Tribological study of elastomeric bearings for marine propeller shaft system

Harish Hirani !, Manish Verma

Department of Mechanical Engineering, IIT Bombay, Powai, Mumbai, Maharashtra 400076, India

a r t i c l e i n f o

Article history:Received 6 August 2007Received in revised form18 July 2008Accepted 29 July 2008Available online 16 September 2008

Keywords:Marine shaft bearingFluid film thicknessCoefficient of wearBearing life

a b s t r a c t

Elastomeric compounds, due to their favourable properties like sufficient hardness, toughness andnatural resistance to abrasion and corrosion, are commonly used as bearing material for propeller shaftsystem of Indian Coast Guard Ships. Recently unequal and non-uniform wear of these bearings hasresulted in unscheduled lay off of the Coast Guard Ships. To solve this problem of bearing wear, a mixedlubrication analysis of sea-water lubricated journal bearing has been attempted in the present study. Acomputer code was written to estimate lubricating film thickness for a given set of load and speedcondition, and to predict the lubrication regime for the specified surface roughness parameters. Tovalidate the theoretical analysis performed in the present study, the results obtained from the computersimulation have been compared with the established studies on the water lubricated bearing.

To understand the uneven wear of marine bearings, actual geometric clearances of new and wornout bearings recorded by the ship maintenance team, and the operational data (load, speed andoperating hours), obtained from the log books of ICGS Sangram (AOPV) of Indian Coast Guard, are listedin the present paper. The dynamic viscosity of sea water, surface roughness of propeller shaft andbearings, and particulate contamination has been measured. Finally, the suggestions have been enlistedfor proper operation of shaft-bearing system so as to maintain the wear within the permissible limitsduring ship’s operational cycle.

& 2008 Elsevier Ltd. All rights reserved.

1. Introduction

Indian Coast Guard’s Advanced Offshore Patrol Vessels (AOPVs)are fitted with twin propeller shafts, each powered with 6400 HPengine. Each propeller shaft is 24m long, weighs 10 tons androtates inside four elastomeric journal bearings which arelubricated with sea water. These bearings are plain in lower halfand grooved in upper half as shown in Fig. 1.

Fig. 2 shows the propeller shaft supported in four elastomericbearings. Dimensional details of these bearings are mentioned inTable 1. In addition, Table 1 lists the bearing reactions which aretaken from Ref. [1]. Recently excessive wear of these bearings hasresulted in unscheduled lay off of the Coast Guard Ships. Firstship, ICGS Samar, commissioned in 1995 had successive bearingfailure on both the shafts during two operational cycles (periodbetween two successive dockings). ICGS Sangram, commissionedin 1997 had similar failure in both the shafts on two operationalcycles and one shaft (port) in one operational cycle. Similarfailures were seen in ICGS Sarang which was commissioned in

1999. The aim of the present study is to analyse this problem ofbearing wear, and suggests a remedial action.

Wear down of a water lubricated elastomeric bearing is anobvious phenomenon. However, elastomeric bearings used in allthree ships (ICGS Samar, Sangram, and Sarang) were speciallydesigned and fabricated (by Thordon bearings Inc.). In one lab(David Taylor Research Centre, Bethesda, MD) test, bearing wearof 0.002 in (0.051mm) was observed in over 2000 working hoursof propeller shaft running at 10 rpm against designed elastomericbearing. Expected life of such a bearing is more than 20 years andreplacement of bearings after wear depth of 4.7mm is recom-mended. However, elementary moment analysis [1] reveals thatwearing of bearing 4 redistributes the load, as listed in Table 1,among bearings. As per the gearbox manufacturer, the load ratioforce on bearing 5=force on bearing 6 should not exceed 1.3. Thisconstraint is imposed to restrict the impact loading on gear pairs.At the design stage this ratio was 1.1145 (F5/F6 ¼ 182906/164109).However, after 0.3mm wear of bearing 4, load ratio (F5/F6)increases to (206 453/147401) to 1.4. Therefore, there is aneed to review and establish a new limit on bearing wearfor the replacement of elastomeric bearings compared to believ-ing on the limit of 4.7mm wear depth suggested by bearingmanufacturer.

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Contents lists available at ScienceDirect

journal homepage: www.elsevier.com/locate/triboint

Tribology International

0301-679X/$ - see front matter & 2008 Elsevier Ltd. All rights reserved.doi:10.1016/j.triboint.2008.07.014

! Corresponding author. Tel.: +912225767535; fax: +91222 5726875.E-mail address: hirani@iitb.ac.in (H. Hirani).

Tribology International 42 (2009) 378–390

Bearing wear can be completely avoided by establishing thehydrodynamic lubrication mechanism in water lubricated journalbearings. However, the bearing wear observed in the presentstudy indicates that bearings were operated under mixed and/orboundary lubricated regimes. One possible reason of mixedlubrication condition is that all patrol vessels, fitted with twoengines, operate with single engine during patrol. Under suchconditions, the idling shaft speed, which (due to propulsion of themotoring shaft) ranges between 50 and 70 rpm, may beinsufficient to form hydrodynamic film. To understand thisbearing wear problem, a mixed lubrication analysis of sea-waterlubricated journal bearing is required. Further, to streamline theanalysis, answers to the following questions are essential:

(a) Is hydrodynamic fluid film not getting formed due to:(i) Low rpm operations and/or(ii) Increased radial clearance

(b) What are the implications of increased radial clearance ofbearings?

(c) What is the regime of lubrication in which the bearingoperates with the increased radial clearance at a given rpm?

