Tribology International 58 (2013) 12–19

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Tribology International

0301-67

http://d

n Corr

E-m

adateph1 Pr

Energy

mobile:

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

An investigation of tribological behaviors of dynamically loaded non-groovedand micro-grooved journal bearings

Hakan Adatepe a,1, Aydın Bıyıklıoglu b, Hasan Sofuoglu b,n

a Department of Mechanical Engineering, Aksaray University, Aksaray, Turkeyb Department of Mechanical Engineering, Karadeniz Technical University, Trabzon 61080, Turkey

a r t i c l e i n f o

Article history:

Received 18 December 2011

Received in revised form

13 September 2012

Accepted 13 September 2012Available online 26 September 2012

Keywords:

Micro-grooved bearings

Dynamic load

Friction force

Coefficient of friction

9X/$ - see front matter & 2012 Elsevier Ltd.

x.doi.org/10.1016/j.triboint.2012.09.009

esponding author. Tel.: þ90 462 3772931; fa

ail addresses: hakan.adatepe@giresun.edu.tr,

akan@yahoo.com (H. Adatepe), sofuoglu@ktu

esent address: Giresun University, Faculty of

Systems Engineering, Turkey. Tel.: þ90 454

þ90 505 899 23 85.

a b s t r a c t

Tribological performances of non-grooved and micro-grooved journal bearings were studied under

dynamic loading. Numerous experiments were performed using purpose-built test rig and then

simulated using various numerical methods. Friction force, friction coefficient, shaft center orbit, and

film thickness were determined experimentally and numerically. The experimental and numerical

results were in good agreement and the friction forces progressively increased on plain and

circumferential, herringbone, and transversally micro-grooved bearing. The results show that it is

necessary to complete detailed investigation about the tribological properties of the micro-grooved

journal bearing by taking their shape, depth and operating condition into account.

& 2012 Elsevier Ltd. All rights reserved.

1. Introduction

It is extremely important for journal bearing designers to knowtribological performances of journal bearings used in a number ofmachines such as internal combustion engines, jet engines, com-pressors, piston pumps, mechanical presses, and rolling millsduring the design process. The journal bearing surfaces encoun-tered in many studies were assumed to be smooth. However, thepossibility of improving bearing performance by modifying bearingsurface geometry has attracted attention of many researchers andthey have performed several theoretical studies on hydrodynamiclubrication field for rough journal bearing surfaces in recentyears [1]. Zhang and Qiu [2] conducted a theoretical investigationon effects of geometric structure of journal bearing surfaces underdynamic loading and hydrodynamic lubrication conditions. Theyinvestigated effects of surface roughness of dynamically loadedjournal bearings on longitudinal, transversal and isotropic basis. Intheir analysis, they used a statistical method (Stochastic Model)that was based on estimation principles developed by Christensenet al. [3,4]. They found that the maximum oil film pressure on thetransversally rough bearing was higher than those of the long-itudinal and isotropic rough conditions. They also concluded that

All rights reserved.

x: þ90 462 3773336.

.edu.tr (H. Sofuoglu).

Engineering, Department of

215 02 80;

oil film thickness for an isotropic rough condition was lessthan those of the transversal and longitudinal rough bearings.Hata et al. [5] investigated effects of frictional characteristics onsurface roughness under mixed and hydrodynamic lubricationconditions by using a test device that operates on pin-discprinciples. The roughness on bearing surface was cut precisely ina triangular and trapezoidal shape. It was found that the friction ontransversally grooved surface was higher than the friction on plainand longitudinally grooved bearing surfaces. Nakahara et al. [6]performed a study to measure effects of surface roughness onfriction characteristics after cutting regular and irregular threadson rectangular specimens spinning on a small axis. They also useda test device that operates on pin-disc principles. They found thatthe results obtained from the experiments correlated with theore-tical results. They also came to conclude that the effects of rough-ness on transverse rough surface were greater than those on thelongitudinal rough surface. Nakahara [7] then remarked that it wasextremely difficult to achieve accurate measurements for surfaceroughness on hydrodynamic lubrication characteristics. This wasrelated to the difficulties in measuring film layer thicknessesamong rough surfaces.

