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Wear characteristics with mixed lubrication conditions in a full scale journal bearing
Wear 232 Ž1999. 192–198 www.elsevier.comrlocaterwear
Wear characteristics with mixed lubrication conditions in a full scale journal bearing Massimo Del Din a , Elisabet Kassfeldt b
b,)
a ABB Stal AB, Dept. TMS, SE-612 82 Finspang, ˚ Sweden DiÕision of Machine Elements, Lulea˚ UniÕersity of Technology, SE-971 87 Lulea, ˚ Sweden
Abstract Increased awareness of environmental problems has stressed the importance of switching from traditional lubricants to more environmentally friendly alternatives. Different investigations with standard test methods indicate that such a switch is possible without loss of lubricating power. The question arises if the statement above is true for industrial lubricating conditions, without laboratory cleanliness. This paper presents a study of friction and wear in a two grooved journal bearing at different shaft speeds, oil temperatures and contamination levels. A number of tests have been conducted combined with a theoretical analysis of film thickness and lubricating regime. The aim was to investigate whether an environmentally adapted rape seed-synthetic ester oil could replace a traditional mineral oil in a full scale application. The results show that the rape seed-synthetic ester oil gives significantly lower values of wear regardless of the operating conditions and there is also a tendency of lower values of frictional torque compared with the mineral oil. Mainly operating in the mixed lubrication regime, no signs of impending bearing failure have been registered, even though a large amount of silica particles added to the oil gave higher wear values than with uncontaminated oil. q 1999 Elsevier Science S.A. All rights reserved. Keywords: Wear characteristics; Journal bearing; Mixed lubrication; Environmentally adapted lubricant
1. Introduction In today’s society there is an increased need to take an environmental point of view of our day to day activities for both people in general and companies as well as for governmental institutions. Awareness on what pollution does to the environment is widely spread and this has forced actions to be taken in order to reduce the amount of environmentally harmful waste. According to Ahlbom and Duus w1x, and Bartz w2x a considerable amount of the annual lubricant consumption is likely to end up in nature instead of being properly disposed of. Answering the demands for environmentally adapted products, lubricant manufacturers have developed lubricants that do not affect people handling them or nature where the waste eventually ends up. If however the new environmentally adapted lubricants are to replace traditional oils, the lubrication performance of the former must be no worse than the latter. Several investigations of lubrication performance
) Corresponding author. Tel.: q46-920-912-40; fax: q46-920-910-47; E-mail: [email protected]
have been carried out. According to Rieglert w3x, environmentally adapted hydraulic fluids can give lower friction values than a mineral base oil and Kabuya and Bozet w4x found that vegetable oils can easily match mineral oils from a lubrication point of view. 1.1. Aim of the study The aim of the investigations performed was to compare the wear and friction performance of a fully formulated environmentally adapted oil with an equally formulated traditional mineral oil, in a full scale journal bearing application. The work was primarily focused on the slow
Table 1 Operating conditions Designation
Rotational speed wrpmx
Oil temperature w8Cx
Operating time whx
Load wkNx
1 2 3 4
100 100 10 10
70 25 70 25
48 48 48 48
35 35 35 35
0043-1648r99r$ - see front matter q 1999 Elsevier Science S.A. All rights reserved. PII: S 0 0 4 3 - 1 6 4 8 Ž 9 9 . 0 0 1 4 5 - 3
M. Del Din, E. Kassfeldtr Wear 232 (1999) 192–198 Table 2 Physical properties for test oils
193
2.1. Experimental design and oil properties
Notation
A
B
Base oil
Mineral
Viscosity q258C wmPasx Viscosity q708C wmPasx Viscosity index Density q158C wkgrm3 x
52 9 101 874
Rape seedsynth. ester 65 15 213 928
rotational speed associated with a bearing operating in the boundary lubrication regime.
The parameters altered during the investigation were oil type, oil temperature, rotational speed, and contamination level. All tests were performed both with uncontaminated oil, that is oil directly from the manufacturer, and with oil that has been contaminated with silica particles. The particle size covered a range from infinitely small to a maximum of 50 mm, with the majority between 10 and 20 mm. The duration time was 48 h for all the tests performed. The operating conditions are presented in Table 1. The physical properties of the two oils used in the tests are presented in Table 2. Both oils are fully formulated and can be bought from an ordinary oil retailer.