(d) What is the effect of particulate contamination in sea water?Is it responsible for abrasive and erosive wear of bearingsurface?

(e) What is the effect of surface roughness of bearing and shaftsleeve?

(f) What is the effect of viscosity of sea water? Is it changing withrpm and temperature?

(g) Is it possible to control the wear by changing the dimensionsof bearing and shaft?

In the present study a systematic design methodology hasbeen employed to answer the abovementioned questions andsuggest a suitable remedy to avoid bearing wear in the future.

2. Literature survey

To understand the theoretical aspects of journal bearingoperation two books, ‘Engineering Tribology’ by Stachowiak andBatchelor [2], and ‘Applied Tribology’ by Khonsari and Booser [3]were referred. Hirani et al. [4] have used two pressure correction

factors go and gs to give an analytical expression for maximumpressure. The design table [4] listed in their paper can be used tofind out film thickness, pressure and load carrying capacity.However, the formulations in their paper [4] pertained to hydro-dynamic lubrication regime only. Johnson et al. [5] have stipulatedthat provided a major part of the load is carried by elastohydro-dynamic action, the separation between the two rough surfaces isgiven by the film thickness which would exist between two smoothsurfaces under the same conditions of load, speed and lubricant.Kraker et al. [6] have described a mixed EHL model for finite lengthelastic journal bearings. These authors have used commerciallyavailable finite element code SEPRAN to discretise the Reynoldsequation. Bayer [7] has provided algebraic expression for calcula-tion of depth of bearing wear under aligned and misaligned journalconditions. Messimo Del Din et al. [8] have used an experimentalset up to investigate the utility of environmentally adapted rapeseed–synthetic ester oil over traditional mineral oil. The wearmeasurements evaluate the coefficient of wear using Archard’sequation of wear. The methodology discussed in the paper pertainsto experimental measurement of wear by difference in weight ofbearing liner before and after the experiment and compares twodifferent oils. Hsu et al. [9] have given a comprehensive view ofwear under lubricated conditions. As per these authors wear underlubricated conditions can be classified into two main classes: well-lubricated systems and marginal lubricated systems. However, themethodology to determine wear is again experimental in nature.Rao and Mohanram [10] have presented comprehensive sets ofexperiments to study mixed lubrication of journal bearings. Thesurface topography changes have been statistically analysed. Safar[11] has presented an analysis of a journal bearing describing amaximum allowable value of misalignment at a length to diameterratio of unity. The author has opined that journal misalignmentinfluences load carrying capacity of the bearing. A misalignedbearing consumes more power due to friction than an aligned one.El-Butch and Ashour [12] have dealt with analysing the perfor-mance of a misaligned tilting-pad journal bearing under transientloading condition. Jakeman [13] has presented a model specificallyintended to represent the dynamically misaligned sterntubebearing, for the purpose of conducting lateral vibration analysesof marine propeller shafting.

The methodologies presented in these papers have been dulytaken into account while carrying out the present study.

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Symbols and notations

C radial clearance, mD journal diameter, mH;HPmax non-dimensional film thickness, film thickness at

maximum pressure locationh depth of wear, mh0 minimum thickness of fluid film, mh̄ EHL film thickness, mht effective film thickness, mK empirical wear factor or Archard’s wear coefficientL bearing length, mN journal rotational speed, rpmP bearing pressure W/(LD), PaPmax maximum pressure, Papf fluid pressure (percent of total pressure)pa asperity pressure (percent of total pressure)R journal radius, mS sliding distance, mt tilt ratio or non-dimensional tilt, t ¼ m/C,

T operating time, sU journal surface velocity, m/sV wear volume, m3

W dimensional load capacity, NWe, Wf dimensional load capacity along and perpendicular to

line of centres, NWZ ratio of dimensional load capacity to viscosity, m2/sWeZ, WfZratio of We, Wf to viscosity, m2/sz coordinate in axial direction, mgO, gS pressure correction factors for Ocvirk’s and Sommer-

field bearingsL slenderness ratio (L/D)e eccentricity ratiof attitude angle, radianZ viscosity coefficient of lubricant, Pa sy coordinate in circumferential direction, radiansA surface roughness of bearing surface, mmsB surface roughness of shaft sleeve surface, mms combined roughness of two surfaces, mml film thickness parameter

H. Hirani, M. Verma / Tribology International 42 (2009) 378–390 379

3. Approach for analysis of journal bearing parameters

The following approaches have been adopted:

(a) Reynolds equation, simplified for unidirectional velocityapproximation [2,3], has been used. Expression for filmthickness given in the design table of Hirani et al. [4] modifiedto take care of the misalignment.

(b) Elastohydrodynamic lubrication film thickness and surfaceroughness parameter have been incorporated to predict theregime of lubrication. With the given load, rpm and radialclearance each bearing’s film thickness, regime of lubricationand asperity to fluid pressure ratio has been calculated usingavailable formulations. The results have been compared withsimilar study carried out by Kraker et al. [6] to validate theapproach.

(c) Wear model based on boundary lubrication regime hasbeen used to assess the depth of wear and associated life ofbearing [7].

(d) Asperity and fluid pressure ratio given by Johnson et al. [5]have been used and correlated with regime of lubrication.

