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Comparing tribological effects of various chevron-based surface textures under lubricated unidirectional sliding
Tribology International 146 (2020) 106205
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Comparing tribological effects of various chevron-based surface textures under lubricated unidirectional sliding Slawomir Wos, Waldemar Koszela, Pawel Pawlus * Department of Manufacturing Technology and Production Engineering, Rzeszow University of Technology, Poland
A R T I C L E I N F O
A B S T R A C T
Keywords: Surface texturing Friction Pin-on-disc
Experiments were carried out using a pin-on-disc tribological tester in a flat-on-flat configuration in unidirec tional sliding. Both counterparts were made from steel with 50 HRC hardness. The normal load was 20 N, the sliding speed was 0.4 m/s, the number of revolutions was 10,000. Before each test, one drop of L-AN-46 oil (0.08 ml) was supplied to the inlet side of the contact zone. The discs were textured using abrasive jet machining. Tribological effects of eight chevron-based textures were compared with that of the surface having circular oil pockets. All textures had the same pit-area ratios and depths. The untextured assembly was also tested for comparison. It was found that chevron-shaped dimples were sensitive to their angular orientation to the direction of sliding. The best tribological performances were achieved for chevrons inclined to the outer sides of the rotating discs.
1. Introduction Surface texturing, which is the option of surface engineering, de pends on creating the holes (dimples, oil pockets or cavities) on the sliding surface to reduce friction under full and particularly mixed and boundary lubrications [1–5]. Due to surface texturing seizure and abrasive wear resistances can be improved. In the early works, the dimples were circular [1], because of the common opinion that the effect of the oil pockets shape on an improvement of tribological properties of sliding elements was negli gible, compared to an increase in the cost of creation of more compli cated oil pockets. However, it was found later that the geometric shapes of the dimples had substantial impact on the tribological performance of the sliding assemblies. Oil pockets of various shapes were created on machined elements typically on the basis of previous modeling. The effect of the shape of the dimples on the load-carrying capacity under gas-lubricated sliding was modeled by Qiu et al. [6]. Six shapes were taken into consideration (spherical, ellipsoidal, circular, elliptical, triangular and chevron-shaped). It was found that the ellipsoidal shape of oil pockets oriented perpendicular to the sliding direction led to the highest load-carrying capacity. Yu et al. [7,8] found that the elliptical dimple with the major axis perpendicular to the sliding direction generated the higher load-carrying capacity, compared to the ellipse positioned along
the sliding direction, circle, and triangle. These results are consistent with works [9–12] revealing that grooves positioned perpendicular to the sliding direction caused the higher reduction in the friction force that those oriented along the sliding direction. Vladescu et al. [13] created on a steel surface dimples of various shapes and similar depths. Grooves positioned perpendicularly to the sliding direction led to the best tribological performance followed by the chevron layout in condi tions of mixed and boundary lubrications under the reciprocating mo tion. Lu and Khonsari [14] found experimentally superiority of elliptical dimples over circular oil pockets in terms of reducing the friction coef ficient under mixed lubrication. Uddin et al. [15] after numerical modeling found that dimples of triangular shape gave the smallest coefficient of friction compared to oil pockets of other basis shapes. Zhang et al. [16] and Wang et al. [17] found that orientation of the triangular oil pocket to the sliding direction is important in minimizing friction. The smaller coefficient of friction corresponded to situation when one side border perpendicular to the sliding direction first entered the contact area. Zhang et al. [18] found superiority of the square dimples over triangular, circular and rectangular oil pockets in terms of optimizing tribological properties of sliding pair under lubrication in reciprocating motion. Rahmani et al. [19] optimized numerically partially textured parallel thrust bearings with squared dimples. Lu at al [20]. found that textured samples with squared dimples in line contact led to lower
* Corresponding author. E-mail address: [email protected] (P. Pawlus). https://doi.org/10.1016/j.triboint.2020.106205 Received 25 October 2019; Received in revised form 17 December 2019; Accepted 18 January 2020 Available online 23 January 2020 0301-679X/© 2020 The Authors. Published by Elsevier Ltd. This is an open (http://creativecommons.org/licenses/by-nc-nd/4.0/).
