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
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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
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BLANK
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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
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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.
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Table of Contents
ABSTRACT ......................................................................................................................... iii
LIST OF PUBLICATIONS.................................................................................................xiv
STATEMENT OF ORIGINALITY.......................................................................................xv
KEY WORDS .....................................................................................................................xvi
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
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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
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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
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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
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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)
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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.
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)
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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
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
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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 )
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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
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
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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
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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.
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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
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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
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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
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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.
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APPENDIX A- PUBLICATIONS
halla
These articles are not available here.Please consult the hardcopy thesis available from QUT Library
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.
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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.
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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)
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
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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 = ?
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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.
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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
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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
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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
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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
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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
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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
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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)
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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
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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
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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
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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
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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
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Figure E. A2.2 OR After Test A2
Figure E. A3.1 OR before Test A3
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Figure E. A3.2 OR after Test A3
Figure E. A4.1 OR before Test A4
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Figure E. A4.2 OR after Test A4
Figure E A5.1 OR before Test A5
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Figure E A5.2 OR after Test A5
Figure E.A6.1 Before Test A6
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Figure E. A6.2 OR after Test A6
Figure E. A6.3 Shaft sleeve OR before Test A6
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Figure E. A6.4 Shaft sleeve Or After Test A6
Figure E-A7.1Trace before Test A7
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Figure E-A7.2 Trace after Test A7
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Figure E A7.3 Process of wear depth measurement from the trace
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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
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Figure F- A2.3 Shaft sleeve roughness after Test A2
Figure F- A2.4 Shaft sleeve transverse roughness after test A2
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Figure F- A3.1 Bearing roughness after Test A3
Figure F- A3.2 Bearing transverse roughness after Test A3
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Figure F –A3.3 Shaft sleeve roughness after Test A3
Figure F- A3.4 Shaft sleeve transverse roughness after Test A3
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Figure F–A4.1 Bearing roughness after Test A2
Figure F- A4.2 Bearing transverse roughness after Test A4
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Figure F- A4.3 Shaft sleeve roughness after Test A4
Figure F- A4.4 Shaft sleeve transverse roughness after Test A4
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Figure F- A6.1 Bearing roughness after Test A6
Figure F- A6.2 Bearing transverse roughness after Test A6
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Figure F – A6.3 Shaft sleeve roughness after Test A6
Figure F- A6.4 Shaft sleeve transverse roughness after Test A6
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Figure F- A7.1 Bearing roughness after Test A7
Figure F- A7.2 Bearing transverse roughness after Test A7
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Figure F- A7.3 Shaft sleeve roughness after Test A7
Figure F- A7.4 Shaft sleeve transverse roughness after Test A7