3.1. Derivation of pressure expression

The closed form solution to Reynolds equation, for infinitelyshort bearing (Ockvirk’s solution) and infinitely long bearing(Sommerfeld solution) is combined using two correction factors goand gs to approximate the analytical expression for pressure for afinite bearing [3]. The formulations, given in the design table havebeen arrived at considering hydrodynamic lubrication regime.Misalignment factor has been provided in the film thicknessexpression. In misaligned shaft, critical minimum film thicknesswill occur at the edge of the bearing, as shown in Fig. 3a. The basicparameter to describe the tilt of the shaft is the tilt ratio given by

t ¼mC

(1)

Where t is the tilt ratio or non-dimensional tilt, m is the distancebetween the axes of the tilted and non-tilted shaft measured atthe edges of the bearing and C is the radial clearance. Accordingly,expression for non-dimensional film thickness H ( ¼ h0/C) wasmodified as

H ¼ 1þ !#ztL

! "cos y (2)

where Z is the axial coordinate and L the length of bearing and h0the minimum film thickness between smooth surfaces.

The analytical pressure terms given by Ockvirk and Sommer-feld viz PO and PS and modified with two correction factors go andgs [3] is given by

1P¼

gOPO

þgSPS

(3)

where go and gs are given by [3]

gO ¼ 1þ !L1:2½e!5# 1% (4)

gS ¼ eð1#!Þ3

(5)

where L ¼ L=D, is slenderness ratio. To validate the proposedanalytical approach, bearing data given by Sun and Changlin [14]have been used to obtain and plot the pressure profile at variousangular misalignments. Pressure profiles shown in Figs. 3b–e arecomparable to pressure profiles provided in Ref. [14]. Sun andChanglin used finite-difference method to obtained pressureprofile. Further, a comparative study between the values ofmaximum pressure obtained using the present analytical ap-proach and finite-difference method is listed in Table 2. Thesegraphical and tabular results indicate that the proposed analyticalapproach can be used to analyze misaligned hydrodynamicjournal bearing.

3.2. EHL film thickness and surface roughness parameter

The elastomeric water lubricated bearing may experienceelastic deformation [6]. Kraker et al. [6] described a mixedelastohydrodynamic lubrication (EHL) model for finite lengthelastic journal bearing. They employed the finite element methodto solve the coupled system of fluid and structural equations tocompute Stribeck curves at constant load. In the present study, ananalytical approach has been used to evaluate pressure profile andminimum film thickness. To incorporate elastohydrodynamicmodel and compare the results with Kracker et al. [6], the EHLfilm thickness in transverse direction (relevant to the present

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Fig. 1. Bearing with housing.

H. Hirani, M. Verma / Tribology International 42 (2009) 378–390380

study) is obtained using [5]

h ¼ h0 1þ76

sh0

! "2" #

(6)

where h is the EHL film thickness and ho is the minimum filmthickness for smooth surface. The combined roughness of two

surfaces s is given by (sA2+sB

2)0.5. The surface roughness parameterl is calculated to ascertain the regime of lubrication.

l ¼h0s (7)

Bearing is considered to be in boundary lubrication regimewhen lp3.

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6m 6m 6m

UDL 4KN/m

Propeller Weight

40 KN

1 234

P R O P E L L E R S H A F T

5

Mech. S

eal

Gearbox

6

Fig. 2. Propeller shaft with bearings: (a) Coast Guard Ship in Dry dock, (b) shaft withdrawn and showing location of bearings and (c) schematic of propeller shaft withbearings.

H. Hirani, M. Verma / Tribology International 42 (2009) 378–390 381

The formulations given in Eqs. (2)–(7) and those given inRef. [4] have been used to calculate the load carrying capacity ofthe bearing. The calculated load is compared with the given loadand till the time it matches within the prescribed limits, iterationsare continued by incrementing the eccentricity ratio. The flowchart for MATLAB code is shown in Fig. 4. To check the correctnessof programme, bearing data provided in Ref. [6] and listed inTable 3 were tried out. Output results are listed in Tables 4 and 5and plotted in Figs. 5 and 6. Logarithmic plots have beenpresented depicting change in surface roughness parameter andcoefficient of friction with increasing radial clearance andincreasing composite rms surface roughness values of tribo-pair.The results generated by the computer code used in the presentstudy tally, both quantitatively and qualitatively, with the plotspublished by Kraker et al. [6].

3.3. Wear model

Having established the correctness of computer code forjournal bearing parameters, the study was extended to estimatethe bearing life. Journal bearing’s wear and performance areusually related to the increase in clearance between the journaland the bearing. Archard’s equation defines depth of wear h as afunction of bearing load W and sliding distance S. Coefficient ofwear K in the Archard’s equation is a proportionality constant,indicating that wear follows linear relationship. Wear modelsuggested by Bayer [7], although takes linear relationshipbetween volume of wear V, and K and the depth of wear h hasnon-linear relation with length L, radius R and volume of wear V.The expression for wear volume V and h is given by

V ¼ UTKWa (8)

where U is the sliding speed in m/s, T the time in s, K the wearcoefficient in m2/N and Wa the asperity load in N. The empiricalequation for localized depth of wear h, proposed by Bayer [7] foraligned shaft is given by

h ¼ 0:66R#1=3L#2=3V2=3 (9)

where R and L are radius and length of bearing, respectively. Inthis wear model, the underlying assumption is that negligiblewear occurs in the journal. With the known values ofdepth of wear h, surface velocity U, time T and load shared byasperities Wa, the coefficient of wear K for individual bearingsusing Eqs. (8) and (9) can be calculated by substitution andrearrangement:

K ¼ð1:5hÞ3=2R1=2L2pRNTWa

or

K ¼ 0:2924ðhÞ3=2L

R1=2NTWa

(10)