2. Investigation method 2.2. Experimental equipment An investigation was carried out to compare wear and friction of bearing liners in a two grooved journal bearing with different oil types when operating at various oil temperatures, shaft speeds and contamination levels. The investigation was both experimental, with several tests, and theoretical with calculations of oil film thickness and lubrication regime. The tests were carried out on a full scale test rig with the intention of being as close to real operating conditions as possible.
The bearing used in the test can normally be found in large synchronous machines. The full scale test rig consisted of four main parts; a frame, a two groove journal bearing, a large shaft and an electric motor ŽFig. 1.. The frame was constructed from welded steel beams and carried the shaft supported by two spherical roller bearings and connected to the motor via a belt transmission. The test bearing was mounted on the shaft and the
Fig. 1. Test rig.
M. Del Din, E. Kassfeldtr Wear 232 (1999) 192–198
194
During a test the temperature of the liner and the oil were measured using two thermocouples and these data were continuously recorded. The rotational speed of the shaft was measured with a tachometer and the speed controlled with velocity feedback. 2.3. Friction torque measurements In order to measure the friction torque a horizontal bar was connected to the upper side of the bearing and to a vertical bar equipped with two strain gauges ŽFig. 2.. Before each test the strain gauges were calibrated using weights and this made it possible to calculate a friction torque.
Fig. 2. Torque measuring equipment.
load applied by a hydraulic cylinder fixed to the upper side of the bearing at one end, and to the upper most beam of the frame at the other end. Oil was supplied to the bearing by a pump from a tank and the oil temperature adjusted with the assistance of an electrical hot plate. While travelling from the tank to the bearing, the oil is filtered according to the recommendations of the bearing manufacturer. Experience from previous work indicates that the contamination level during a test never falls below 0.02 wt.% in spite of the filtration. For the low temperature tests a water-cooled copper pipe was lowered into the tank to prevent temperature increase. The bearing housing divided into two parts and a liner, made out of babbit, fitted in a spherical cup covered with Teflon which was then assembled into the housing.
2.4. Wear measurements In order to measure the wear of the bearing liners a number of methods was used. Each liner was weighed before and after the tests in order to calculate the weight loss. Liners were also examined with a perthometer and optically inspected. Liners with a wear pattern differing from the normal were also inspected with a microscope and examined with SEM. 2.5. Lubricating film thickness and coefficient of wear With the test rig used in this investigation there was no possibility of measuring the oil film thickness, h, and in order to estimate the lubricating conditions in the bearing,
Table 3 Complete list of results a
Oil
T w8Cx
n wrpmx
Debr.
Visc. wmPasx
e
h min wmmx
R a wmmx
l
m loss wgx
K U 10y7
Mf wNmx
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23
A A A A B B B B B B B B A A A A B B B A A A A
q70 q25 q25 q70 q70 q25 q25 q70 q25 q70 q25 q70 q25 q70 q25 q70 q40 q40 q25 q25 q25 q70 q25
100 100 10 10 100 100 10 10 100 100 10 10 100 100 10 10 100 10 100 100 10 100 100
y y y y y y y y q q q q q q q q q q y q q q q
9 52 52 9 15 65 65 15 65 15 65 15 52 9 52 9 36 36 65 52 52 9 52
0.975 0.880 0.985 0.997 0.952 0.852 0.981 0.994 0.852 0.952 0.981 0.994 0.880 0.975 0.985 0.997 0.920 0.989 0.852 0.880 0.985 0.975 0.880
3.5 16.8 2.1 0.4 6.7 20.7 2.7 0.8 20.7 6.7 2.7 0.8 16.8 3.5 2.1 0.4 11.2 1.54 20.7 16.8 2.1 3.5 16.8
0.374 0.496 0.504 0.481 0.511 0.504 0.328 0.359 0.711 0.862 0.267 0.259 0.458 0.389 0.198 0.114 0.711 0.473 0.427 0.603 0.298 0.359 1.12
6.4 26.4 3.3 0.6 10.3 32.2 5.2 1.5 25.4 7.1 5.6 1.7 27.6 6.3 4.7 1.0 13.7 2.5 35.4 23.2 4.2 6.5 14.1
0.09 0.02 0.07 1.13 0.00 0.00 0.02 0.13 0.21 0.04 0.57 0.00 2.25 0.84 1.88 0.21 0.01 0.19 0.00 – 0.96 0.86 2.34
4.7 1.0 36.3 586.2 0 0 10.4 67.4 10.8 2.1 295.7 0 116.7 43.6 975.2 108.9 0.5 98.5 0 – 498.0 44.6 121.4
176 271 200 205 181 123 85 160 215 215 184 297 200 190 172 374 246 139 188 262 146 270 276
T s Oil temperature; h min s Minimum film thickness; m loss s Weight loss of liner; n s Shaft speed; R a s Surface roughness; K s Coefficient of wear; e s eccentricity; l s Film thickness parameter; Mf s Friction torque.
