and. Sinha, P., (1975), "Lubrication Theory for Micro-polar
Fluids and its Application to a Journal Bearing", Int. Journal of Engineering Sciences, Vol. 13, pp. 217-232.
106. Rabinowicz, E., (1965), “Friction and Wear of Materials”, John Wiley and
Sons, New York. 107. Rabinowicz, E., et al., (1961), "A Study of Abrasive Wear Under Three-
body Conditions", Wear, Vol. 4, pp. 345-355. 108. Rigney, D. A., (1994), "The Role of Hardness in the Sliding Behaviour of
Materials", Wear, Vol. 175, pp. 63-69. 109. Roach, A. E. and D. Mich (1950), “Performance of Oil-Film Bearings with
Abrasive-Containing Lubricant”,. Annual ASME Meeting on Lubrication and Machine Design, ASME, New York.
110. Ronen, A. and Malkin, S., (1981), "Wear Mechanism of Statically Loaded
Hydrodynamic Bearing by Contaminant Abrasive Particles", Wear, Vol. 68, pp. 371-389.
111. Ronen, A. and. Malkin, S., (1983), "Investigation of Friction and wear of Dynamically Loaded Hydrodynamic Bearings with Abrasive Contaminants", ASME Transactions, Journal of Lubrication Technology, Vol. 105, pp. 559-568.
200
112. Ronen, A, et al., (1980), "Wear of Dynamically Loaded Hydrodynamic
Bearings by Contaminant Particles", Transaction. of ASME, Vol. 102(October), pp. 452-458.
113. Rowe, C. N. (1967), "A Relation Between Adhesive Wear and Heat of
Adsorption in the Function of Boundary Lubricants", ASLE Transactions, Vol. 9, pp. 110-111.
114. Rowe, C. N., (1970), "Role of Additive Adsorption in the Mitigation of
Wear", ASLE Trans., Vol. 13, pp. 179-188. 115. Rowe, C.N., (1980), “ Lubricated Wear”, “ in wear Control Handbook,
Peterson M.B and Winer, W.O, Ed., 93-102.. American Society of Mechanical Engineers, New York (1980), pp 93-102.
116. Rylander, H. G., (1952), “Effects of Solid Inclusions in Sleeve-Bearing Oil
Supply. Semi-Annual Meeting of the ASME,Ohio, USA. 117. Scharrer, J. K., et. al., (1991), "The Eeffects of Wear on the Rotor dynamic
Coefficients of a Hydrostatic Journal Bearing", ASME Journal of Tribology, Vol. 113, pp. 210-213.
118. Scott, W. and D. Hargreaves, (1991), “A Case of Reclaimed Lubricating
Oils", Proceedings of the International. Mechanical Engineering Congress- Mech'91', Inst. of Engineers Australia, pp.1-4.
119. Sep, J. and. Kucaba-Pietal A., (2001), "Experimental Testing of Journal
Bearings with Two-Component Surface Layer in the Presence of an Oil Abrasive Contaminant", Wear, Vol. 249(12), pp. 1090-1095.
120. Sep, J.,(2004), “Three-dimensional Hydrodynamic Analysis of Journal
Bearing with a Two-component Surface Layer”, Tribology International, Vol. 38, pp. 97-104
121. Sharma,S.C., K.L. Awasthy and B.S. Tak, “ Experimental Studies of Wear
in Metal Working Conditions ”, Proc. Fourth International Tribology Conference, UWA, Perth, Australia, 5- 8 Dec. 1994 ed. G.W. Stachowiak, pp. 825 – 828.
122. Sharma, S., C. and Hargreaves, D (2001), “Effect of Solid Contaminants on
Journal Bearing Performance”, World Tribology Conference, Vienna, pp.1-4
123. Sharma, S. C., Hargreaves, D. and Scott, W., (2004), Influence of Errors in
Measuring the Radial Clearance of Journal Bearing Performance. 1st International Conference on Advanced Tribology, Singapore. pp.1-3
. 124. Sinha, P., Singh, C. and Prasad, K.R., (1981), "Effect of Viscosity Variation
Due to Lubricant Additives in Journal Bearings", Wear, Vol. 66, pp. 175-
201
188. 125. Tallian, T. E., et.al., (1964), "Lubricant film in Rolling Contact of Rough
Surfaces", American Society of Lubrication Engineers (ASLE), Vol. 7(2), pp. 109-126.
126. Tanaka, K. S. and Musuzuki, A., (1994), “Development of Gear Lubricant
for High Speed Shinkansen Electric Rail-Cars- A Study of New Thermal-Resistant Gear Oils for 300 km/h Class Shinkansen Ttrains”, Quarterley Report of Railway Technical Research Institute, Japan, V35, No.4, Nov. 1994, pp. 224-228.
127. Tao, F. F. and Appeldoorn, J. K., (1969),"An Experiment Study of the Wear
Caused by Loose Abrasive Particles in Oil", ASLE Transactions, Vol. 13(169), pp. 169-178.
128. Thompson, R. A. and Bocchi, W., (1971), “Model of Asperity Load Sharing
in Lubricated Contacts”, Lubrication Conference ASLE/ASME, Oct.5-7, 1971, Pittsburgh USA, pp. 67-69.
129. Tipei, N., (1979), "Lubrication with Micro-polar Fluids and its Application
to Short Bearings", Journal of Lubrication Technology, Vol. 101, pp. 350-363.
130. Tonder, K., (1986), "The Lubrication of Unidirectional Straighted
Roughness Consequences for Some General Roughness Theories", Journal of Tribology Transactions of ASME, Vol 108, pp. 167-170.
131. Trott, M., (2004), “The Mathematica GuideBook for Programming”,
Springer Verlag, New York, http://www.mathematicaguidebooks.org. 132. Truscott, G. F., (1972), "A Literature Survey on Abrasive Wear in Hydraulic
Machinery", Wear, Vol. 20, pp. 29-50. 133. Tsann-Rong, L., (1996), "Hydrodynamic Lubrication of Journal Bearings
Including Micro-polar Lubricants and Three-dimensional Irregularities", Wear, Vol. 192, pp. 21-28.