3.4. Asperity and fluid pressure ratio

In many instances of EHL, direct contact between the deformedasperities will still occur in spite of the presence of EHL film. If thelubricating film separating the surfaces is such that it allows somecontact between the deformed asperities then this type oflubrication is called mixed or partial lubrication. Johnson et al.[5] have propounded that, provided a major part of the load iscarried by elastohydrodynamic action, the separation between thetwo rough surfaces is given by the film thickness which wouldexist between two smooth surfaces under the same conditions ofload, speed and lubricant. The authors have shown that anincrease in total load is carried by an increase in fluid pressure anda small increase in asperity contact pressure. The fluid pressure tototal pressure is given by following expression [5]:

pfp

¼h0h

! "6:3

(11)

The pressure shared by the asperity can be deduced as

pap

¼ 1#pfp

(12)

These equations have been used to ascertain pressure sharingbetween fluid and asperity at various rpm/radial clearanceconditions. Under mixed lubrication conditions use of Eqs. (11)and (12) is required to find out the fraction of load (Wa) bared byasperities. Calculated Wa can be used in Eqs. (10) to determine thewear constant for mixed lubrication.

4. Experiments

Elastomeric bearings, used in Indian Coast Guard’s AdvancedOffshore Patrol Vessels (AOPVs), are lubricated with sea water thatmay contain particulates. Further, for reliable mixed lubricationanalysis the measurement of dynamic viscosity of sea water andsurface roughness of propeller shaft and bearings is essential.Therefore, it was necessary to perform particle analysis to checkthe influence of particulate contamination in sea water (forerosive and abrasive wear), to measure the dynamic viscosity ofsea water and its thixotropic behaviour and to determine thesurface roughness of bearing and shaft surface. Experiments forviscosity and particulate contamination were conducted at IITLabs with sample of sea water taken from Naval DockyardMumbai, India. Pertometer M2, available at Instrumentation Lab

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Table 1Bearing dimensions and reactions [1]

Bearing number (PORT andstarboard shafting)

Bearing type Length (mm) Diameter (mm) Dynamic bearingreactionsa (N)

Bearing reactions on0.3mm wear onbearing no. 4

1 Water lubricated slidingbearings

735 365 #60 089 #602162 735 355 #27151 #26 4143 735 350 #33145 #353594 735 345 #26123 #17679

5 (gearbox aft) Oil lubricated rolling bearings 190 265 #182906 #206 4536 (gearbox fwd) 190 265 #164109 #147401

a The reactions are considered positive in the downward direction.

H. Hirani, M. Verma / Tribology International 42 (2009) 378–390382

of IIT was used to check the surface roughness of bearing sampleobtained from M/S Vanson Mumbai. The details of readings andits possible influence on wear analysis are elucidated in forth-coming paragraphs.

4.1. Measurement of surface roughness by perthometer M2

The surface roughness of the bearing samples was measuredusing perthometer M2 in IIT machine tool lab. The summary of

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Fig. 3. Misaligned hydrodynamic journal bearing: (a) misaligned shaft in a journal bearing [1], (b) pressure profile for aligned bearing, (c) pressure profile for 0.0041misalignment, (d) misalignment equal to 0.0071 and (e) pressure profile for 0.011 misalignment.

H. Hirani, M. Verma / Tribology International 42 (2009) 378–390 383

various surface roughness parameters for the bearing surfaces islisted in Tables 6 and 7. Roughness value of 15mm for bearingsurface and 3mm for shaft sleeve surface has been considered inthe present study, which is standard roughness limit stipulatedprior fitment in ship.

4.2. Measurement of viscosity of sea water

Viscosity of sea water was measured at ONGC JRC Lab at IITB.The equipment used was Brookfield make Rheometer. The read-ings taken during the experiment is given in Table 8. The average

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Table 2Values of the maximum fluid pressure in misaligned hydrodynamic journalbearing

Misalignment(deg)

Present study(MPa)

Finite differencemethod (MPa) [14]

% difference

0.000 34.75 33.06 4.860.004 43.91 39.60 9.820.007 73.96 63.58 14.030.01 412.51 415.35 0.69

Start

Acquire parameters

Initialise eccratio

CalculateLoad by integratingPressureterm twice

Compare calculatedload with given Load

Difference > terminating residual

Incrementecc ratio

Yes

Get min filmthickness

No

Get film thicknessparameter

λ > 3

HDLregime Get bearingparameters

λ ≤ 3

Mixed/BL Regime Calculatedepth of wear

END

Preset ‘max h’

Compare ‘h’ with‘max h’

LESS

Get the bearing life

Equal

Fig. 4. Flow chart for programme rpm_vs_ecc_regime.

Table 3Design parameters of reference bearing

Description Parameter Value Units

Bearing radius R 25 mmBearing length L 100 mmRadial clearance C 0.125 mmComposite surface roughness Sq or s 0.424 mm

H. Hirani, M. Verma / Tribology International 42 (2009) 378–390384

value of viscosity at 150 rpm was 0.96mPa s, at 250 rpm it was1.08mPas and at 500 rpm the average value of viscosity was 1.50.In the present study a curve fit equation

0:8331þ 0:0006 rpmþrpm1000

# $2

has been used to represent the water viscosity as a function ofrotational speed of shaft.

4.3. Measurement of particle size in sea water using particle sizeanalyser

The variation in particle size diameter indicates inconsistencyof particulate contamination in sea water. The particulatecontamination varies depending upon depth, distance from shoreline and turbulence level of sea. It was observed that particulatecontamination is high during rough weather when sea is

turbulent. The particle size in sea-water sample had meandiameter of 40–120nm and the average speed (at 250 rpm ofshaft speed) of particle is less than 3m/s, which is not of anyappreciable consequence as far as wear of marine shaft bearing isconcerned. The particulate contamination was therefore, notconsidered in the present study.