M. Del Din, E. Kassfeldtr Wear 232 (1999) 192–198
195
Fig. 3. Weight loss vs. operating condition for uncontaminated oil. Fig. 5. Friction torque vs. operating condition for uncontaminated oil.
a theoretical approach was used. With the experience of Wikstrom ¨ et al. w5x, who showed that viscous effects are negligible, an isoviscous analysis of the bearing was made and the oil film thickness calculated. From the film thickness it was possible to determine the film parameter, l, which according to Hamrock w6x is defined as
ls
h
(R
2 2 q1 q R q2
Ž 1.
l s film thickness parameter; h s film thickness wmx; R q s equivalent surface roughness wmx. Hamrock suggests that the value of l has to be 5 or more to ensure full film lubrication but for journal bear-
Fig. 4. Weight loss vs. operating condition for contaminated oil.
ings higher values in the range 8 to 10 are commonly used. Because of the poor knowledge of how the surface roughness of a liner changes during a test, the calculated values of the film parameter presented should be used with caution. As R q is often very similar to R a the latter was used in the calculations. Surface roughness varies over the worn surface although from the surface measurements one value of R a , considered to reflect the main roughness, was determined. To obtain a comparable value of wear, the coefficient of wear, K, was used instead of weight loss. K was calculated using Archard’s equation of wear Ž2. which is dimen-
Fig. 6. Friction torque vs. operating condition for contaminated oil.
196
M. Del Din, E. Kassfeldtr Wear 232 (1999) 192–198
Fig. 7. Wear for uncontaminated oil.
sionless, which is an advantage when comparing tests with different sliding distance. V FN V H sK ´Ks Ž 2. S H S FN V s Wear volume wm3 x; S s Sliding distance wmx; K s Coefficient wear; H s Materials hardness wMPax; FN s Load wNx. 3. Results The results for wear and friction torque gave the general impression, with some exceptions, that oil B gives lower values than oil A in both cases. For tests where uncommonly high weight loss values have been recorded, the tests were repeated and a mean value for weight loss and friction torque used in the evaluation. The complete results can be seen in Table 3. 3.1. Wear
Fig. 9. Liner with a v-shaped pattern.
cles, leaving any embedded particles to be weighed with the liner. The scale used had an accuracy of "0.01 g. The temperature of the oil during the tests varied "48C and the rotational speed "1 rpm. As can be seen in Fig. 3, the weight loss of the liners for test with uncontaminated oil vs. operating condition, only minimal weight loss occurred, the highest weight loss values registered for the operating condition with a rotational speed of 10 rpm and an oil temperature of q708C. When comparing both oils, oil B gave the lowest wear values. As can be seen in Fig. 4, the weight loss for the liners tested with contaminated oil vs. operating condition, the wear values were higher than with uncontaminated oil but with oil B still giving the lowest values of wear. The remarkable low values at high temperature and low rotational speed may be a result of the fact that different wear mechanisms determines the wear at different operating conditions. When examining the liners optically, only mild wear could be detected, with some polishing and a few scratches
Before weighing the bearing liners they were cleaned with solvent to remove oil and loose contamination parti-
Fig. 8. Wear for contaminated oil.
Fig. 10. SEM recordings ŽThe white line to the right of the photo represents the length, L..
M. Del Din, E. Kassfeldtr Wear 232 (1999) 192–198
197
4. Discussion and conclusions
Fig. 11. Friction torque for uncontaminated oil.
for the uncontaminated oil and a greyish wear pattern for the contaminated oil. For some tests with the contaminated oil the worn areas formed equally divided stripes in the circumferential direction, Žsee Fig. 9.. This wear pattern was also reported by V. Wikstrom. ¨ For some tests with contaminated oil a v-shaped area starting in the central region of the liner could be seen near the outlet region. This has not been mentioned in previous investigations. 3.2. Friction torque The friction torque was measured using the method described in Section 2.3 and it was continuously recorded during each test. The friction torque showed little variation over time Ž"10%. and therefore a mean value was calculated for every test run. The friction torque vs. operating condition for contaminated and uncontaminated oil are given in Figs. 5 and 6, oil B showing lower values with two exceptions. Irrespective of the level of contamination, the results are reasonable even when comparing the oils at the same operating condition. However, the level of the friction torque values are higher for contaminated oil than for uncontaminated oil and the maximum torque is reached at different operating conditions.