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Tribology International 146 (2020) 106205
friction coefficient under boundary and mixed lubrication regimes up to 15%. It was found in numerical and then experimental tests that textures of the bullet and fish shapes led to the lower friction force in unidirectional sliding than dimples of the circular shape [21]. Shen and Khonsari [22] numerically optimized the shapes of dimples for generating the greatest load-carrying capacity. They found that pairs of trapezoids were the best for bidirectional sliding, while chevrons for unidirectional sliding. Akbarzadeh and Khonsari [23] combined coating and surface texturing for the piston ring surface. Dimples had trape zoidal shape. This combination allowed for a reduction in both the friction force and the running-in time. Shen and Khonsari [24] tried to improve tribological properties of seal-like samples by adding dimples of different shapes. The numerical followed by experimental investigations revealed that chevron shapes with flat fronts behaved the best. Costa and Hutchings [25] analysed oil pockets of various shapes (circles, grooves and chevrons) and found experimentally that chevrons gave the greatest improvement in hydro dynamic film thickness in reciprocating motion. A numerical and then experimental studies of the effect of chevron-based textured patterns on the friction reduction were performed [26]. Dimple pattern, depth, the angle between arms (inclusion angle) and pit-area ratio varied. All tested textured patterns showed the significant friction reduction compared to the untextured surface (up to 27%). Morris et al. [27] found numerically that shallow chevrons (with a depth of 1 μm) created on cylinder liner surface at piston reversals caused the friction reduction avoiding large lubricant loss. Wos et al. [28] tested textured disc sur faces with four shapes of dimples (circle, oval, triangle and chevron) in contact with pin under lubricated unidirectional sliding. They found that dimples of chevron shape gave a decrease in the friction force compared to the untextured sample independently of the dimple array (spiral or radial). Wang et al. [29] revealed that dimples with asym metric chevron shape led to the best tribological behavior performance of the gas face seal. One can see from the literature analysis that the chevron is the promising shape of dimple in the unidirectional sliding. Under starved lubrication due to the presence of the centrifugal force in the rotary motion chevrons oriented asymmetrically to the sliding direction should be taken into consideration. It is difficult to find the results of experi mental research in this field. The aim of this work is to compare tribo logical effects of various chevron-based surface textures under lubricated unidirectional sliding. Laser texturing is the common technique of dimples creation [1,20, 23,26]. The other techniques can be also applied, like etching [14,25], honing [30] and burnishing [31,32]. Abrasive jet machining is an interesting alternative to laser texturing [28,33–36].
Fig. 1. The scheme of conformal contact: 1 – disc, 2 – small disc, 3 - pin.
volume of the one drop of oil is enough to fill dimples. Tests were carried out in starved lubrication regime to simulate conditions of a poor oil environment occurred in the start-up or stoppage of real sliding as semblies. Full lubricating mechanism can be easily simulated in contrast to mixed friction. Therefore experimental investigations under the conditions of the starved lubrication are needed. Surfaces of counterparts were polished to obtain the roughness height, determined by the Ra parameter, in the following range: 0.05–0.07 μm. Surface topographies of counterparts were measured by a white light interferometer Talysurf CCI Lite with vertical resolution of 0.01 nm. Disc samples were textured using abrasive jet machining (Fig. 2). Table 1 presents parameters of abrasive jet machining. The results of tribological behavior of sliding pairs containing discs with chevron-based textured patterns were compared with those of untextured assemblies and the sliding pair with the most popular cir cular shape. The attempt was done to obtain the same pit-area ratio of 15 � 0.5% and average depth of 6–8 μm to study the tribological effect of only shape of dimples. The designed surface area of a single dimple was 0.195 mm2. The spiral pattern of dimples was used. Dimensions of dimples and the type of pattern were selected on the basis of earlier research [28, 35, 36]. Arms of chevrons first entered the contact area. Fig. 3 presents schemes of sliding pairs. Each experiments was repeated five times. Fig. 4 presents the dimple shape and the contour plot of the disc CIR with circular oil pockets. There were eight types of chevron-based patterns. Fig. 5 presents shapes of chevrons, the direction of textured disc rotation is marked by red arrows, while the centrifugal force – green arrows. Fig. 6 presents contour maps of discs containing chevron-based dimples. Parameters of textured surfaces are presented in Table 2. Differences between the target area of a single dimple and obtained one was caused