5. Data collection

Having obtained the values of characteristic parameters,it was time now to obtain operational data from a ship. It wasdecided to analyze the wear of propeller shaft bearing ofIndian Coast Guard Ship Sangram, an AOPV, which had experi-enced problems of unequal and unsymmetrical wear in itsshaft system bearings. Initial and final radial clearances recordedduring the construction of ship, CGS Sangram, at Goa Shipyard

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Table 4Results of computer code for design parameters at different rpms (s ¼ 0.43 kept constant)

RPM Radial clearance 0.00025mR/C ¼ 0.01

Radial clearance 0.000125mR/C ¼ 0.005

Radial clearance 0.0000625mR/C ¼ 0.0025

Radial clearance 0.00003125mR/C ¼ 0.00125

ht/s m ht/s m ht/s m ht/s m

10 1.38 0.114 1.38 0.114 1.38 0.114 1.38 0.1120 1.38 0.114 1.38 0.114 1.38 0.114 1.38 0.0830 1.38 0.114 1.38 0.114 1.38 0.114 2.24 0.0840 1.38 0.114 1.38 0.114 1.38 0.114 3.04 0.0650 1.38 0.114 1.38 0.114 1.38 0.114 3.93 0.0470 1.38 0.11 1.38 0.11 2.07 0.09 5.5 0.02

100 1.388 0.11 1.388 0.11 3.61 0.05 7.8 0.01200 1.388 0.11 1.87 0.1 7.97 0.01 15.8 0.004300 1.388 0.11 4.73 0.03 11.9 0.004 23.5 0.002400 1.388 0.11 7.28 0.01 15.8 0.004 29.2 0.001500 1.388 0.11 9.63 0.009 19.7 0.003 – –600 1.4 0.11 11.85 0.007 23.7 0.002 – –700 2.18 0.09 13.99 0.005 27.8 0.002 – –800 3.75 0.04 16.06 0.004 – – – –900 5.28 0.02 18.09 0.003 – – – –

1000 6.75 0.01 20.1 0.0025 – – – –1100 8.16 0.005 22.1 0.001 – – – –

Table 5Results of computer code for design parameters at different rpms (radial clearance C ¼ 0.000125m kept constant)

rpm sA ¼ 0.1, sB ¼ 0.2, s ¼ 0.22 sA ¼ 0.2, sB ¼ 0.4, s ¼ 0.43 sA ¼ 0.3, sB ¼ 0.8, s ¼ 0.85

ht/s m ht/s m ht/s m

10 1.38 0.11 1.38 0.11 1.38 0.1130 1.38 0.11 1.38 0.11 1.38 0.1150 1.38 0.11 1.38 0.11 1.38 0.1170 1.38 0.11 1.38 0.11 1.38 0.11

100 1.38 0.11 1.38 0.11 1.38 0.11200 3.7 0.04 1.87 0.1 1.38 0.11300 9.53 0.009 4.73 0.03 2.45 0.08400 14.63 0.004 7.28 0.01 3.78 0.04500 19.3 0.003 9.63 0.009 5.0 0.03600 23.8 0.002 11.85 0.007 6.16 0.02700 28.01 0.002 13.09 0.005 7.28 0.016800 – – 16.06 0.004 8.36 0.012900 – – 18.1 0.003 9.4 0.01

1000 – – 20.1 0.003 10.4 0.0091100 – – 22.1 0.003 11.49 0.0071200 – – – – 12.51 0.006

H. Hirani, M. Verma / Tribology International 42 (2009) 378–390 385

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0

4

8

12

16

20

24

1 10 100 1000 10000RPM

Surfa

ce R

ough

ness

Para

met

er

C = 0.01 C = 0.005 C = 0.0025 C = 0.00125

0

4

8

12

16

20

24

Surfa

ce R

ough

ness

Para

met

er (S

igm

a)

Sigma = 0.22 Sigma = 0.43 Sigma = 0.85

1 10 100 1000 10000RPM

Fig. 5. Variation of surface roughness parameter with (a) increasing radial clearance of bearing and (b) increasing composite rms value (s) of tribo-pair (data refers toTable 4).

0

0.04

0.08

0.12

Coe

ff o

f fri

ctio

n

C = 0.00025 C = 0.000125 C = 0.0000625 C = 0.00003125

0

0.04

0.08

0.12

1 10 100 1000RPM

Coe

ff o

f fri

ctio

n

Sigma = 0.22 Sigma = 0.43 Sigma = 0.85

1 10 100 100001000RPM

Fig. 6. Variation of coefficient of friction with (a) increasing radial clearance of bearing and (b) increasing composite rms value (s) of tribo-pair (data refers to Table 5).

H. Hirani, M. Verma / Tribology International 42 (2009) 378–390386

Ltd and during its two successive docking ships are registeredin Table 9. The recorded clearances were diametric and forthe present study they have been converted into radial clearance.The exploitation pattern and hours of sailing, obtained fromICGS Sangram’s Main Engine Room Log Book, are listed inTable 10.