The two oils had different ISO viscosity grades so to be able to compare them, both wear and friction torque are presented as a function of minimum oil film thickness. Film thickness and the film thickness parameter, l, were calculated following the method presented in Section 2.5 and the complete results presented in Table 3. The wear behaviour with uncontaminated oil, Fig. 7, was as could be expected with very low K values except for the thinnest of oil films. It should be noted that the lines joining the measurement points in the diagram are an aid to differentiate between the oils, not an assumption of the behaviour between them. The wear behaviour seems reasonable given that the major wear mechanisms is through asperity contacts. The vegetable oil B, appears to give the lower wear values of the two oils. For the contaminated oil, Fig. 8, maximum wear did not occur with minimum film thickness as was the case for the uncontaminated oil. This is probably due to a change in the wear mechanism. As was the case with uncontaminated oil, oil B gives the lower K values. Uncontaminated oil gave liners with a bright wear pattern, the bright area increasing with decreasing oil film thickness. The contaminated oil resulted in a more greyish liner surface, some showing the circumferential stripes mentioned in Section 3.1. In addition to the stripes, a v-shaped pattern occurred with some of the liners tested with an oil film thickness of 2.1–3.5 mm ŽFig. 9.. Looking at this wear pattern with SEM, Žsee Fig. 10., the v-shaped area did not appear as scratched as the surrounding area which had many parallel scratch marks. A possible explanation for this is that the combination of shaft bending, due to the applied force, and the curvature of the liner along with the hydrodynamically induced oil film allowed the contamination particles to move more freely towards the outlet region. Thus, the scratches were not as deep as in the greyish area. Irrespective of contamination level, oil B seems to give the lowest friction torque value, see Figs. 11 and 12; the wear pattern being the same for both oils with a dip for h min values of approximately 2–4 mm. The friction torque
Fig. 12. Friction torque for contaminated oil.
198
M. Del Din, E. Kassfeldtr Wear 232 (1999) 192–198
values follow the Stribeck curve which also has a similar dip for low h min values. This is often called the boundary lubrication regime. For both contaminated oil samples the friction has a maximum and the coefficient of wear a minimum. A possible explanation for this could be that for thin oil films the contamination particles are to large to pass between the two surfaces and therefore become embedded in the soft liner material and no wear is registered. In the region where maximum wear occurs, the oil film is of the right size to allow particles to pass without being embedded, tearing off material when rollingrsliding through the contact, and hence resulting in high K values. Thicker oil films show a low coefficient of wear, increasing slightly when higher rotational speed increase the number of abrasive particles passing through the contact region. From the study the following conclusions can be drawn, for the investigated operating conditions, oil type and materials: Ø The semi-synthetic oil led to significantly lower rates of wear than the mineral oil; Ø There is measurable tendency for the semi-synthetic oil to give lower values of friction torque than the mineral oil; Ø Neither of the oils has a higher risk of failure.
Acknowledgements The authors would like to thank their colleagues at the Division of Machine Elements for advice and valuable ¨ discussions and especially Mr. Osten Uusitalo for his skilful assistance during the work with the test rig.
References ¨ w1x J. Ahlbom, U. Duus, 1992. RENA SMORJAN? Smorjmedel ¨ mojligheter till forandring. Rapport fran ¨ ¨¨ ˚ Kemikalieinspektionen 8r92. w2x W.J. Bartz, Lubricants and the environment, Proc. of the World Tribology Congress, London, 1997. w3x J. Rieglert, Lubricating Performance of Environmentally Adapted Hydraulic Fluids, Licentiate Thesis, Lulea˚ University of Technology, Division of Machine Elements, Lulea, ˚ 1997. w4x A. Kabuya, J.L. Bozet, 1995. Comparative analysis of the lubricating power between a pure mineral oil and biodegradable oils of the same iso grade. Lubricants and Lubrication, p. 25–30. w5x V. Wikstrom, R. Larsson, Wear of bearing liners at low ¨ E. Hoglund, ¨ speed rotation of shafts with contaminated oil, Wear 162–164 Ž1993. 996–1001. w6x Hamrock, B.J., 1994. Fundamentals of Fluid Film Lubrication. McGraw-Hill Book.