A new generation of bearings with micro-grooves has now beenused particularly in automotive engines operating at extremeconditions since last two decades. The studies about performancesof micro-grooved bearings have been, theoretically, shown thatcircumferential micro-grooves on journal bearings enhanceddynamic characteristics of bearings [8]. Moreover, it was alsoshown that circumferentially grooved bearings were stronger interms of resistance against deformation and wear as compared

Nomenclature

R bearing radius [m]P oil film pressure [N/m2]h oil film thickness [m]D bearing diameter [m]L bearing length [m]B bearing width [m]C radial clearance [m]F bearing load [N]C radial clearance [m]

M mobility vector [N-m]Z oil viscosity [Ns/m2]o bearing rotational speed [1/s]V circumferential velocity [m/s]e eccentricityg bearing’s width to diameter ratio, (B/D)e eccentricity ratio, (e/C)a eccentricity ratio in x directionb eccentricity ratio in y directionz axial coordinatey circumferential coordinate

1

2

3

456

7

8 9

18 17 16 15 14 13

12

11

10

1. Fixing ring 7. Conic ring 13. Spherical joint 2. Journal 8. Positioning element for bearing 14. O- Ring 3. Test bearing 9. Pulley 15. Piston 4. Bearing cap 10. Optic sensor 16. Nut- Ring 5. Screwed ring 11. Bearing housing 17. Servo-vane connection 6. Bushing 12. Loading shoe 18. Main structure support plate

Fig. 1. Cross-sectional view of the test rig along journal axis [16–18, 20].

H. Adatepe et al. / Tribology International 58 (2013) 12–19 13

to plain journal bearings. In a study conducted by Hargreavesand Armatys [9], the performances of different shapes ofmicro-grooved-journal bearings, transversal and longitudinal, underdifferent static and dynamic loads were investigated. A sample ofdynamic bearing load in a near sinusoidal form and on a single axiswas applied on a bearing by using a cam mechanism. The variationsof frictional moments within liquid friction zone under differentstatic loads suspended on the bearing and at operating speeds of400–2000 rpm were determined. They found that the journalbearings with circumferential grooves had lower frictional momentsas compared to the other shapes at high operating speeds bearings.Watanabe et al. [10] developed the high performance micro-groovedengine bearings by cutting circumferential grooves on plain journalbearings. They also determined the performances of the micro-grooved bearings by applying hydrodynamic and elastohydrody-namic lubrication theory.

According to the recent studies, plain cylindrical journal bear-ings with grooves were used extensively in industry to distribute oilover the entire surface of the bearings and to obtain optimumperformance as mentioned above [11,12]. It was shown that as theoil flow rate increased, the bearing temperature in the micro-grooved bearings became less than that of the bearings withoutmicro-grooves and as the oil remained in the grooves, the possibi-lity for the bearings to fade was low. Moreover, it was found thateffects of micro-grooves had direct impacts on journal bearingsperformances under dynamic loading conditions. It was also statedin the literature that the minimum oil film thickness of micro-grooved bearings was thicker than that of the traditional plainbearings due to the oil being retained in the grooves. Finally, havingmicro-grooves on the surface of journal bearings was proven to bean effective method to enhance the tribological behavior of thejournal bearings under starved lubrication conditions.

However, there has been little discussion on the shape of grooveand the operating conditions, such as friction zones. It became anecessary problem to discuss the shape of the groove and operat-ing conditions of the journal bearing. The objective of this studywas, therefore, to experimentally and theoretically investigate andcompare the tribological behavior of purpose-made micro-groovedand non-grooved (plain) journal bearings loaded dynamically. Inorder to achieve this objective, the plain journal bearings were firsttested at full film lubrication zone by utilizing the journal bearingtest rig under dynamic bearing load. The experiments were thenrepeated for the circumferential, transversally and herringbone(V-shaped) micro-grooved journal bearings for the same bearingparameters to present the effects of different shapes of micro-grooves on the performance of a typical engine crankshaft mainbearing. Later on, a commercial software ORBIT, developed foranalyzing dynamically loaded journal bearings using the mobilityand the finite difference methods [13], was used to numericallysimulate the experiments. Moreover, an additional computationalprogram which operates on the Schaffrath method [14,15]

was utilized to test the tribological behavior of journal bearing.Finally, the experimental results were compared with the numericalpredictions.