6. Results and discussions

Table 10 shows that water lubricated bearings undergotransient phenomena of starts and stops (i.e. 304 time shaft was

started/stopped during May 97 to June 2000). Mokhtar et al. [15]studied the starting behaviour of journal bearing and observed arapid build up of hydrodynamic film in all the cases. Theyconcluded that hydrodynamic film formed in a very short time,after which shaft moved in a spiral shaped whirling locusto the steady state operating position. Similarly they studiedstopping motion and reported that during shutting down theshaft followed a typical hydrodynamic locus until rotationceased and then a squeeze film trajectory to final restingposition was observed. These results indicate that start/stoptransient phenomena does not cause major wear on thebearing surface. Under proper load conditions, shaft takes lesserthan half rotation to lift off from bearing surface. Therefore,in the present study it is assumed that starts/stops do notaffect bearing life significantly. However, excessive load, largeclearance and low rotational speed may have affected the bearinglife and estimating wear constant may indicate the source ofbearing wear.

6.1. Coefficient of wear

There are two approaches to find the wear coefficient. In thefirst approach it is assumed that there are only two lubricationdomains: hydrodynamic and boundary. In the hydrodynamicregime (l43), bearing wear is zero and the applied load is baredby full-fluid film. In the boundary regime (lp3), applied load iscarried by asperities and bearing wear is definite. The value of lcan be determined by computer simulation of algorithm given inFig. 4 and Eq. (7). By employing this approach, MATLAB code wasrun with input data of load, bearing length and diameter as listedin Table 1, and different radial clearances. Fig. 7 shows that atradial clearance greater than 0.6mm, all four bearings operate inboundary lubrication regime, meaning thereby that hydrody-namic film does not form even at the highest operating rpm (254).However, when the radial clearance is reduced to 0.4mm forbearing 2, 3 and 4, the average minimum rpm for hydrodynamic

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Table 6Readings of perthometer M2 for shaft sample

Parameter Reading 1 Reading 2 Reading 3 Reading 4

Sampling length (mm) 5.6 17.5 5.6 17.5Lc (mm) 0.8 2.5 0.8 2.5Ra (mm) 0.88 0.968 0.279 0.401Rz (mm) 3.82 4.85 2.62 3.92Rmax (mm) 4.41 5.25 3.51 6.92Rt (mm) 4.41 5.81 3.88 6.92

Table 7Readings of perthometer M2 for bearing sample

Parameter Reading 1 Reading 2 Reading 3 Reading 4

Sampling length (mm) 5.6 17.5 5.6 17.5Lc (mm) 0.8 2.5 0.8 2.5Ra (mm) 1.510 2.790 1.750 2.02Rz (mm) 9.80 18.6 10.5 13.0Rmax (mm) 12.7 22.2 16.4 14.3Rt (mm) 14.2 22.7 16.4 16.5

Table 9Radial clearances of bearings

Radial clearancesrecorded from

Bearing no. 1 (mm) Bearing no. 2 (mm) Bearing no. 3 (mm) Bearing no. 4 (mm)

Installed May 1997 After 3 years Installed May 1997 After 3 years Installed May 1997 After 3 years Installed May 1997 After 3 years

Port 0.75 1.58 0.725 0.9 0.675 1.4 0.8 1.05Stbd 0.7 1.6 0.775 0.9 0.625 1.15 0.8 0.95

Initial Oct 2000 Final Dec 2002 Initial Oct 2000 Final Dec 2002 Initial Oct 2000 Final Dec 2002 Initial Oct 2000 Final Dec 2002

Port 0.775 1.65 0.65 1.0 0.75 1.175 0.5 0.755Stbd 0.575 1.025 0.45 0.7 0.4 0.625 0.5 0.755

Initial May 2003 Final Sep 2004 Initial May 2003 Final Sep 2004 Initial May 2003 Final Sep 2004 Initial May 2003 Final Sep 2004

Port 0.525 0.775 0.275 0.53 0.75 0.9 0.5 0.67Stbd 1.025 1.64 0.7 0.8 0.625 1.28 0.575 0.85

Table 8Viscosity measurement readings

rpm First run Second run Third run Fourth run Fifth run

Dynamicviscosity(mPa s)

Shear stress(Pa)

Dynamicviscosity(mPa s)

Shear stress(Pa)

Dynamicviscosity(mPa s)

Shear stress(Pa)

Dynamicviscosity(mPa s)

Shear stress(Pa)

Dynamicviscosity(mPa s)

Shear stress(Pa)

150 0.974 0.736 0.954 0.721 0.952 0.720 0.952 0.719 0.954 0.727250 1.083 1.364 1.075 1.354 1.077 1.356 1.078 1.358 1.080 1.361500 1.510 3.805 1.504 3.790 1.5 3.779 1.502 3.783 1.504 3.789

H. Hirani, M. Verma / Tribology International 42 (2009) 378–390 387

film to form is 160–190. Bearing 1 continues to operate inboundary lubrication regime even at radial clearance of 0.4. Effectof increased radial clearance on regime of lubrication can be seenin graph shown in Fig. 8. Surface roughness parameter (l) below 3indicates bearing operates in boundary lubrication regime. Thisanalysis suggests that all four bearings operate under boundarylubrication regime and Eq. (10) can be employed to evaluate thewear constant K. The calculated values of K are listed in Tables 11and 12. The wear depth of bearings of both shaft lines of one shipfor three operational cycles has been obtained by subtractinginitial clearance from final clearance. However, referring toTables 11 and 12, it can be seen that value of K is not consistentfor different bearings within the same operational cycle and forsame bearing within different operational cycles. For bearing 1,the variation in K is largely dependent upon initially set radialclearance; for bearing 2, K is quite consistent irrespective ofinitial set radial clearance; and for bearings 3 and 4, K isinconsistent irrespective of initially set radial clearance.Hence, the average of 24 values of K listed in Tables 11 and 12(0.226(10#17m2/N) cannot be used to estimate the life of waterlubricated bearing. However, the results obtained using thisapproach indicates that bearing 1 is severely loaded and moreprone to wear damage.