134. Tzeng, S. T. and E. Saibel, (1967), "Surface Roughness Effect on Slider
Bearing Lubrication", ASLE Transactions, Vol. 10, pp. 334. 135. Vaidyanathan, K., and Keith, T.G., (1991), "Performance Characteristics of
Cavitated Noncircular Journal Bearings in the Turbulent Flow Regime", Tribology Transactions, Vol. 34, pp. 35-44.
136. Warne, T. M. and Clive, H. A., (1985), "Toxicity of Lubricating
Oils."ASLE Preprints, No. 85-AM-31-1(1), pp 1-9. 137. Watanabe, S., et al., (1985), “Evaluation of Wear Life of Journal Bearing
Lubricated by Contaminated Oils”, Proceedings of JSLE Int'l Tribology
202
Conference, Tokyo, Japan. pp 114-118. 138. Weisstein, E. W., (2004), "Simpson's Rule From Math World-A Wolfram
Web Resource", http://mathworld.wolfram.com/SimpsonRule.html. 139. William, J. A. and Hyncica, A. M., (1992), "Mechanism of Abrasive Wear
in Lubricated Contact", Wear, Vol. 152, pp. 57-75. 140. Winer, W. O., (1967), "Molybdenum disulphide as a Lubricant: A review of
the Fundamental Knowledge", Wear, Vol. 10, pp. 422. 141. Wlkstrom, V., et al., (1993), "Wear of Bearing Liners Low Speed Rotation
of Shafts with Contaminated Oil", Wear, Vol. 162-164, pp. 996-1001. 142. Xuan, J. L., al., (1989), "Hardness Effect on Three-Body Abrasive Wear
under Fluid Film Lubrication", Journal of Tribology, Vol. 111(January), pp. 35-40.
143. Yuan, C. Q., et.al. (2004), “Effects of Temperature on Sliding Wear Process
under Contaminated Lubricant Conditions”, Wear, Vol. 257, Issue 7-8, pp 812-822.
144. Yuji, Y. and Gondo, S., (1989), “Friction and wear characteristics of
molybdenum dithiocarbamate and molybdenum dithiophosphate”, Tribology Transactions, Vol. 32, pp. 251-257.
145. Zaheeruddin, K. and. Isa, M., (1978), "Micro-polar Fluid Lubrication of
One Dimensional Journal Bearing", Wear, Vol. 50, pp. 211-220. 146. Zheng, P. Y., et. al., (1986), “The Mechanism of Friction Reduction of
Sulfurised Oxymolybdenum dichloroethylhexyl-phosphorodiothate under Boundary Lubrication”, ASLE/ASME Tribology Conference, Pittsburgh, USA,, pp 211-215.
147. Zielinski, P. (1997), “The Influence of Surface Roughness on the
Performance of Hydrodynamic Journal Bearings”, Ph.D. Thesis,. School of Mechanical, Manufacturing and Medical Engineering, Queensland University of Technology, Brisbane, pp. 133.
.
203
APPENDICES
SUMMARY OF APPENDICES
APPENDIX A- Publications
1. Sharma, S., C. and Hargreaves, D (2001), “Effect of Solid Contaminants on Journal Bearing Performance”, World Tribology Conference, Vienna, pp.1-4
2. Sharma, S. C., Hargreaves, D. and Scott, W., (2004), Influence of Errors in Measuring the Radial Clearance of Journal Bearing Performance. 1st International Conference on Advanced Tribology, Singapore. pp.1
3. Sharma, S., Hargreaves, D., Scott, W., (2008), “Journal bearing metrology and manufacturing issues”, 9th Global Congress on Manufacturing and management (GCMM 2008), 12-14 November 2008, pp (paper accepted).
4. Sharma, S., Hargreaves, D., Scott, W., (2008), “Characterisation of antiwear additives”, 9th Global Congress on Manufacturing and management (GCMM 2008), 12-14 November 2008, pp. (abstract accepted- paper to be published in The Journal of Computational materials and Surface Engineering).
5. Sharma, S., Hargreaves, D., Scott, W. (2008), “Characterisation of additives using out-of roundness traces”, 2nd International Conference on Advance Tribology 2008 (ICAT 2008), 3-5 December 2008, Singapore (paper accepted)
APPENDIX B-Fortran Program
A Fortran program was developed for minimum oil film thickness calculations.
These calculations were used to find out eccentricity at no load condition, and also to
calculate the minimum oil film thickness for ideal conditions when there is no
contaminant mixed with the oil.
APPENDIX C- ESDU output
ESDU OUTPUT: Some examples of theoretical calculations for different test
conditions have been presented. These were acquired using ESDU program 84031
(version 1996). The output gives bearing performance parameters such as: minimum
204
oil film thickness, attitude angle, lubricant flow rate etc.
APPENDIX D- Micro-graphs Selected micrographs have been shown for
examining the embedding effect of particles. These micrographs are prepared for
both shaft sleeve as well as journal bearing.
APPENDIX E- Out-of-roundness traces
Selected out-of-roundness traces have been included for demonstrating the effect of
wear and change in roundness of the test bearings. Shaft sleeves have not been
included.
APPENDIX F- Surface roughness traces
Selected surface roughness traces after the tests are presented in the appendix. These
were obtained from Taylor Hobson’s Surtronic 3+ and Talysurf for bearings and
shaft sleeve, in circumferential as well as transverse directions have been shown in
this Appendix. These traces are examples of the roughness in the wear zones of test
bearing elements, and are as close to the average values as possible.