Wear characteristics with mixed lubrication conditions in a full scale journal bearing Massimo Del Din a , Elisabet Kassfeldt b
b,)
a ABB Stal AB, Dept. TMS, SE-612 82 Finspang, ˚ Sweden DiÕision of Machine Elements, Lulea˚ UniÕersity of Technology, SE-971 87 Lulea, ˚ Sweden
Abstract Increased awareness of environmental problems has stressed the importance of switching from traditional lubricants to more environmentally friendly alternatives. Different investigations with standard test methods indicate that such a switch is possible without loss of lubricating power. The question arises if the statement above is true for industrial lubricating conditions, without laboratory cleanliness. This paper presents a study of friction and wear in a two grooved journal bearing at different shaft speeds, oil temperatures and contamination levels. A number of tests have been conducted combined with a theoretical analysis of film thickness and lubricating regime. The aim was to investigate whether an environmentally adapted rape seed-synthetic ester oil could replace a traditional mineral oil in a full scale application. The results show that the rape seed-synthetic ester oil gives significantly lower values of wear regardless of the operating conditions and there is also a tendency of lower values of frictional torque compared with the mineral oil. Mainly operating in the mixed lubrication regime, no signs of impending bearing failure have been registered, even though a large amount of silica particles added to the oil gave higher wear values than with uncontaminated oil. q 1999 Elsevier Science S.A. All rights reserved. Keywords: Wear characteristics; Journal bearing; Mixed lubrication; Environmentally adapted lubricant
1. Introduction In today’s society there is an increased need to take an environmental point of view of our day to day activities for both people in general and companies as well as for governmental institutions. Awareness on what pollution does to the environment is widely spread and this has forced actions to be taken in order to reduce the amount of environmentally harmful waste. According to Ahlbom and Duus w1x, and Bartz w2x a considerable amount of the annual lubricant consumption is likely to end up in nature instead of being properly disposed of. Answering the demands for environmentally adapted products, lubricant manufacturers have developed lubricants that do not affect people handling them or nature where the waste eventually ends up. If however the new environmentally adapted lubricants are to replace traditional oils, the lubrication performance of the former must be no worse than the latter. Several investigations of lubrication performance
) Corresponding author. Tel.: q46-920-912-40; fax: q46-920-910-47; E-mail: [email protected]
have been carried out. According to Rieglert w3x, environmentally adapted hydraulic fluids can give lower friction values than a mineral base oil and Kabuya and Bozet w4x found that vegetable oils can easily match mineral oils from a lubrication point of view. 1.1. Aim of the study The aim of the investigations performed was to compare the wear and friction performance of a fully formulated environmentally adapted oil with an equally formulated traditional mineral oil, in a full scale journal bearing application. The work was primarily focused on the slow
Table 1 Operating conditions Designation
Rotational speed wrpmx
Oil temperature w8Cx
Operating time whx
Load wkNx
1 2 3 4
100 100 10 10
70 25 70 25
48 48 48 48
35 35 35 35
0043-1648r99r$ - see front matter q 1999 Elsevier Science S.A. All rights reserved. PII: S 0 0 4 3 - 1 6 4 8 Ž 9 9 . 0 0 1 4 5 - 3
M. Del Din, E. Kassfeldtr Wear 232 (1999) 192–198 Table 2 Physical properties for test oils
193
2.1. Experimental design and oil properties
Notation
A
B
Base oil
Mineral
Viscosity q258C wmPasx Viscosity q708C wmPasx Viscosity index Density q158C wkgrm3 x
52 9 101 874
Rape seedsynth. ester 65 15 213 928
rotational speed associated with a bearing operating in the boundary lubrication regime.
The parameters altered during the investigation were oil type, oil temperature, rotational speed, and contamination level. All tests were performed both with uncontaminated oil, that is oil directly from the manufacturer, and with oil that has been contaminated with silica particles. The particle size covered a range from infinitely small to a maximum of 50 mm, with the majority between 10 and 20 mm. The duration time was 48 h for all the tests performed. The operating conditions are presented in Table 1. The physical properties of the two oils used in the tests are presented in Table 2. Both oils are fully formulated and can be bought from an ordinary oil retailer.