2. Experimental details Experiments were conducted using a pin-on-disc tribological tester in a flat-on-flat scheme in unidirectional sliding. Both counterparts were made from 42CrMo4 steel with 50 HRC hardness after the heat treat ment. The special construction was used to achieve the conformal con tact between the co-acted surfaces (Fig. 1). It was used in previous research [28,35,36]. Disc 1 (sample) had diameter of 25.4 mm, while small disc 3 (counter sample) 5 mm. Tests were carried out at the ambient temperature of 21 � C, the normal load was set to 20 N. The friction radius was 8 mm, the sliding speed was set to 0.4 m/s. The test duration, determined by the number of revolutions was 10,000, the sliding distance was 502.6 m. The relative humidity was between 35 and 40%. During each test, the sliding speed, the normal load and the test duration were constant. Before each test, one drop of L-AN-46 oil (0.08 ml) was supplied to the inlet side of the contact area. The lubricant was not further supplied during the tests, therefore the experiments were carried out in starved lubrication regime. It was found in the previous work [35] that the
Fig. 2. Device for disc texturing; 1- compressed air, 2- abrasive, 3 - accelerated abrasive, 4- mask. 2
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[34] and analysis of the literature [32]. Because the experiment was conducted under unidirectional sliding in rotary motion at starved lubrication regime, the conditions of lubrication were different from two sides of the sliding direction due to the presence of the centrifugal force. Chevrons CH 600 and CH 900 were oriented symmetrically to the di rection of sliding, chevrons: CH 600 ( 200) and:CH 900 ( 200) were inclined of 200 to the inner, while chevrons: CH 600 (þ200) and CH 900 (þ200) to the outer side of the disc. The mentioned chevrons had the same lengths of arms. Two other chevrons of 900 inclusion angle were symmetrical to the direction of motion, while had different lengths of chevron arms; lower CH 900 AS1 and higher CH 900 AS2 from the inner sider of the disc. Chevron of different lengths of arms, similar to CH 900 AS2 was recommended in Reference [35] on the basis of the numerical analysis. The arm thickness was 0.2 mm, the radius of the arm end was 0.1 mm. The untextured disc sample was called UNT.
Table 1 Parameters of abrasive jet machining. Parameter
Value
Work pressure: Nozzle diameter Distance between nozzle and processed surface Abrasive Grain size
0.6 MPa 8 mm 100 mm Aluminium oxide F150 75–106 μm
by the heat affecting zone on the polypropylene mask during abrasive jet machining. This area due to overheating is sensitive to the action of abrasive particles. This was the reason why the real area of a single dimple was larger than the presumed (designed) area. Fig. 7 shows examples of extracted profiles from textured surfaces. The inclusion angles were 600: CH 600 ( 200), CH 600 (þ200) and CH 600 and 900: CH 90 0 ( 200), CH 900 (þ200), CH 900, CH 900 AS1 and CH 900 AS2. These angles were selected based on the initial research
Fig. 3. Scheme of sliding parts containing the untextured disc (a), textured discs with circular (a) and chevron-based oil pockets (c).
Fig. 4. Dimple shape (a) and contour plot of disc with circular oil pockets (CIR) (b). 3
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Fig. 5. Shapes of chevrons with their assumed dimensions.
3. Results and discussions
Addition of circular oil pockets caused both reductions in values and fluctuations of the resistance to motion (Fig. 9). After few revolution the coefficient of friction obtained the stable value a little higher than 0.02, on average for all test repetitions. The coefficient of friction marginally increased as test progressed with the low fluctuation; only in few cases abrupt increases occurred. These growths were the results of presence of the adhesive junctions caused probably by penetration of the abrasive particles to the contact zone. The reduction in the coefficient of friction was more than twice compared to the assembly with the untextured disc sample. Fig. 10 shows the friction coefficient runs in time for sliding pair
Fig. 8 shows the variation of the friction coefficient with the number of revolutions for the sliding pair with the untextured disc sample. Each curve corresponds to the test repetition. The coefficient of friction μ after initial sharp growth obtained the stable value after about 300 s. In four test runs this value was a little smaller than 0.04. The friction coefficient was stable as test progressed with a small fluctuation; the final value was about 0.04. In the fifth run, the stabilized friction coefficient was a little smaller than 0.06. This value was constant during test and increased in the final 2000 revolutions to about 0.063. 4
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Fig. 6. Contour maps of discs containing chevron-based dimples.