205
APPENDIX A- PUBLICATIONS
210
Abstract accepted: 9th Global Congress on Manufacturing and management (GCMM 2008), 12-14
November 2008
Journal bearing metrology and manufacturing issues
Subhash Sharma*, Doug Hargreaves and Will Scott
School of Systems Engineering, Queensland University of Technology, 2 George St , Brisbane, Australia School of *Aerospace, Mechanical and Manufacturing Engineering, RMIT University, Melbourne3001, email: [email protected]
Abstract:
Journal bearings lubricated with oils containing solid contaminants are subjected to
malfunctioning and can cause premature failures. The malfunctioning of machines
results in poor quality output and premature failures, leading to increased downtime.
In an experimental study of journal bearing lubrication, the radial clearance in a
bearing was measured. The investigation revealed that the radial clearance varies
along the periphery of the bearing. This clearance varies so much so that the small
solid particles of the size of the minimum oil film thickness in the bearing contact
can trap and hamper the performance of a machine and reduce the bearing life. Thus
the metrology of bearing clearance measurement, manufacturing process and the
procedures for bearing design need to be reviewed, if a bearing is subjected to
contaminated environment. A simply statement that the hydrodynamic bearings
should operate on lambda ratio 10 is not enough.
211
Abstract accepted: 9th Global Congress on Manufacturing and management (GCMM 2008), 12-14
November 2008
Characterisation of additives using out-of-roundness traces
Subhash Sharma*, Doug Hargreaves and Will Scott
School of Systems Engineering, Queensland University of Technology, 2 George St , Brisbane, Australia School of *Aerospace, Mechanical and Manufacturing Engineering, RMIT University, Melbourne3001, email: [email protected]
ABSTRACT:
In a tribological study, out-of-roundness traces have been used for characterising the
antiwear additives. Wear tests have been conducted on a journal bearing and wear in
bearings was estimated using out-of-roundness traces. The study showed that the
Talyrond instrument can be successfully used for studying the wear performance of a
bearing. This study has been further extended, and antiwear additives used for
treating the solid contaminants contained in the lubricating oils were characterised
for their wear performance. Wear characteristic numbers have been derived for some
additives using out-of-roundness traces. At present, there are various types of
antiwear additives available in the market and their manufacturers claim high about
their efficacy without any substantial proof. The proposed wear characteristic
numbers can be useful in selecting the most appropriate additive for applications
where bearings are exposed to solid contaminants.
212
Characterisation of antiwear additives
Subhash Sharma*, Doug Hargreaves and Will Scott School of Systems Engineering, Queensland University of Technology, 2 George St , Brisbane, Australia School of *School of Aerospace, Mechanical and Manufacturing Engineering, RMIT University, Melbourne3001, email: [email protected] KEY WORDS: Journal bearing, antiwear additives, out-of-roundness, contaminants, wear ABSTRACT In this tribological study, out-of-roundness traces have been used for characterising antiwear additives. Wear tests have been conducted on a journal bearing and wear depth has been computed from the out-of-roundness traces obtained before and after the wear tests on a bronze journal bearing. The traces give wear depth at different locations and also the maximum wear depth in the bearing. The wear depth results computed from this method have been compared with the measured values and were found a good comparison. Wear characteristic Equations have been derived from this data which gives unique wear signature of wear reducing behaviour of an additive. INTRODUCTION: Use of Antiwear additives is very common in machines operating in dusty environments. There are several products available in the market and manufacturers make unsubstantiated claims about efficacy of their products. There is a need for a tool by which a user can rank these products for their efficacy and choose one that suits his requirements. Journal bearings are worst affected bearing component There are very few studies carried out on the bearings operating with oils containing solid contaminants treated with antiwear additives. Some studies have been carried in this area without using antiwear additives (Mckee, 1927, Roach and Mich, 1950, Elwell, 1977, Duchowski, 1998, Maru 2006). The objectives of their studies were different such as: to study the wear modes, embedding of particles, friction. In a recent study on journal bearing wear with contaminants treated with antiwear additives was carried by Maru et.al. (2006) but his main concern was to study the wear modes and mechanisms. Rowe (1986) determined the coefficient of wear for some antiwear additives using a four-ball test machine and clean oils. Thus there is a need to conduct experiments aimed to study the effect of antiwear additives on bearings operating with oils containing solid contaminants and find a method by which efficacy of these additives could be adjudged. It is difficult to measure small quantities of wear in journal bearings directly, with precision especially when lubricated bearings are dust laden. This study deals with wear tests under the said conditions, where a method has been developed to characterise the antiwear additives using out-of roundness traces of the journal bearings. Using this method Method: Wear tests were conducted on a journal bearing test rig keeping the same operating conditions and specifically the sliding distance 7536 m and ‘K’ ratio (minimum film thickness to particle size ratio) close to 1. Aluminium oxide Al2O3 of 16 microns size was chosen for this study. The wear tests were run on a journal bearing tests rig where a bronze bearing of 40 mm diameter and L/D ratio equal to one was used. The bearing shaft was made of steel sleeve. The out-of-roundness traces were prepared before and after the tests as shown in Figure 1a and b. Wear depths in the wear zone were computed from these traces and their values were compared with the measured values. The process has been shown graphically in Figure 2a and 2b and has been described step-by-step as below: Steps1) Prepare two out-of-roundness traces of a journal bearing i.e. before and after the wear tests (Figure 1 a and b) respectively). Step2) Enlarge both the traces to a suitable magnification (same magnification) and superimpose them. Step 3) Mark the wear zone, by labelling inner crescent and outer crescent, Step 4) mark the probable Trace centre (O) and mark the longest distance point on outer crescent D Step 5) Draw a line from the trace centre ‘O’ to the outer crescent such that it is the longest distance point at the outer crescent mark this point D and mark point W where it intersects the inner crescent, giving WD as maximum wear depth WD. Step 6) Locate point D1 and w17 on either side of D at 60 degree angular displacement either sides respectively, and so the w 1 to w 17 on the inner
213
Graduations marked
magnified traces convert the marked division in microns per millimetre called Scale Factor (SF). Thus record the wear depth at each node. As reported in Table 1.
a) b) Figure 5.4 Out-of-roundness traces a) before wear test b) after wear test
Figure 2a Actual trace of a test bearing with redrawn shape
Discussion: The results show that there is a variation up to 20 %. The measurements with hole-test-gauge showed that it is difficult to locate the location of wear depth manually. This becomes even harder when wear quantity is too small. Traces clearly indicate the wear zone even when the wear depth is too small. This is mainly due to the property of out-of-roundness measurement equipment which amplifies the departure of bearing circumference without amplifying the overall dimensions of the work piece. Though the magnification on Talyrond equipment is limited it was further enhanced while photocopying the traces. Thus the computed values of wear depths are more precise than the measured values from conventional devices.