2. Experimental background and test procedure

In this study, the purpose-built laboratory test rig was utilizedto investigate tribological behavior of the dynamically loadedengine journal bearing. It uses a direct method where the frictiontorque of only the test bearing is measured without any inter-ference of the shaft-supporting bearings. The test rig wasdesigned to measure friction force under dynamic loading condi-tions by Bıyıklıoglu [16] and then modified to measure the orbitof the journal center under dynamic load by Bıyıklıoglu et al.[17–19]. Fig. 1 shows the cross-sectional view of the test rig alongthe shaft axis while the measurement system constructed spe-cially to measure friction force is illustrated in Fig. 2. The detailed

F

n

Tension bar (23)

Lubricating oil

Squeezed oilJournal bearing (3)

Bearing cap (4)

Bushing (6)

Conic ring (7)

Bearing housing (11)

Journal (2)

Fixing ring (1)F

Fig. 2. Schematic representation of the developed friction force measurement

system [16–18, 20].

34

51.8

8047

.750

3.5

1

3

Fig. 3. The shapes and dimensions of the journal bearing.

Fig. 4. Photographs of (a) plain, (b) circumferential, (c) her

H. Adatepe et al. / Tribology International 58 (2013) 12–1914

information about the test rig can be obtained from the abovementioned references and [20].

The journal was made of AISI 1070 hardened steel with RockwellC hardness of 55 and supported at each end by bronze sup-port bearings with a radial clearance of 0.025 mm. The bearings ofS-844-GS5049 M series used in the experiments were supplied bymanufacturer located in Turkey. The geometry and the dimensionsof the bearings were shown in Fig. 3 while Fig. 4 illustrated thephotographs of plain, circumferential, herringbone, and transversallygrooved journal bearings. The circumferential, transversally andherringbone grooves were fabricated onto the plain engine bearing.The circumferential grooves were made with a hard 601 anglethread cutting tool tip whereas the transversally and the herring-bone grooves were made onto the plain engine bearing by pinchingstraight and 451 inclined grooves via knurling wheel used forgeneral non-cutting forming purposes, respectively. The surfaceprofiles of the journal bearings were measured prior and after thetests by using Perthometer M1, surface roughness measuring device.The average depths of all micro-grooves were taken as 30 mm in thisstudy. For example, the journal bearing with circumferentialgrooves, the average roughness height (Ra) before the tests wasfound to be 7.815 mm while the average roughness depth was29.7 mm. On the other hand, the average roughness height of thepresumably plain and smooth journal bearing was 1.040 mm whilethe average roughness depth was 4.86 mm [21].

By using a belt-pulley system, the power taken from 5.5 kWmotor was transmitted to the driven shaft of diameter d¼47.7 mmrotated at 1240 rpm. When the journal was driven, the journalbearings housing also turned in the same direction as the journaldue to the friction between the test bearing and the rotatingjournal. The oscilloscope was included into the system to indicatethe time history of net load along each axis passing through twoopposing pairs of cylinders. The polar loads, two bearing

ringbone and (d) transverse grooved journal bearings.

H. Adatepe et al. / Tribology International 58 (2013) 12–19 15

load components (vertical and horizontal), and friction forceswere then recorded by a DSO 400 GOULD model oscilloscope.An optic sensor was used to measure the crank angle during theexperiments while four transducers were utilized to measurethe journal center trajectory. The frictional moment occurringbetween the journal bearings and bearing cap was measured byusing a Wheatstone bridge circuit via the strain gauges. A full-bridge circuit was set up by using four strain gauges. The signaltaken from the bridge circuit was amplified with a KWS/T-5model amplificatory and sent to a computer via a data acquisitionsystem. The conversion to a bending force was executed on thefour measurement beams on which strain gauges attached toboth sides gave a signal proportional to the frictional moment.The signal values were then converted into friction torque andfriction force values, which were determined using a calibrationline obtained from the strain gauges. Journal load, friction force,and displacement of journal center were measured from therecord by every 151 crank angle intervals.

The test bearings were run for 2 h in order to flush out weardebris during the running-in process under low speed and lowloading conditions. In order to provide temperature stability, thesystem was run for about 30 min and the lubricant temperaturereached 23 1C. The experiments were carried out with commercialbase oil.

Fig. 5. Dynamic bearing load views taken from the oscilloscope screen: (a) Polar

load, (b) horizontal and (c) vertical components.