2. Investigation method 2.2. Experimental equipment An investigation was carried out to compare wear and friction of bearing liners in a two grooved journal bearing with different oil types when operating at various oil temperatures, shaft speeds and contamination levels. The investigation was both experimental, with several tests, and theoretical with calculations of oil film thickness and lubrication regime. The tests were carried out on a full scale test rig with the intention of being as close to real operating conditions as possible.
The bearing used in the test can normally be found in large synchronous machines. The full scale test rig consisted of four main parts; a frame, a two groove journal bearing, a large shaft and an electric motor ŽFig. 1.. The frame was constructed from welded steel beams and carried the shaft supported by two spherical roller bearings and connected to the motor via a belt transmission. The test bearing was mounted on the shaft and the
Fig. 1. Test rig.
M. Del Din, E. Kassfeldtr Wear 232 (1999) 192–198
194
During a test the temperature of the liner and the oil were measured using two thermocouples and these data were continuously recorded. The rotational speed of the shaft was measured with a tachometer and the speed controlled with velocity feedback. 2.3. Friction torque measurements In order to measure the friction torque a horizontal bar was connected to the upper side of the bearing and to a vertical bar equipped with two strain gauges ŽFig. 2.. Before each test the strain gauges were calibrated using weights and this made it possible to calculate a friction torque.
Fig. 2. Torque measuring equipment.
load applied by a hydraulic cylinder fixed to the upper side of the bearing at one end, and to the upper most beam of the frame at the other end. Oil was supplied to the bearing by a pump from a tank and the oil temperature adjusted with the assistance of an electrical hot plate. While travelling from the tank to the bearing, the oil is filtered according to the recommendations of the bearing manufacturer. Experience from previous work indicates that the contamination level during a test never falls below 0.02 wt.% in spite of the filtration. For the low temperature tests a water-cooled copper pipe was lowered into the tank to prevent temperature increase. The bearing housing divided into two parts and a liner, made out of babbit, fitted in a spherical cup covered with Teflon which was then assembled into the housing.
2.4. Wear measurements In order to measure the wear of the bearing liners a number of methods was used. Each liner was weighed before and after the tests in order to calculate the weight loss. Liners were also examined with a perthometer and optically inspected. Liners with a wear pattern differing from the normal were also inspected with a microscope and examined with SEM. 2.5. Lubricating film thickness and coefficient of wear With the test rig used in this investigation there was no possibility of measuring the oil film thickness, h, and in order to estimate the lubricating conditions in the bearing,
Table 3 Complete list of results a
Oil
T w8Cx
n wrpmx
Debr.
Visc. wmPasx
e
h min wmmx
R a wmmx
l
m loss wgx
K U 10y7
Mf wNmx
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23
A A A A B B B B B B B B A A A A B B B A A A A
q70 q25 q25 q70 q70 q25 q25 q70 q25 q70 q25 q70 q25 q70 q25 q70 q40 q40 q25 q25 q25 q70 q25
100 100 10 10 100 100 10 10 100 100 10 10 100 100 10 10 100 10 100 100 10 100 100
y y y y y y y y q q q q q q q q q q y q q q q
9 52 52 9 15 65 65 15 65 15 65 15 52 9 52 9 36 36 65 52 52 9 52
0.975 0.880 0.985 0.997 0.952 0.852 0.981 0.994 0.852 0.952 0.981 0.994 0.880 0.975 0.985 0.997 0.920 0.989 0.852 0.880 0.985 0.975 0.880
3.5 16.8 2.1 0.4 6.7 20.7 2.7 0.8 20.7 6.7 2.7 0.8 16.8 3.5 2.1 0.4 11.2 1.54 20.7 16.8 2.1 3.5 16.8
0.374 0.496 0.504 0.481 0.511 0.504 0.328 0.359 0.711 0.862 0.267 0.259 0.458 0.389 0.198 0.114 0.711 0.473 0.427 0.603 0.298 0.359 1.12
6.4 26.4 3.3 0.6 10.3 32.2 5.2 1.5 25.4 7.1 5.6 1.7 27.6 6.3 4.7 1.0 13.7 2.5 35.4 23.2 4.2 6.5 14.1
0.09 0.02 0.07 1.13 0.00 0.00 0.02 0.13 0.21 0.04 0.57 0.00 2.25 0.84 1.88 0.21 0.01 0.19 0.00 – 0.96 0.86 2.34
4.7 1.0 36.3 586.2 0 0 10.4 67.4 10.8 2.1 295.7 0 116.7 43.6 975.2 108.9 0.5 98.5 0 – 498.0 44.6 121.4
176 271 200 205 181 123 85 160 215 215 184 297 200 190 172 374 246 139 188 262 146 270 276
T s Oil temperature; h min s Minimum film thickness; m loss s Weight loss of liner; n s Shaft speed; R a s Surface roughness; K s Coefficient of wear; e s eccentricity; l s Film thickness parameter; Mf s Friction torque.