Inner Crescent Arc
Nodes
w1d1
w17 d17
WD
Out-of-roundness trace
d2
d6
O
d3
D
W
w 2
d 16
Trace Centre
Outer Crescent Arc
Computed shape before test
Graduations
Max. wear depth
CSWA
Inner Crescent Arc
214
Figure 2b Graphical representation of out-of-roundness trace method
Results:The wear depth data is recoded in Table 1.
Table 1. Computed maximum wear depth data Nodes/ wear depth (wndn)
Grid distance mm
(wndn) (µm) A2
(wndn) (µm) A3
(wndn) (µm) A4
(wndn) (µm) A5
(wndn) (µm) A6
(wndn) (µm) A7
1 2.616 0 0 0 0 0 0 2 5.233 2.4 0 2 0 1.5 1 3 7.85 8.8 2 5 3 2.5 5 4 10.466 15.2 6 13 11 5 10 5 13.083 22.4 14 16 16 8 14 6 15.7 30.4 18 21 21 9 16 7 18.316 42.4 23 28 27 10 20 8 20.933 43.2 24 34 29 10 21 9 23.55 41.6 29 35 31 9 22 10 26.166 38.4 28 33 31 7.5 21 11 28.783 30.4 24 36 30 5.5 18 12 31.4 24.8 21 34 29 4 12 13 34.016 18.4 15 29 24 2 5 14 36.633 11.2 7 26 20 1 0 15 39.25 7.2 2 16 11 0 0 16 41.866 3.2 0 10 0 0 0 17 44.483 0 0 0 0 0 0
These wear depths were also measured (WDmax) using hole-test-gauge the results have been reported and compared graphically in Figure 2.
215
APPENDIX –B BEARING FORTRAN PROGRAM
C PROGRAM WORN BEARING PERFORMANCE (HRWD1)
C ****************************************************************
C *MODIFIED WEAR DEPTH ADDED AT EACH GRID
C * THIS PROGRAM IS FOR AXIAL BEARING IT CALCULATES FILM THICK *
C * BY PRESUMING EE=0.01 AND COMPARES LOAD CALCULATED AND APPLIED*
C *BY INCRESING EE VALUE IN STEPS. FILM THICKNESS IS MODIFIED BY *
C * INCORPORATING WEAR PROFILE EQUATION DEVELOPED IMERICALLY *
C ****************************************************************
REAL C,ETA,WR,EE,AR,AL,WB,WW,ORF
INTEGER N,M,WA,AN
DIMENSION P(50,20), H(50),C1(50),C2(50),C3(50),F1(20),F2(20)
C DIMENSION WD(50), THW(50)
C INPUT: N= 46 THETA & M=9 IN Z DIRECTION'
C NORMALLY C=0.000080 METERS, WA=500N,AN=500RPM
C READ (5,*)N,M,C,WA,AN
WRITE(6,*)'INPUT DATA: N,M,C,WA,AN,ETA'
READ (5,*)N,M,C,WA,AN,ETA
C OVERRELAXATION FACTOR,ORF=1.3 IN GENERAL WR=0 STATIC BRG
C R IS L/D RATIO = 1
R=1.0
C CHANGE ETA IF FILON CASE
216
C ETA=0.042
ORF=1.3
EE=0.01
WR=1.0
C CALCULATE NON DIM. LOAD
c LENGTH OF BEARING IS L=40MM, AR IS THE RADIUS =20MM=AL/2
AL=0.040
AR=AL/2.
C NON DIEMSIONAL BEARING CHARACTERISTIC NUMBER HAMROCK
PI=4.*ATAN(1.)
C WW=(1.6*WA*C**2)*6/(ETA*AL*AN*AR**3)
C BEARING SPEED WB IN RADIANS PER SECOND
WB=2.*PI*AN/60.
WRITE (6,*)'WB=',WB
WRITE (6,*)'AN=',AN
WW=WA/(ETA*WB*AL*AR*(AR/C)**2)
WRITE (6,*)'WW=',WW
DEE=0.0
DPHII=0.0
RR=(1./R)**2
C GRID NUMBERING
I5=N
I6=I5+1
I4=I5-1
I3=I4-1
217
AI3=I3
PI=4.*ATAN(1.)