3. Numerical background

ORBIT program was applied in analyses of dynamically loadedjournal bearings in this study. This computer aided engineeringprogram was first used to find solutions with the mobility methoddeveloped by Booker [22,23]. In the mobility method, the mobi-lity vector, M, which is the dimensionless squeeze velocity withwhich a non-rotating shaft moves to support the load, is calcu-lated by converting the effect of rotation to an equivalent squeezemotion under steady load. For a rotating shaft, the squeezevelocity (e) is given by [13]

e0 ¼9F9½C=R�2

LDZ=CMþoe ð1Þ

In dynamic analyses, the components of mobility vector (M)can generally be expressed as [13]

M¼ f ðe,gÞ ð2Þ

The analytical equations for mobility components obtained fromthe curve fitting of mobility components are expressed in thefollowing manners [13,24]:

Mx¼f ða,gÞpg2

ffiffiffiffiffiffi3pp�0:24g2eþ

b2

1�a ð1�0:4ffiffiffiffiffiffiffiffiffiffi1�ap

Þþa2gð4=3�gÞ)(

ð3Þ

My¼�bð1�aÞ f ða,gÞ

54 þ

ag7 þ

a2

8 ð1þaÞ�0:3g2 1þ a3

2ffiffi3p

� �þ

2b2

15ð1�aÞ

�0:016 1� 0:034g2

� �1

ð1:03�aÞ

9>=>;

8><>:

ð4Þ

f ða,gÞ ¼ ð1�aÞ21=2

pg21þ

g2

2ð1�aÞ

� �ð5Þ

The second method provided by this commercial software isbased on a direct finite difference solution of the Reynoldsequation [13]:

@

@z

h3

12Z@P

@z

!þ

1

R2

@

@yh3

12Z@P

@y

!¼

1

R

V

2

@h

@yþo2

@h

@zþ@h

@tð6Þ

with the following boundary conditions:

z¼ 0 Pðz,yÞ ¼ P1,j

z¼ L Pðz,yÞ ¼ Pn,j ð7Þ

The Schaffrath method [14,15] is the last computationalprogram used in this study to test performances of dynamicallyloaded engine journal bearing. This method calculates the frictionforce and the coefficient of friction by using the followingequations:

Fs ¼e

2RF sinbþB

Zc

ro 2pffiffiffiffiffiffiffiffiffiffiffi1�e2p þ2pZBR _g cos2g 1ffiffiffiffiffiffiffiffiffiffiffi

1�e2p �1

� �ð8Þ

m¼ e

2Rsinbþ

BZroFc

2pffiffiffiffiffiffiffiffiffiffiffi1�e2p þ

2pZBR

F_g cos2g 1ffiffiffiffiffiffiffiffiffiffiffi

1�e2p �1

� �ð9Þ

4. Results and discussion

Fig. 5 showed the dynamic bearing load views taken directlyfrom the oscilloscope screen, the polar load and the horizontaland vertical components of bearing load. It was clear from thefigure that the horizontal and vertical components of applied loadwere completely variable (dynamic) and act on the journalbearing in a sinusoidal form. In Fig. 6, the variations of horizontaland vertical components of applied dynamic bearing load withtheir resultant were given with respect to bearing’s crank angle.These figures clearly indicated that the maximum bearing loadwas reached its maximum value at 16.68 kN which correspondsto the crank angle of 97.51 while the minimum bearing load of172.5 N was obtained at the crank angle of 172.51.

Fx Fy Fsum

Fig. 6. Variation of the resultant applied bearing load with vertical and horizontal

components as a function of crank angle.

Fig. 7. Variations of measured friction forces formed on oil film of journal bearing

for (a) transversally, (b) herringbone, and (c) circumferential grooved and (d) plain

journal bearings.

Fig. 8. Variations of friction force for four different bearing types under the same

dynamic bearing load.

Fric

tion

Coe

ffic

ient

, µ

Fig. 9. Variations of friction coefficients for four different bearing types under the

same dynamic bearing load.

H. Adatepe et al. / Tribology International 58 (2013) 12–1916

The distributions of measured friction forces taken directlyfrom the oscilloscope screen during the test under dynamicbearing load for all bearings tested in this study were given inFig. 7 while Fig. 8 illustrated their variations against the crankangle. As concluded from Fig. 8 that the highest friction force of21.12 N was obtained at 127.51 crank angle for the transversallygrooved bearing followed by 20.02 N at 112.51 crank angle forherringbone grooved bearing, 19.54 N at 67.51 crank angle forcircumferential grooved bearing, and 11.63 N at 82.51 crank anglefor plain bearing. The average increase in the friction force basedon to the plain journal bearing was obtained to be approximately1.81, 1.72, and 1.68 times for transversally, herringbone, andcircumferentially grooved bearings, respectively. It was observedfrom Figs. 7 and 8 under the dynamic loading condition that the

effect of groove’s shape on the friction force varied considerablywith the crank angle. Both figures clearly indicated that thelowest friction force was obtained from the non-grooved journalbearing while the highest friction force was measured for thetransversally grooved journal bearing as opposed to enhancingperformances of the micro-grooved journal bearing given in theliterature. The reason for the higher frictional values occurred onthe micro-grooved bearings can be attributed to a higher disrup-tion of the oil flow on the micro-grooved bearings as compared tonon-grooved bearing.