containing disc sample with chevron-type dimples for the inclusion angle of 600. For symmetric orientation of chevrons to the measurement direction (CH 600 sample - Fig. 10a) the coefficient of friction after initial fluctuations obtained low value (about 0.01) after 2000 revolu tions. Then, the coefficient of friction increased in time with various
rates. The final values of the coefficient of friction were in the range: 0.012–0.03. In the fifth run the coefficient of friction was similar to 0.02 after 200 revolutions, then it increased as test progressed and obtained the final value of about 0.06 – this behavior was probably caused by not proper pattern of chevrons – the low distance between dimples in the 5
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2000 revolutions and then they increased with time, obtaining the final values between 0.02 and 0.035. However, in the remaining runs, the friction coefficients were comparatively high in initial (0.04–0.057 after 2000 revolution) and final parts of tests (0.065–0.083). In contrast, the inclination of chevrons to the outer side of the disc CH 600 (þ200) sample (Fig. 10c), led to the improvement of the tribo logical behaviors of tested assembly, particularly in the initial part of the test. For all test runs the coefficients of friction obtained values about 0.01 after 10,000 resolutions, so in this case both values and variations of the friction force was small. Unfortunately, similar to other chevronbased disc surface textures, the coefficients of friction increased as tests progressed with various rates and obtained the final values in the ranges: 0.013–0.033. Fig. 11 shows the results of investigations for sliding pairs containing disc samples with chevron-shaped dimples for the inclusion angle of 900 and equal lengths of chevron arms. Various results were obtained when chevrons were located sym metrically to the direction of motion (disc CH 900 – Fig. 11a). For two test runs the coefficient of friction after obtaining shortly the stable value increased to the range: 0.023–0.035. However, in three other runs the coefficients of frictions were much higher; the final values were between 0.045 and 0.082 – higher frictional resistances were obtained compared to the analogic disc with chevrons of the smaller inclusion angle CH 600 and even the untextured disc. Inclinations of chevrons to the inner disc side; disc CH 900 ( 200) Fig. 11b, did not change significantly tribological performance of the sliding pair. In two cases the coefficients of friction were low (the final values were about 0.015) with small variation. However, for three other runs much higher resistances to motion were obtained; the final co efficients of friction were between 0.06 and 0.085. After inclinations of chevrons to the outer disc side; CH 900 (þ200) Fig. 11b, low and comparatively stable resistance to motion was ach ieved. The coefficient of friction obtained stable value after 1000 revo lutions, and then slowly increased. In four cases the final coefficients of friction were between 0.013 and 0.017; only for one, non-stable run the
Table 2 Parameters of obtained textured surfaces. Chevron shape
Total number of dimples on sliding track
Number of dimples contacted between the disc and counter sample
Average area of single dimple (based on contour map analysis), mm2
Pit-area ratio on the sliding track, %
Maximum depth of dimple, μm
CIR CH 60� ( 20� ) CH 60� (0� ) CH 60� (þ20� ) CH 90� ( 20� ) CH 90� (0� ) CH 90� (þ20� ) CH 90� AS1 CH 90� AS2
144 144
8–11 9–11
0.196 0.228
14.04 16.33
9.8 10.6
144
9–11
0.211
15.11
7.42
144
9–11
0.202
14.47
9.4
144
9–11
0.209
14.97
8.88
144
9–11
0.227
16.26
9.4
144
9–11
0.210
15.04
8.26
144
9–11
0.214
15.33
9.3
144
9–11
0.221
15.83
8.65
circumferential direction and the large in the radial direction. This could lead to the fluctuation of the hydrodynamic lift due to the irregular placement of dimples. Owing to the higher lengths of chevrons in circumferential direction the pressure generated area was larger in comparison to the circular shape of dimples. This could lead to vibra tions that disrupted the oil film generation. The inclination of chevrons to the inner side of the disc - CH 600 ( 200) sample (Fig. 10b), didn’t cause the improvement of tribological properties of the analysed sliding pair. In three runs after initial changes the coefficients of friction reached values between 0.017 and 0.02 after
Fig. 7. Extracted profiles of textured surfaces: with circular dimples CIRC (a), with chevron-based dimples CH 900 - profile of arm (b).