C GRID SPACING DEL= 2*PI/46-2 IN RADIANS, DELZ=M/9-2
C SINCE ONE GRID WILL BE OVERLAPPED
DEL=2.*PI/AI3
DELZ=1./AJ3
DEL2=DEL**2
DELZ2=DELZ**2
C WRITE(6,*)'AI3=',AI3,'AJ3=',AJ3,'DEL=',DEL,'DELZ=',DELZ
C INITIALIZATION OF PRESSURE
100 DO I=1,I6
DO J=1,J6
P(I,J)=0.0
ENDDO
ENDDO
ITER=1
SPO=0.0
DO I=1,I5
DO J=1,J5
SPO=SPO+P(I,J)
ENDDO
ENDDO
CO=RR*(DEL2/DELZ2)
DO I=1,I5
AI=I-2
THETA=AI*DEL
H(I)=1.+EE*COS(THETA)
218
C1(I)=(3.*EE*SIN(THETA)*DEL)/(2.*H(I))
C2(I)=(6.*EE*SIN(THETA)*DEL2)*(1.-2.*WR*DPHII)/(H(I)**3)
C3(I)=(12.*WR*DEE*COS(THETA)*DEL2)/(H(I)**3)
ENDDO
C++++++++++++++++++++++++++++++++++++++++++++++++++++++++++++
C WRITE(6,*)C1(I),C2(I)
C WEAR DEPTH CALCULATIONS
C DO I=IWS,IWF
C THETA=(I-2)*DEL
C THW=(I-IWS)*DEL
C X=THW
C ESTEROL WEAR DEPTH EQUATION
C 20 GRID SECOND SIXTH ORDER WEAR EQUATION FOR TEST A7
C WD=((-11.458*X*X+26.095*X-4.3477)*0.000001)/C
C 20 GRID SIXTH ORDER WEAR EQUATION FOR TEST A7
C WD=(13.276*X**6-96.637*X**5+285.4*X**4-436.74*X**3
C /+321.45*X**2-52.569*X+1.9385)*1000000/C
C H(I)=1.+EE*COS(THETA)-WD
C +++++++++++++++++++++++++++++++++++++++++++++++++++++++++++++
C " FILM THICKNESS WITH WD ADDED AT EACH NODE"
C ADDITIVE /TEST NUMBER
IF (H(I).GE. H(16))THEN
H(16)=H(16)+0.0
H(17)=H(17)+0.0
H(18)=H(18)+2*1.0E-06/C
219
H(19)=H(19)+4*1.0E-06/C
h(20)=H(19)+6*1.0E-06/C
H(21)=H(19)+8*1.0E-06/C
H(22)=H(19)+10*1.0E-06/C
H(23)=H(19)+12*1.0E-06/C
H(24)=H(19)+14*1.0E-06/C
H(25)=H(19)+12*1.0E-06/C
H(26)=H(19)+10*1.0E-06/C
H(27)=H(19)+8*1.0E-06/C
H(28)=H(19)+6*1.0E-06/C
H(29)=H(19)+4*1.0E-06/C
H(30)=H(19)+2*1.0E-06/C
H(31)=H(19)+0.0
H(32)=H(19)+0.0
ENDIF
C1(I)=(3.*EE*SIN(THETA)*DEL)/(2.*H(I))
C2(I)=(6.*EE*SIN(THETA)*DEL2)*(1.-2.*WR*DPHII)/(H(I)**3)
C3(I)=(12.*WR*DEE*COS(THETA)*DEL2)/(H(I)**3)
IF (H(I) .GE. H(33)) THEN
H(I)=1.+EE*COS(THETA)
ENDIF
C1(I)=(3.*EE*SIN(THETA)*DEL)/(2.*H(I))
C2(I)=(6.*EE*SIN(THETA)*DEL2)*(1.-2.*WR*DPHII)/(H(I)**3)
C3(I)=(12.*WR*DEE*COS(THETA)*DEL2)/(H(I)**3)
C WRITE (6,*)H(15),H(16),H(24),H(26),H(32),H(33),H(36)C
C WRITE(6,*)C1(18),C2(18)
C OPEN(55,FILE='HWEAR AREA.DAT')
220
C WRITE(55,555)(H(I),I=1,47)
C555 FORMAT(2X,8F8.2)
500 DO I=1,I6
P(I,1)=P(I,3)
ENDDO
DO J=2,J4
DO I=2,I4
TERM=(P(I+1,J)+P(I-1,J)+CO*(P(I,J+1)+P(I,J-1))
/-C1(I)*(P(I+1,J)-P(I-1,J))+C2(I)-C3(I))/(2.*(1.+CO))
ERR=TERM-P(I,J)
P(I,J)=P(I,J)+ORF*ERR
IF (P(I,J) .LT. 0.0)P(I,J)=0.0
IF (I.EQ.2) GO TO 101
IF (I .EQ.I4) Go TO 102
GO TO 50
101 P(I5,J)=P(2,J)
GO TO 50
102 P(I,J)=P(I4,J)
50 ENDDO
ENDDO
C ++++++++++ AXIAL GRROVE CONDITION++++++++++
C DO I=26,29
C DO J=2,7
C P(I,J)=101.3*c**2*60.0/(ETA*AR**2*31.4*AN)
C WRITE (6,*)'XP(I,J)=', P(26,2),P(29,7)
C P(I,J)=0.123
c P(I,J)=1.0
221
c ENDDO
c ENDDO
c WRITE (*,*)P(26,2),P(28,3),P(29,7)
C +++++++ UPTO THIS POINT MODIFICATIONS+++++++++
C TEST FOR CONVERGENCE
SPN=0.0
DO I=1,I5
DO J=1,J5
SPN=SPN+P(I,J)
ENDDO
ENDDO
ERR=1.-SPO/SPN
IF(ABS(ERR) .LE. 0.01) GO TO 400
SPO=SPN
ITER=ITER+1
GO TO 500
400 SPO=SPN
C WRITE(6,7) ((P(I,J),I=1,I5),J=1,J5)
C7 FORMAT(1X,8F10.3/)
C OPEN(51,FILE='PRESSURE1.DAT')
C DO I=2,I5
C WRITE(51,52)I,P(I,2)
C52 FORMAT(I3,1X,E10.3)
C ENDDO
C EVALUATION OF LOAD BEARING CAPACITY
DO J=2,J6
F1(J)=0.0
222
F2(J)=0.0
DO I=2,I3,2
AI=I-2
TH1=AI*DEL
TH2=(AI+1.)*DEL
TH3=(AI+2.)*DEL
TERM1=P(I,J)*COS(TH1)+4.*P(I+1,J)*COS(TH2)+P(I+2,J)*COS(TH3)
TERM2=P(I,J)*SIN(TH1)+4.*P(I+1,J)*SIN(TH2)+P(I+2,J)*SIN(TH3)
F1(J)=F1(J)+TERM1*DEL/3.
F2(J)=F2(J)+TERM2*DEL/3.
ENDDO
ENDDO
WX=0.0
WY=0.0
DO J=2,J4,2
TERM1=F1(J)+4.*F1(J+1)+F1(J+2)
TERM2=F2(J)+4.*F2(J+1)+F2(J+2)
WX=WX+TERM1*DELZ/3.