Fig. 9 depicts the variation of coefficients of friction with crankangle for four types of bearings (plain and grooved) under theeffect of dynamic bearing load. It was seen from this figure thatthe smallest value of the coefficient of friction was exhibited onthe plain journal bearing while the largest one was obtained fromthe transversal grooved journal bearing similar to the distributionof the friction force, as expected. The maximum values of coeffi-cient of friction obtained from the experiments were 0.0036,0.0024, 0.0021 and 0.0015 for the transversally, the herringbone,and the circumferential grooves, and for the plain journalbearings, respectively. The increasing ratios for the coefficient of

Fric

tion

Coef

ficie

nt, µ

Fig. 11. Variation of friction coefficient under the same dynamic bearing load

obtained from numerical and experimental analyses.

1 20°

H. Adatepe et al. / Tribology International 58 (2013) 12–19 17

friction with respect to the plain journal bearing were approxi-mately 1.4, 1.6 and 2.4 for the circumferential, the herringbone,and the transversally micro-grooved bearings, respectively.These numbers indicated that all the experiments on the journalbearings took place within the liquid friction zone. Therefore,Fig. 9 illustrated that with the same dynamic load applied, themicro-grooved journal bearings yielded a higher coefficient offriction than that of non-grooved bearing as opposed to thosegiven in literature.

This disagreement can be attributed to the fact that studieswere performed in different working conditions. For example,motors work in mixed lubrication zone during startup andstopping motions in which metal-to-metal contact could be morepossible to occur due to starved lubrication conditions. Themicro-grooves, in that case, help oil to easily lubricate the surfaceof the journal bearing in mixed lubrication zone. It can, therefore,be concluded that micro-grooves on the surface of journal bear-ings can be used to improve the tribological behavior of thejournal bearings under starved lubrication conditions. On theother hand, the experiments and the numerical studies, in ourcase, were pursued in the full film lubrication (the liquid friction)zone so that there was no metal-to-metal contact and the micro-grooves struck the lubricating oil to flow on the surface of micro-grooved bearings. It can, therefore, be deduced that the operatingcondition plays an important role on the performance of micro-grooved journal bearing.

When Figs. 6, 8, and 9 were evaluated together, it could beconcluded that the friction force and the coefficient of frictionreached to peak values at the crank angle in which the bearingload was also maximum (in the vicinity of 1001). Since the bearingload is dynamics, it is possible to have different crank angles forthe peak values of the friction force and the coefficient of friction.

As for the theoretical study, the tribological behavior of theplain journal bearing was determined by using ORBIT softwarewhich utilizes two numerical methods, the mobility method andthe finite difference methods. Moreover, another computationalsolution was determined by using the Schaffrath method.The friction forces obtained numerically and experimentally werethen plotted at the same figure for comparison purpose, shown inFig. 10. It was found that the numerical and the experimentalresults obtained in the full film lubrication zone were in goodagreement. It was seen that the highest friction force was 9.94 Nat 1001 crank angle obtained by the mobility method while it

Fig. 10. Variation of friction force under the same dynamic bearing load obtained

from numerical and experimental analyses.

was 10.16 N at 1001 crank angle for the finite difference method.On the other hand, the highest friction force obtained bythe Schaffrath method at 1501 crank angle was 11.29 N while11.63 N was determined for the plain bearing at 82.51 crankangle.All results for the coefficient of friction obtained from theexperimental and the numerical studies were then demonstratedin Fig. 11. It was clear from the figure that the values for thecoefficients of friction were very close to each other, especially forthose obtained from the mobility and finite difference methods,as well as the corresponding crank angles (between 1501 and1701). The exception was occurred for the maximum value of thefriction coefficient obtained by the Schaffrath method, similar to

0.2 0.4 0.6 0.8 1.0

43

Mobility Method Finite Difference Method Schaffrath Method Polar Load

270°

360°

30.000 N

90°

360°

180°

270°

90°

180°

360°

90°

180°

270°

180°

270°

Fig. 12. Polar load diagram for the plain journal bearing and journal center orbits

obtained from different numerical methods.