M. Del Din, E. Kassfeldtr Wear 232 (1999) 192–198
195
Fig. 3. Weight loss vs. operating condition for uncontaminated oil. Fig. 5. Friction torque vs. operating condition for uncontaminated oil.
a theoretical approach was used. With the experience of Wikstrom ¨ et al. w5x, who showed that viscous effects are negligible, an isoviscous analysis of the bearing was made and the oil film thickness calculated. From the film thickness it was possible to determine the film parameter, l, which according to Hamrock w6x is defined as
ls
h
(R
2 2 q1 q R q2
Ž 1.
l s film thickness parameter; h s film thickness wmx; R q s equivalent surface roughness wmx. Hamrock suggests that the value of l has to be 5 or more to ensure full film lubrication but for journal bear-
Fig. 4. Weight loss vs. operating condition for contaminated oil.
ings higher values in the range 8 to 10 are commonly used. Because of the poor knowledge of how the surface roughness of a liner changes during a test, the calculated values of the film parameter presented should be used with caution. As R q is often very similar to R a the latter was used in the calculations. Surface roughness varies over the worn surface although from the surface measurements one value of R a , considered to reflect the main roughness, was determined. To obtain a comparable value of wear, the coefficient of wear, K, was used instead of weight loss. K was calculated using Archard’s equation of wear Ž2. which is dimen-
Fig. 6. Friction torque vs. operating condition for contaminated oil.
196
M. Del Din, E. Kassfeldtr Wear 232 (1999) 192–198
Fig. 7. Wear for uncontaminated oil.
sionless, which is an advantage when comparing tests with different sliding distance. V FN V H sK ´Ks Ž 2. S H S FN V s Wear volume wm3 x; S s Sliding distance wmx; K s Coefficient wear; H s Materials hardness wMPax; FN s Load wNx. 3. Results The results for wear and friction torque gave the general impression, with some exceptions, that oil B gives lower values than oil A in both cases. For tests where uncommonly high weight loss values have been recorded, the tests were repeated and a mean value for weight loss and friction torque used in the evaluation. The complete results can be seen in Table 3. 3.1. Wear
Fig. 9. Liner with a v-shaped pattern.
cles, leaving any embedded particles to be weighed with the liner. The scale used had an accuracy of "0.01 g. The temperature of the oil during the tests varied "48C and the rotational speed "1 rpm. As can be seen in Fig. 3, the weight loss of the liners for test with uncontaminated oil vs. operating condition, only minimal weight loss occurred, the highest weight loss values registered for the operating condition with a rotational speed of 10 rpm and an oil temperature of q708C. When comparing both oils, oil B gave the lowest wear values. As can be seen in Fig. 4, the weight loss for the liners tested with contaminated oil vs. operating condition, the wear values were higher than with uncontaminated oil but with oil B still giving the lowest values of wear. The remarkable low values at high temperature and low rotational speed may be a result of the fact that different wear mechanisms determines the wear at different operating conditions. When examining the liners optically, only mild wear could be detected, with some polishing and a few scratches
Before weighing the bearing liners they were cleaned with solvent to remove oil and loose contamination parti-
Fig. 8. Wear for contaminated oil.
Fig. 10. SEM recordings ŽThe white line to the right of the photo represents the length, L..
M. Del Din, E. Kassfeldtr Wear 232 (1999) 192–198
197
4. Discussion and conclusions
Fig. 11. Friction torque for uncontaminated oil.
for the uncontaminated oil and a greyish wear pattern for the contaminated oil. For some tests with the contaminated oil the worn areas formed equally divided stripes in the circumferential direction, Žsee Fig. 9.. This wear pattern was also reported by V. Wikstrom. ¨ For some tests with contaminated oil a v-shaped area starting in the central region of the liner could be seen near the outlet region. This has not been mentioned in previous investigations. 3.2. Friction torque The friction torque was measured using the method described in Section 2.3 and it was continuously recorded during each test. The friction torque showed little variation over time Ž"10%. and therefore a mean value was calculated for every test run. The friction torque vs. operating condition for contaminated and uncontaminated oil are given in Figs. 5 and 6, oil B showing lower values with two exceptions. Irrespective of the level of contamination, the results are reasonable even when comparing the oils at the same operating condition. However, the level of the friction torque values are higher for contaminated oil than for uncontaminated oil and the maximum torque is reached at different operating conditions.