WY=WY+TERM2*DELZ/3.
ENDDO
W=SQRT(WX*WX+WY*WY)
IF (W.GE.WW) THEN
GO TO 200
ELSE
EE=EE+0.01
GO TO 100
ENDIF
223
200 ATT=-ATAN(WY/WX)*180/PI
DO I=1,I5
AI=I-2
THETA=AI*DEL
H(I)=(1.+EE*COS(THETA))*C*1000000.0
ENDDO
C OPEN(52,FILE='OPERATING DATA.DAT')
C WRITE(52,4)
C 4 FORMAT(10X,'l/D',5X,'E',7X,'W',8X,'PHI'/)
C WRITE(52,3)R,EE,WW,ATT
C WRITE (52,3)R,EE,WW,ATT,WA
C 3 FORMAT(2X,F7.2,1X,F8.5,F9.5,F10.5,F5.2/)
OPEN(51,FILE='FILMTHICKNESS.DAT')
WRITE(51,111)
111 FORMAT(2X,'NAME OF THE ADDITIVE.....')
WRITE(51,112)
112 FORMAT(2X,'WA',7X,'AN',6X,'EE',6X,'WW',6X,'PHI'
/,5X,'ETA',5X,'L/D',5X,'C'/)
WRITE(51,113)WA,AN,EE,WW,ATT,ETA,R,C
113 FORMAT(1X,F6.1,4X,F6.1,3X,F6.4,2X,F6.2,2X,F6.2,
/2X,F6.2,4X,F3.1,4X,F7.6)
WRITE(51,7)(H(I),I=1,47)
C7 FORMAT(1X,8E12.3)
WRITE(51,8)(P(I,2),I=1,46)
7 FORMAT(1X,F8.2)
8 FORMAT(2X,F14.6)
224
C OPEN(52,FILE='INPUT DATA.DAT')
C WRITE(52,8)
C 8 FORMAT(4X,'LOAD',5X,'SPEED',5X,'VISCOSITY',5X,'NDLOAD'/)
C WRITE(52,9)WA,AN,ETA,WW
C 9 FORMAT(4X,F6.2,1X,F8.2,1X,F10.4,1X,F10.3//)
WRITE(6,*) 'WW=',WW
STOP
END
225
APPENDIX –C EXAMPLES OF ESDU BEARING OUTPUT
ESDU A9305
* Example 1 of Item 84031 * ~~~~~~~~~~~~~~~~~~~~~~~ * Two bearings are required to support the weight of the rotor of an * electric motor. The load on each bearing is 38 kN under both start- * up and running conditions. It is anticipated that the rotor would * be started and stopped once each day. The rotor has a steel journal * 0.25 m diameter and rotates unidirectional at 6.67 rev/s (400 rev/min). * The shaft angular deflection is calculated to be 2.0E-4 rad at the * bearing. Space limitations restrict the maximum width of the bearing * to 0.25 m. It is assumed that the feed temperature will be 40 deg C * and the feed pressure 1.0E5 N/m^2 (1 bar). lubDat name of lubricant database = MIN.DAT lubID lubricant identifier = VG46 Tf lubricant feed temperature = 40.0 deg C Pf lubricant feed pressure = 1.0E5 N/m^2 d diameter of journal = 0.04 m b axial length of bearing = 0.04 m a groove axial length = 0.036 m wg circumferential width of lubricant groove = 0.01 m Cd diametral clearance (minimum case) = 160E-6 m
N frequency of rotation of journal = 6.66
rev/s
W running load on bearing = 500 N Ws start-up load on bearing = 500 N beta angular misalignment = ? rad � Plain Text Attachment [ Download File | Save to my Yahoo! Briefcase ] ---------------------------------------------------------------------- ESDU International plc. PROGRAM A9305 ESDUpac Number: A9305
226
ESDUpac Title: Calculation methods for steadily loaded,
axial groove hydrodynamic journal bearings Data Item Number: 93005 Data Item Title: Calculation methods for steadily loaded, axial groove hydrodynamic journal bearings (Guide to use of computer program A9305). ESDUpac Version: 1.1, June 1996. (See Data Item for full input/output specification and interpretation) ---------------------------------------------------------------------- Name of input data file EXP1_SHARMA.IN INPUT DATA ~~~~~~~~~~ * Example 1 of Item 84031 * ~~~~~~~~~~~~~~~~~~~~~~~ * Two bearings are required to support the weight of the rotor of an * electric motor. The load on each bearing is 38 kN under both start- * up and running conditions. It is anticipated that the rotor would * be started and stopped once each day. The rotor has a steel journal * 0.25 m diameter and rotates unidirectionally at 6.67 rev/s (400 rev/min). * The shaft angular deflection is calculated to be 2.0E-4 rad at the * bearing. Space limitations restrict the maximum width of the bearing * to 0.25 m. It is assumed that the feed temperature will be 40 deg C * and the feed pressure 1.0E5 N/m^2 (1 bar). lubDat name of lubricant database = MIN.DAT lubID lubricant identifier = VG46 Tf lubricant feed temperature = 40.0 deg C Pf lubricant feed pressure = 1.0E5 N/m^2 d diameter of journal = 0.04 m b axial length of bearing = 0.04 m a groove axial length = 0.036 m wg circumferential width of lubricant groove = 0.01 m Cd diametral clearance (minimum case) = 160E-6 m N frequency of rotation of journal = 6.66 rev/s W running load on bearing = 500 N
Ws start-up load on bearing = 500 N
beta angular misalignment = ?