Fig. 15. Variation of minimum oil film thicknesses with respect to crank

and position angles.

H. Adatepe et al. / Tribology International 58 (2013) 12–1918

the case in the friction force. It is normal to obtain the peak valuesof the friction force and the coefficient of friction at the differentcrank angles due to the nature of the dynamic bearing load. How-ever, although the crank angles corresponding to the peak frictionforces differed from each other, the variations of the friction forcethrough one cycle of the crank angle looked similar; i.e., increas-ing and decreasing regions almost coincided. This was held truefor the coefficient of friction more strongly.

The orbits drawn by the shaft center based on three differentcomputational methods and polar load diagram for the dynamicallyloaded plain journal bearings were shown in Fig. 12. The figureshowed that the orbits drawn by the shaft center complied well witheach other. The small discrepancies might be caused by the differentcomputational method used. It was also seen from the figure that themaximum eccentricity ratio existed for two methods, namely;mobility and finite difference methods, was completely coincided atthe position angle of 1871 with a value of 0.76 while the Schaffrathmethod had a very small divergence from these two methods havingthe value of 0.79. The graph plotted in order to determine therelationship between the experimentally obtained coefficients offriction on the plain journal bearing and the minimum oil filmthickness obtained from different numerical methods was given in

Fig. 14. Variation of minimum oil film thickness with respect to bearing load.

Fig. 13. Variation of coefficients of friction and minimum oil film thickness

obtained from different numerical methods.

Fig. 13. This figure illustrated that when the oil film thicknessdecreased the coefficient of friction increased and wherever thecoefficient of friction reached to maximum value the oil film thick-ness was attained to minimum value. In Fig. 14, the variation ofminimum oil film thickness against the bearing load was shown.It was demonstrated that the minimum oil film thicknesses werefound to be 6.17 mm, 6.11 mm, and 5.04 mm by utilizing the mobility,the finite difference and Schaffrath methods, respectively. They allexisted at around crank angle of 1411. Moreover, it was seen that thecrank angle position of the minimum oil film thickness, was justoccurred 43.51 beyond to that of the maximum bearing load (97.51).The variation of minimum oil film thickness versus the crank andthe position angles was illustrated in Fig. 15. This figure showed thatthe smallest oil film thickness was just occurred at the position angleof 1861.

5. Conclusions

Before designing micro-grooved bearings, it is necessary topredict performances of the micro-grooved bearings accurately.We, therefore, investigated the effects of micro-grooves on thetribological behavior of the journal bearing within liquid frictionzone using the purpose-built friction test rig in this study. Thegrooves in micron scale basis were made onto the plain journalbearings as circumferential, transversal and herringbone shapesand they were then subjected to dynamic load. This investigationyielded the following results:

1.

The experiments showed that the lowest value of friction forcewas determined on the plain journal bearing. The experimen-tally obtained values increased progressively for the circum-ferential and the herringbone micro-grooved bearings and,finally, the highest frictional force distribution was exercisedon the transversal micro-grooved bearing.

2.

The shape of micro-groove affects the coefficient of friction,attaining the lowest value to the plain journal bearing andresulting in the maximum value for the transversally micro-grooved journal bearing, similar to the trend of the friction force.

3.

The friction force distributions as well as the variations in thecoefficient of friction obtained from the different numericalmethods for the plain journal bearing were very similar andagreed well with each other and also with those of theexperimentally obtained results.

H. Adatepe et al. / Tribology International 58 (2013) 12–19 19

4.

The minimum oil film thickness value appeared at the furthestposition from the point of the maximum bearing load valueexisted.

In conclusion, the results of this research showed that theshapes of micro-grooves played an important role on the oil filmthickness, the coefficient of friction, and the friction force underthe dynamic load. This study further indicated that the micro-grooves developed the worse tribological properties on thejournal bearings when they operated in the liquid friction zonethan the non-grooved journal bearings as opposed to knowledgewritten in the literature. It can, therefore, be concluded that theuse of generalized results regardless of shape of micro-groove andoperating condition may result in erroneous estimates of thetribological properties of the micro-grooved journal bearing. It is,therefore, recommended that performing detailed investigation inconjunction with shape, depth and operating condition is neces-sary to obtain reliable data regarding the tribological properties ofthe micro-grooved journal bearing.

Acknowledgment

This work was supported by the Research Fund of KaradenizTechnical University, Project no. 2002.112.003.03.

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