Fig. 8. Variation of the coefficient of friction with the number of revolutions for the sliding pair containing the untextured disc sample. 6
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Fig. 9. The coefficient of friction versus the number of revolutions for the sliding pair containing the disc sample with circular oil pockets CIR.
Fig. 10. Variation of the coefficient of friction with the number of revolutions for sliding pairs containing disc samples with chevron-shaped oil pockets CH 600 (a), CH 600 ( 200) (b) and CH 600 (þ200) (c). 7
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Fig. 11. The coefficient of friction versus the number of revolutions for sliding pairs containing disc samples with chevron-shaped oil pockets CH 900 (a), CH 900 ( 200) (b) and CH 900 (þ200) (c).
final coefficient of friction was 0.025. Unfortunately, sharps increases of the friction force occurred in some cases during tests, resulting from the presence of adhesive junctions. Fig. 12 presents the results of investigations for sliding pairs con taining disc samples with chevron-type dimples for the inclusion angle of 900 and different lengths of chevron arms. When the arm was shorter in the inner disc part (disc CH 900 AS1 Fig. 12a) the coefficient of friction increased, reached the local maximal value (typically after 100–200 revolutions) and then decreased. In two cases the coefficient of friction next increased after 2000 revolution. However, for three remaining test runs the coefficient of friction after obtaining the maximal value decreased or stabilized. This performance was different to that obtained for other sliding parts having chevronshaped dimples. Perhaps a reduction in the length of the inner arm caused the lower oil flow to the vertex point of the chevron which led to smaller oil loss. A decrease in friction resulted from matching of the
contacting surfaces. The final coefficients of friction were between 0.012 and 0.05. When the arm length was larger in the inner disc part (disc CH 900 AS2 – Fig. 12b) for all test runs the coefficient of friction after initial increasing decreased and slowly increased – this behavior was common for sliding pairs with chevron-shaped dimples. In three cases the final coefficients of friction were comparatively small (about 0.02), in two other cases they were higher (0.04). Fig. 13 presents average values and scatters of the coefficient of friction at the beginning of tests (2000–4000 revolutions), at the end of tests (8000–10000 revolutions) and at whole tests (2000–10000 revo lutions). Due to initial perturbations the first 2000 revolutions were not taken into consideration. In most cases, disc surface texturing resulted in reductions of the coefficients of friction. Circular dimples led to a reduction in the resis tance to motion about 2 times compared to the sliding pair with 8
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Fig. 12. Variation of the coefficient of friction with the number of revolutions for sliding pairs containing disc samples with chevron-type oil pockets CH 900 AS1 (a), and CH 900 AS2 (b).
untextured disc; due to texturing, the variation of the friction coefficient also much decreased. A further decrease of the coefficient of friction was achieved for sliding pairs containing discs with chevron-type dimples inclined to the outer side of discs CH 600 (þ200) and CH 900 (þ200). Discs CH 600 (þ200) behaved better at the beginning (lower values and scatters of the friction coefficients), while disc CH 900 (þ200) at the end of the tests and at whole tests. Inclinations of chevrons to the outer side of the disc did not give beneficial results; the friction forces and their scatters even increased for the lower inclusion angle CH 600 ( 200) compared to the disc with symmetrical chevrons to the direction of motion CH 600. For symmetrically oriented chevrons to the sliding di rection CH 600 and CH 900, lower values and variations of the co efficients of friction were achieved for the disc with lower inclusion angle CH 600. Various lengths of chevron arms CH 900 AS1 and CH 900 AS1 resulted in reductions in the friction coefficient values and scatters compared to the disc with chevrons of the same length of arms CH 900. A characteristic feature of sliding pairs with discs containing chevron-type dimples was growths of the coefficients of friction, at tests progressed. Due to high hardness levels of both co-acting parts, wear levels of discs were difficult to characterize. The highest peaks of disc surfaces were removed. In addition the grooves were created. In order to increase wear levels, the test duration increased 3 times. Fig. 14 presents contour
plots and extracted profiles free of dimples of selected discs after tribological tests. The height of these grooves were higher (typically between 2.5 and 5 μm) when the inclusion angles of chevrons were higher (900) compared to the smaller inclusion angle (600) – in this case heights of grooves were smaller than 2 μm. Higher wear occurred in the inner parts of the textured disc, probably due to the lack of the lubricant in the final test parts. Typically the highest frictional resistance was obtained for the untextured assembly. In this case due to the lack of dimples mixed or mainly boundary friction occurred. Circular dimples led to a reduction in the friction coefficient about two times and its stabilization compared to the sliding pair with untextured disc. The beneficial effect of adding chevron-shaped dimples instead of circular oil pockets was visible particularly in the initial test part, when chevrons led to much smaller coefficient of friction. This superiority is the consequence of the oil flow inside two chevron arms, which causes an increase in the oil pressure at the vertex leading to the higher hydrodynamic lift. However, this oil flow caused also the oil loss due to the presence of the centrifugal force leading to an increase in the friction force with time. The smaller length of the inner arm is one of method taken against the friction increase as test progresses. In contrast, much smaller oil flow inside circular dimples resulted in the stable coefficient of friction in time. Therefore the 9
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Fig. 13. Average values and scatters of the coefficient of friction at the beginning (a), at the end (b) and at whole tests (c).