227
rad OUTPUT DATA ~~~~~~~~~~~ LUBRICATION SYSTEM ~~~~~~~~~~~~~~~~~~ Lubricant database : MIN.DAT Lubricant description: ISO VG 46 mineral oil rho density of lubricant = 0.850 E+03 kg/m^3 rhoC volumetric heat capacity = 1.70 E+06 J/m^3 K kappa thermal diffusivity of lubricant = 80.0 E-09 m^2/s Pf lubricant supply pressure = 0.100 E+06 N/m^2 Tf lubricant supply temperature = 40.0 deg C BEARING DIMENSIONS ~~~~~~~~~~~~~~~~~~ d diameter of journal = 40.0 E-03 m b axial length of bearing = 40.0 E-03 m Cd diametral clearance = 0.160 E-03 m a axial length of lubricant groove = 36.0 E-03 m wg circumferential width of groove = 10.0 E-03 m dg recommended minimum groove depth = 3.20 E-03 m bdRat bearing length to diameter ratio = 1.00 CdRat diametral clearance ratio = 4.00 E-03 abRat groove length to bearing length ratio = 0.900 theta angular extent of lubricant groove = 28.6 deg OPERATIONAL PARAMETERS ~~~~~~~~~~~~~~~~~~~~~~ Default values have been assigned to the unspecified operational parameters. N frequency of shaft rotation = 6.66 rev/s W applied load = 0.500 E+03 N Ws load at start-up = 0.500 E+03 N Prun specific loading for running load = 0.313 E+06 N/m^2 Pstart specific loading for start-up load = 0.313 E+06 N/m^2 beta angular misalignment = 0.00 rad RESULTS OF ANALYSIS ~~~~~~~~~~~~~~~~~~~ Re bearing Reynolds number = 1.49
ReCrit critical Reynolds number = 0.809 E+03
NOTE: Bearing lubricant flows are laminar.
228
Results of laminar flow analysis ~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~ Q lubricant flow rate into bearing = 3.02 E-06 m^3/s H power loss = 2.61 W Tmax bearing maximum temperature = 40.6 deg C Tout lubricant outlet temperature = 40.2 deg C eccRat eccentricity ratio (for aligned bearing) = 0.794 psi attitude angle = 30.6 deg hmin minimum film thickness (at edge of bearing) = 16.5 E-06 m hs safe allowable minimum film thickness = 4.28 E-06 m Ra recommended maximum surface roughness value = 0.55 E-06 m ---------------------------------------------------------------------- ***Normal end Plain Text Attachment [ Download File | Save to my Yahoo! Briefcase ] ESDU A9305 * Example 1 of Item 84031 * ~~~~~~~~~~~~~~~~~~~~~~~ * Two bearings are required to support the weight of the rotor of an * electric motor. The load on each bearing is 38 kN under both start- * up and running conditions. It is anticipated that the rotor would * be started and stopped once each day. The rotor has a steel journal * 0.25 m diameter and rotates unidirectionally at 6.67 rev/s (400 rev/min). * The shaft angular deflection is calculated to be 2.0E-4 rad at the * bearing. Space limitations restrict the maximum width of the bearing * to 0.25 m. It is assumed that the feed temperature will be 40 deg C * and the feed pressure 1.0E5 N/m^2 (1 bar). lubDat name of lubricant database = MIN.DAT lubID lubricant identifier = VG46 Tf lubricant feed temperature = 40.0 deg C Pf lubricant feed pressure = 1.0E5 N/m^2
d diameter of journal = 0.04 m b axial length of bearing = 0.04 m a groove axial length = 0.036 m wg circumferential width of lubricant groove = 0.01 m Cd diametral clearance (minimum case) = 178E-6 m N frequency of rotation of journal = 8.33 rev/s W running load on bearing = 500 N Ws start-up load on bearing = 500 N beta angular misalignment = ? rad
�
229
Plain Text Attachment [ Download File | Save to my Yahoo! Briefcase ] ---------------------------------------------------------------------- ESDU International plc. PROGRAM A9305 ESDUpac Number: A9305 ESDUpac Title: Calculation methods for steadily loaded, axial groove hydrodynamic journal bearings Data Item Number: 93005 Data Item Title: Calculation methods for steadily loaded, axial groove hydrodynamic journal bearings (Guide to use of computer program A9305). ESDUpac Version: 1.1, June 1996. (See Data Item for full input/output specification and interpretation) ---------------------------------------------------------------------- Name of input data file EXP1_HT_SHARMA89.IN INPUT DATA ~~~~~~~~~~ * Example 1 of Item 84031 * ~~~~~~~~~~~~~~~~~~~~~~~ * Two bearings are required to support the weight of the rotor of an * electric motor. The load on each bearing is 38 kN under both start- * up and running conditions. It is anticipated that the rotor would * be started and stopped once each day. The rotor has a steel journal * 0.25 m diameter and rotates unidirectionally at 6.67 rev/s (400 rev/min). * The shaft angular deflection is calculated to be 2.0E-4 rad at the * bearing. Space limitations restrict the maximum width of the bearing * to 0.25 m. It is assumed that the feed temperature will be 40 deg C * and the feed pressure 1.0E5 N/m^2 (1 bar). lubDat name of lubricant database = MIN.DAT lubID lubricant identifier = VG46 Tf lubricant feed temperature = 40.0 deg C Pf lubricant feed pressure = 1.0E5 N/m^2 d diameter of journal = 0.04 m b axial length of bearing = 0.04 m a groove axial length = 0.036 m wg circumferential width of lubricant groove = 0.01 m Cd diametral clearance (minimum case) = 178E-6 m
N frequency of rotation of journal = 8.33
230
rev/s