The second approach for wear analysis of water lubricatedbearing is to first evaluate the load shared by asperities using Eqs.

(11) and (12) and employ the calculated load to determine thewear constant using Eqs. (8) and (9). To understand this approachlet us consider that total applied, 60089N as shown in Table 1, onthe bearing 1 is shared by fluid film and asperity contact.Determine the asperities pressure and calculate the load baredby asperities by area integration of asperity pressure. For bearings1, 2 and 4 under different speed and clearance conditions, asperityloads are listed in Table 13. The results of this table indicate thatasperity load will decrease with increasing rotational speed of theshaft. Further, the minimum value of wear constant equal to0.35(10#11mm2/N, the maximum value equal to 1.7(10#11

mm2/N, and average value equal to 1.0(10#11mm2/N have beenestimated. These results indicate more than 50% variation in thevalue of wear constant. One possible reason for such variation isthat the individual bearing (i.e. bearing 1, bearing 2) has beenanalyzed as a component that is subject to a particular load andspeed condition. In reality load shared by all four bearings iscompletely coupled and dynamic. Larger clearance in one bearingincreases the load on other bearings and vise-versa. Table 9 clearlyindicates that radial clearance for bearing 1 in first load cycle was200mm greater than the clearance provided in third load cycle.Similarly radial clearance for bearing 2 was 500mm greater in firstload cycle compared to clearance in third load cycle. Such highvariation in radial clearance drastically changes the load capacityof bearings.

In hydrodynamic journal bearing often radial clearance (C) inthe range of 0.00075 to 0.001 times shaft radius is preferred. Theradial clearances shown in Table 9 demonstrate that two to fourtimes higher clearance was provided for water lubricated bear-ings. The load capacity of such bearings is proportional to 1/C2

therefore load capacity can be increase by decreasing radialclearance, which is shown in Table 14. This table indicates that ifradial clearance of bearing is approximately equal to 0.001 timesshaft radius, then there is significant decrease in asperity load.Hence, it can be concluded that major problem in elastomericbearing used in the Indian Coast Guard Ships is radial clearance ofthe bearing. Radial clearance between 0.00075 and 0.001 timesshaft radius should be maintained for longer life of all fourbearings.

7. Conclusions

Unequal and non-uniform wear of elastomeric bearings, usedin Indian Coast Guard Ships, have been analyzed. Following

ARTICLE IN PRESS

0

100

200

300

400

500

600

0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9Radial Clearance (mm)

RP

M

Brg 1 Brg 2 Brg 3 Brg 4

Fig. 7. Rpm below which bearing performs under boundary lubrication.

Table 10Exploitation pattern of ship for 03 operational cycles

Description May 97–Jun 2000 (H) Oct 2000–Dec 2002 (H) May 2003–Sep 2004 (H)

Total engine running hours 4109 5129 3264Total engine running hours in clutched condition 3800 4714 3198Number of starts and stops 304 191 122Hours at engine rpm 500 (shaft rpm 130) 549 – –Hours at engine rpm 550 (shaft rpm 143) 67 368 366Hours at engine rpm 600 (shaft rpm 156) 140 15 25Hours at engine rpm 650 (shaft rpm 170) 472 613 189Hours at engine rpm 700 (shaft rpm 182) 1709 2447 1677Hours at engine rpm 750 (shaft rpm 195) 618 28 –Hours at engine rpm 800 (shaft rpm 209) 177 864 753Hours at engine rpm 850 (shaft rpm 221) 64 375 173Hours at engine rpm 900 (shaft rpm 235) 01 04 14Hours at engine rpm 950 (shaft rpm 247) 02 – 01Hours at engine rpm 975 (shaft rpm 254) 01 – –

H. Hirani, M. Verma / Tribology International 42 (2009) 378–390388

conclusions can be drawn from the experimental and theoreticalstudy presented in the paper:

) Sea water shows thixotropic behaviour and its dy-namic viscosity increases with increase in rotationalspeed.

) Particle size of sea-water contamination is very small(40–120nm). Such small size of particles does not causebearing wear.

) All four elastomeric bearings, used to support propeller shaft,operated in mixed lubrication at operating conditions providedin Table 10.

) All four bearings have unplanned excessive radial clearance,which reduce their load capacity and result uneven rapid wear.

) Bearing life can be enhanced by proper selection of radialclearance for all four bearings.

Acknowledgement

Our sincere thanks to Indian Coast Guard Ship Sangram, fromwhere valuable and dependable data of shaft clearance and enginerunning hours were provided.

ARTICLE IN PRESS

@ 150 RPM

0

1

2

3

4

0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9Radial Clearance (mm)

0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9Radial Clearance (mm)

Bearing 1 Bearing 2 Bearing 3 Bearing 4

Bearing 1 Bearing 2 Bearing 3 Bearing 4

@ 185 RPM

0

1

2

3

4

5

Sur

face

Rou

ghne

ssP

aram

eter

S

urfa

ce R

ough

ness

Par

amet

er

Fig. 8. Effect of increased radial clearance on surface roughness parameter.