beneficial effect of chevron-shaped dimples is visible especially for the larger amount of supplied oil. Tribological behavior of the sliding elements depends on orientation of chevrons to the sliding direction. The inclination of chevrons to the outer sides of discs resulted in the best tribological performance (low value and variation of the coefficient of friction). Fig. 15 will be helpful to explain this behavior. When the chevron is aligned parallel to sliding (Fig. 15a) because of similar oil flows inside two chevron arms the centrifugal force caused the slow oil loss. When chevron is inclined to the inner side of the disc (Fig. 15b) the larger amount of oil is displaced in the inner arm (closer to the disc center) to the vertex point of the chevron compared to the outer arm, leading to the higher oil loss,
compared to the action of only the centrifugal force. However, when the chevron is inclined to the outer disc side (Fig. 15c) thanks to displace ment of larger amount of the lubricant in the outer chevron arms, the oil loss due to the action of the centrifugal force is reduced and conse quently the film thickness is increased and the coefficient of friction is reduced. This analysis explains tribological performance of sliding pairs with discs having chevrons of smaller inclusion angle (600) and partially behavior of assemblies with chevron-based textures of larger inclusion angle (1200). However, the chevron lubricating mechanism can be simulated by CFD or analysed by fluid hydrodynamic lubrication. The lower values and scatters of the coefficient of friction were achieved for discs with the lower angle between chevron arms (600), 10
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Fig. 14. Contour plots (a, c, e, g) and extracted profiles free of dimples (b, d, f, h) of disc samples: UNT (a, b), CH 600 (c, d), CH 900 ( 200) (e, f) and CH 900 AS1 (g, h) after tribological tests.
thanks to better oil flow to the chevron vertex point, leading to gener ation of the higher oil pressure, compared to discs with the larger in clusion angle. Further investigation in this area would be carried out for larger amount of supplied oil to the contact zone. The inclination of chevrons to the outer side of the disc would be combined with a shorter arm in the inner side. It would be also good to study the tribological effects of
chevrons with the sloped bottoms and of different chevron-based pattern. Surface texturing should be dedicated to a special tribological sys tem. The obtained finding are valid to conformal contact in unidirec tional sliding under starved lubrication regime.
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Fig. 15. The effect of orientation of chevrons in relation to the sliding direction. Bold arrows A indicate directions of oil flow caused by disc rota tions (in addition sliding speeds are various), yel low arrows B indicate direction of small oil flow caused by the presence of the centrifugal force and red allows C indicate directions of oil flows inside the chevron: chevron aligned parallel to sliding (a), chevron inclined to the inner side of the disc (b), chevron inclined to the outer side to the disc (c). (For interpretation of the references to colour in this figure legend, the reader is referred to the Web version of this article.)
4. Conclusions
Acknowledgments
In most cases, surface texturing resulted in a reduction of frictional resistance and led to stabilization of the co-action between contacting surfaces compared to untextured samples. Tribological behaviors of chevron-shaped oil pockets are sensitive to their angular position to the direction of sliding. This orientation affects the oil flow during rotary unidirectional motion. The best tribological performances were achieved for the chevrons inclined to the outer sides of the discs. Tribological performance of the sliding pair with chevron-based disc surface texture was better for larger amount of supplied oil. During test a constant loss of lubricant due to the presence of the centrifugal force led to an increase or a fluctuation of the friction force. For chevron vertex aligned parallel to sliding the lower values and scatters of the frictional resistance were achieved for the discs with the lower angle between chevron arms. Circular dimples led to the low and stable resistance to motion.