W running load on bearing = 500 N Ws start-up load on bearing = 500 N beta angular misalignment = ? rad OUTPUT DATA ~~~~~~~~~~~ LUBRICATION SYSTEM ~~~~~~~~~~~~~~~~~~ Lubricant database : MIN.DAT Lubricant description: ISO VG 46 mineral oil rho density of lubricant = 0.850 E+03 kg/m^3 rhoC volumetric heat capacity = 1.70 E+06 J/m^3 K kappa thermal diffusivity of lubricant = 80.0 E-09 m^2/s Pf lubricant supply pressure = 0.100 E+06 N/m^2 Tf lubricant supply temperature = 40.0 deg C BEARING DIMENSIONS ~~~~~~~~~~~~~~~~~~ d diameter of journal = 40.0 E-03 m b axial length of bearing = 40.0 E-03 m Cd diametral clearance = 0.178 E-03 m a axial length of lubricant groove = 36.0 E-03 m wg circumferential width of groove = 10.0 E-03 m dg recommended minimum groove depth = 3.56 E-03 m bdRat bearing length to diameter ratio = 1.00 CdRat diametral clearance ratio = 4.45 E-03 abRat groove length to bearing length ratio = 0.900 theta angular extent of lubricant groove = 28.6 deg OPERATIONAL PARAMETERS ~~~~~~~~~~~~~~~~~~~~~~ Default values have been assigned to the unspecified operational parameters. N frequency of shaft rotation = 8.33 rev/s W applied load = 0.500 E+03 N Ws load at start-up = 0.500 E+03 N
Prun specific loading for running load = 0.313 E+06
N/m^2
Pstart specific loading for start-up load = 0.313 E+06 N/m^2 beta angular misalignment = 0.00
rad
231
RESULTS OF ANALYSIS ~~~~~~~~~~~~~~~~~~~ Re bearing Reynolds number = 2.09 ReCrit critical Reynolds number = 0.766 E+03 NOTE: Bearing lubricant flows are laminar. Results of laminar flow analysis ~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~ Q lubricant flow rate into bearing = 4.19 E-06 m^3/s H power loss = 3.63 W Tmax bearing maximum temperature = 40.9 deg C Tout lubricant outlet temperature = 40.2 deg C eccRat eccentricity ratio (for aligned bearing) = 0.793
psi attitude angle = 30.6
deg hmin minimum film thickness (at edge of bearing) = 18.4 E-06
m
hs safe allowable minimum film thickness = 4.28 E-06 m
Ra recommended maximum surface roughness value = 0.55 E-06 m
---------------------------------------------------------------------- ***Normal end
232
233
234
235
APPENDIX-D MICROGRAPHS (SURFACE IMAGES)
Figure D-A1.1Bearing before Test A1 Figure D-A1.2 Bearing after Test A2X20
Figure D-A1.3 Bearing after Test A2 X50 Figure D- A1.4 Sleeve after test A2 X20
236
Figure D A1.5 Sleeve before test A1 Figure D A1.6Sleeve after Test A1
Figure D- A17 Sleeve Before Test A2 X10 Figure D-A1.8 Sleeve after Test A2 X50
237
FigureD- A3 Bearing before Test Figure D-A3.3 Bearing after Test A3 X20
Figure D A3.3 Bearing after Test A3 X50 Figure D A3.4 Sleeve after Test A3 (edge unworn)
238
Figure D A4.1 bearing after Test A4 X10 Figure D A4.2 earing after Tets A3X50
Figure D A4.3 Sleeve after wear X10 Figure D A4.4 Sleeve afre Test A4 X50
239
Figure D A5.1. bearing after wear X20 Figure D A5.2Bearing after test A5 X 50
Figure D A5.3. Sleeve before Test A5 Figure D A5.4 Sleeve after Test A5 X 40
240
Figure D A6.1 bearing after Test A6X10 Figure D A6.2 Bearing after Test A6X40
Figure D-A6.3 sleeve after Test A6 Figure D A5.1Sleeve after Test A6 X40
241
Figure D A7.1 Bearing after Test A7 X20 Figure D A7.2 Bearing after Test 7X100
Figure D7 A7.3 Sleeve after Test A7 Figure D 7.4 Sleeve after Test A7 X 40
242
APPENDIX-E OUT–OF–ROUNDNESS TRACES
Figure E- A1.1 OR before Test A1
Figure E –A2.1 OR before Test A2/ after Test A1
243
Figure E. A2.2 OR After Test A2
Figure E. A3.1 OR before Test A3
244
Figure E. A3.2 OR after Test A3
Figure E. A4.1 OR before Test A4
245
Figure E. A4.2 OR after Test A4
Figure E A5.1 OR before Test A5
246
Figure E A5.2 OR after Test A5
Figure E.A6.1 Before Test A6
247
Figure E. A6.2 OR after Test A6
Figure E. A6.3 Shaft sleeve OR before Test A6
248
Figure E. A6.4 Shaft sleeve Or After Test A6
Figure E-A7.1Trace before Test A7
249
Figure E-A7.2 Trace after Test A7
250
Figure E A7.3 Process of wear depth measurement from the trace
251
APPENDIX F –ROUGHNESS TRACES OF BEARINGS AND SHAFT SLEEVE
Figure F-A2.1 Bearing roughness after Test A2
Figure F –A2.2 Bearing transverse roughness after Test A2
252
Figure F- A2.3 Shaft sleeve roughness after Test A2
Figure F- A2.4 Shaft sleeve transverse roughness after test A2
253
Figure F- A3.1 Bearing roughness after Test A3
Figure F- A3.2 Bearing transverse roughness after Test A3
254
Figure F –A3.3 Shaft sleeve roughness after Test A3
Figure F- A3.4 Shaft sleeve transverse roughness after Test A3
255
Figure F–A4.1 Bearing roughness after Test A2
Figure F- A4.2 Bearing transverse roughness after Test A4
256
Figure F- A4.3 Shaft sleeve roughness after Test A4
Figure F- A4.4 Shaft sleeve transverse roughness after Test A4
257
Figure F- A6.1 Bearing roughness after Test A6
Figure F- A6.2 Bearing transverse roughness after Test A6
258
Figure F – A6.3 Shaft sleeve roughness after Test A6
Figure F- A6.4 Shaft sleeve transverse roughness after Test A6
259
Figure F- A7.1 Bearing roughness after Test A7
Figure F- A7.2 Bearing transverse roughness after Test A7
260
Figure F- A7.3 Shaft sleeve roughness after Test A7
Figure F- A7.4 Shaft sleeve transverse roughness after Test A7
261