The two oils had different ISO viscosity grades so to be able to compare them, both wear and friction torque are presented as a function of minimum oil film thickness. Film thickness and the film thickness parameter, l, were calculated following the method presented in Section 2.5 and the complete results presented in Table 3. The wear behaviour with uncontaminated oil, Fig. 7, was as could be expected with very low K values except for the thinnest of oil films. It should be noted that the lines joining the measurement points in the diagram are an aid to differentiate between the oils, not an assumption of the behaviour between them. The wear behaviour seems reasonable given that the major wear mechanisms is through asperity contacts. The vegetable oil B, appears to give the lower wear values of the two oils. For the contaminated oil, Fig. 8, maximum wear did not occur with minimum film thickness as was the case for the uncontaminated oil. This is probably due to a change in the wear mechanism. As was the case with uncontaminated oil, oil B gives the lower K values. Uncontaminated oil gave liners with a bright wear pattern, the bright area increasing with decreasing oil film thickness. The contaminated oil resulted in a more greyish liner surface, some showing the circumferential stripes mentioned in Section 3.1. In addition to the stripes, a v-shaped pattern occurred with some of the liners tested with an oil film thickness of 2.1–3.5 mm ŽFig. 9.. Looking at this wear pattern with SEM, Žsee Fig. 10., the v-shaped area did not appear as scratched as the surrounding area which had many parallel scratch marks. A possible explanation for this is that the combination of shaft bending, due to the applied force, and the curvature of the liner along with the hydrodynamically induced oil film allowed the contamination particles to move more freely towards the outlet region. Thus, the scratches were not as deep as in the greyish area. Irrespective of contamination level, oil B seems to give the lowest friction torque value, see Figs. 11 and 12; the wear pattern being the same for both oils with a dip for h min values of approximately 2–4 mm. The friction torque
Fig. 12. Friction torque for contaminated oil.
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M. Del Din, E. Kassfeldtr Wear 232 (1999) 192–198
values follow the Stribeck curve which also has a similar dip for low h min values. This is often called the boundary lubrication regime. For both contaminated oil samples the friction has a maximum and the coefficient of wear a minimum. A possible explanation for this could be that for thin oil films the contamination particles are to large to pass between the two surfaces and therefore become embedded in the soft liner material and no wear is registered. In the region where maximum wear occurs, the oil film is of the right size to allow particles to pass without being embedded, tearing off material when rollingrsliding through the contact, and hence resulting in high K values. Thicker oil films show a low coefficient of wear, increasing slightly when higher rotational speed increase the number of abrasive particles passing through the contact region. From the study the following conclusions can be drawn, for the investigated operating conditions, oil type and materials: Ø The semi-synthetic oil led to significantly lower rates of wear than the mineral oil; Ø There is measurable tendency for the semi-synthetic oil to give lower values of friction torque than the mineral oil; Ø Neither of the oils has a higher risk of failure.
Acknowledgements The authors would like to thank their colleagues at the Division of Machine Elements for advice and valuable ¨ discussions and especially Mr. Osten Uusitalo for his skilful assistance during the work with the test rig.
References ¨ w1x J. Ahlbom, U. Duus, 1992. RENA SMORJAN? Smorjmedel ¨ mojligheter till forandring. Rapport fran ¨ ¨¨ ˚ Kemikalieinspektionen 8r92. w2x W.J. Bartz, Lubricants and the environment, Proc. of the World Tribology Congress, London, 1997. w3x J. Rieglert, Lubricating Performance of Environmentally Adapted Hydraulic Fluids, Licentiate Thesis, Lulea˚ University of Technology, Division of Machine Elements, Lulea, ˚ 1997. w4x A. Kabuya, J.L. Bozet, 1995. Comparative analysis of the lubricating power between a pure mineral oil and biodegradable oils of the same iso grade. Lubricants and Lubrication, p. 25–30. w5x V. Wikstrom, R. Larsson, Wear of bearing liners at low ¨ E. Hoglund, ¨ speed rotation of shafts with contaminated oil, Wear 162–164 Ž1993. 996–1001. w6x Hamrock, B.J., 1994. Fundamentals of Fluid Film Lubrication. McGraw-Hill Book.