Table 11coefficient of wear (port shaft)

Bearingnumber

Coefficient of wear K (m2/N) ((10#17)

Operational cycle1

Operational cycle2

Operational cycle3

1 0.506 0.366 0.0922 0.187 0.231 0.2123 0.159 0.255 0.0794 0.198 0.024 0.101

Table 12Depth of wear and coefficient of wear (Stbd Shaft)

Bearingnumber

Coefficient of wear K (m2/N) ((10#17)

Operational cycle1

Operational cycle2

Operational cycle3

1 0.587 0.15 0.3542 0.090 0.134 0.053 0.47 0.098 0.7174 0.092 0.024 0.247

H. Hirani, M. Verma / Tribology International 42 (2009) 378–390 389

References

[1] Allais B. Calculations of bending moments and bearing reaction calculation.A Technical Note from M/S acbLIPS company, 1991.

[2] Stachowiak GW, Batchelor AW. Engineering tribology. 2nd ed. USA: Butter-worth-Heinemann; 2001.

[3] Khonsari MM, Booser RE. Applied tribology. Wiley; 2001.[4] Hirani H, Rao TVVLN, Athre K, Biswas S. Rapid performance evaluation of

journal bearings. Tribol Int 1997;30:825–34.[5] Johnson KL, Greenwood JA, Poon SY. A simple theory of asperity contact in

elastohydrodynamic lubrication. Wear 1972;19:91–108.[6] Kraker Alex de, van Ostayen Ron AJ, Rixen Daniel J. Calculation of stribeck

curves for (water) lubricated journal bearings. Tribol Int 2007;40:459–69.[7] Bayer RG. Engineering design for wear. 2nd ed. New York, USA: Marcel Dekker

Inc.; 2004.[8] Massimo Del Din, Kassfeldt Elisabet. Wear characteristics with mixed

lubrication conditions in a full scale journal bearing. Wear 1999;232:192–8.

[9] Hsu SM, Munro RG, Shen MC, Gates RS. Boundary lubricated wear.Wear—materials, mechanisms and practice. New York: Wiley; 2005.

[10] Rao AR, Mohanram PV. A study of wear characteristics of journalbearings operating under mixed-lubrication conditions. Wear 1994;172:11–22.

[11] Safar ZS. Energy loss due to misalignment of journal bearings. Tribol Int1984;17:107–9.

[12] El-Butch AM, Ashour NM. Transient analysis of misaligned elastic tilting-padjournal bearing. Tribol Int 2005;38:41–8.

[13] Jakeman RW. Non-linear oil film response model for the dynamicallymisaligned sterntube bearing. Tribol Int 1989;22:3–10.

[14] Sun J, Changlin G. Hydrodynamic lubrication analysis of journal bearingconsidering misalignment caused by shaft deformation. Tribol Int 2004;37:841–8.

[15] Mokhtar MOA, Howarth RB, Davis PB. The behaviour of plain hydro-dynamic journal bearings during starting and stopping. ASLE 1977;20(3):183–90.

ARTICLE IN PRESS

Table 13Asperity load and wear constant for bearings 1, 2 and 4

rpm Load shared by asperities of bearing 1, Wa (N) Load shared by asperities of bearing 2, Wa (N) Load shared by asperities of bearing 4, Wa (N)

Cycle 1 Cycle 2 Cycle 3 Cycle 1 Cycle 2 Cycle 3 Cycle 1 Cycle 2 Cycle 3

130 52467 – – 20887 – – 20708 –143 51767 51942 49 081 20341 19764 13307 20242 17778 17778156 51098 51294 48 199 19784 19142 12320 19811 17116 17116170 50349 50582 47215 19228 18554 11259 19244 16 412 16 412182 49 658 49 951 46 421 18732 17995 10342 18813 15831 15831195 48 951 49 279 – 18185 17452 – 18342 15208 –209 48257 48 500 44543 17606 16 809 8556 17840 14507 14507221 47574 47922 43674 17125 16260 7715 17408 13937 13937235 46 822 47135 42693 16502 15707 6647 16923 13248 13248247 46125 – 41885 16029 – 5925 16 470 – 12671254 45690 – – 15704 – – 16207 – –

Wear constant (mm2/N) ((10#11)1.7 1.43 0.35 0.44 1.0 1.6 0.76 0.73 0.49

Table 14The effect of clearance reduction on asperity load

Description May 97–Jun 2000(h)

Load shared by asperities of bearing 1, Wa (N) Load shared by asperities of bearing 4, Wa (N)

Bearing clearance(0.75mm)

Bearing clearance(0.175mm)

Bearing clearance(0.8mm)

Bearing clearance(0.175mm)

Total engine running hours in clutchedcondition

3800

Hours at engine rpm 500 (shaft rpm 130) 549 52467 36933 20708 8807Hours at engine rpm 550 (shaft rpm 143) 67 51767 35032 20242 7477Hours at engine rpm 600 (shaft rpm 156) 140 51098 33669 19811 6176Hours at engine rpm 650 (shaft rpm 170) 472 50349 31157 19244 4811Hours at engine rpm 700 (shaft rpm 182) 1709 49658 29736 18813 3667Hours at engine rpm 750 (shaft rpm 195) 618 48 951 27905 18342 2452Hours at engine rpm 800 (shaft rpm 209) 177 48 257 25972 17840 1172Hours at engine rpm 850 (shaft rpm 221) 64 47574 23877 17408 366Hours at engine rpm 900 (shaft rpm 235) 01 46 822 22109 16923 0Hours at engine rpm 950 (shaft rpm 247) 02 46125 20 044 16 470 0Hours at engine rpm 975 (shaft rpm 254) 01 45690 18304 16207 0

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