Part of this work is was supported by the National Science Centre (Decision No. 2018/31/B/ST8/02946) “The effect of disc surface texturing on tribological properties of pin-on-disc assembly”. References [1] Etsion I. State of the art in laser surface texturing. ASME J Tribol 2005;125:248–53. [2] Nilsson B, Rosen B-G, Thomas TR, Wiklund D. Oil pockets and surface topography: mechanism of friction reduction. In: Proceedings of the XI international colloquium on surfaces, chemnitz, Germany; 2004. [3] Erdemir A. Review of engineered tribological interfaces for improved boundary lubrication. Tribol Int 2005;38:249–56. [4] Gachot C, Rosenkranz A, Hsu SM, Costa HL. A critical assessment of surface texturing for friction and wear improvement. Wear 2017;372–373:21–41. [5] Rosenkranz A, Grützmacher PG, Gachot C, Costa HL. Surface texturing in machine elements - a critical discussion for rolling and sliding contacts. Adv Eng Mater 2019:1900194. [6] Qiu M, Deli A, Raeymaekers B. The effect of texture shape on the load-carrying capacity of gas-lubricated parallel slider bearings. Tribol Lett 2012;48(3):315–32. December. [7] Yu H, Wang X, Zhou F. Geometric shape effects of surface texture on the generation of hydrodynamic pressure between conformal contacting surfaces. Tribol Lett 2010;37. 123–3. [8] Yu H, Deng H, Huang W, Wang X. The effect of dimple shapes on friction of parallel surfaces. Proc Inst Mech Eng Part J Eng Tribol 2011;225(8):693–703. [9] Pettersson U, Jacobson S. Friction and wear properties of micro textured DLC coated surfaces in boundary lubricated sliding. Tribol Lett 2004;17:553–9. [10] Moronuki N, Furukawa Y. Frictional properties of the micro-textured surface of anisotropically etched silicon. CIRP Ann - Manuf Technol 2003;52:471–4. [11] Yuan S, Huang W, Wang X. Orientation effects of micro-grooves on sliding surfaces. Tribol Int 2011;44:1047–54. [12] Zum Gahr KH, Wahl R, Wauthier K. Experimental study of the effect of microtexturing on oil lubricated ceramic/steel friction pairs. Wear 2009;267: 1241–51. [13] Vladescu SC, Olver AV, Pegg G, Reddyhoff T. The effects of surface texture in reciprocating contacts - an experimental study. Tribol Int 2015;82. 28–4. [14] Lu X, Khonsari MM. An experimental investigation of dimple effect on the stribeck curve of journal bearings. Tribol Lett 2007;27:169–76. [15] Uddin MS, Liu YW. Design and optimization of a new geometric texture shape for the enhancement of hydrodynamic lubrication performance of parallel slider surfaces. Biosurf Biotribol 2016;2:59–69. [16] Zhang H, Hua M, Dong GN, Zhang DY, Chin KS. A mixed lubrication model for studying tribological behaviors of surface texturing. Tribol Int 2016;93:583–92. [17] Wang W, Huang Z, Shen D, Kong L, Li S. The effect of triangle-shaped surface textures on the performance of the lubricated point-contacts. J Tribol 2013;135. 021503–1-11.
Author contributions Conceptualization: S.W., W.K., P.P.; methodology, investigation and formal analysis: S.W., W.K., P.P.; writing—original draft preparation: S. W., W.K., P.P.; writing—review and editing: S.W., W.K., P.P. Ethical statement Author states that the research was conducted according to ethical standards. Funding body National Science Centre. Declaration of competing interest None declared.
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Tribology International 146 (2020) 106205 [27] Morris N, Rahmani R, Rahnejat H, King PD, Howell-Smith S. A numerical model to study the role of surface textures at top dead center reversal in the piston ring to cylinder liner contact. J Tribol Apr 2016;138(2):021703. [28] Wos S, Koszela W, Pawlus P. The effect of oil pockets shape on the friction force. Tribologia 2018;4. [29] Wang X, Shi L, Dai Q, Huang W, Wang X. Multi-objective optimization on dimple shapes for gas face seals. Tribol Int 2018;123:216–23. [30] Grabon W, Pawlus P, Wos S, Koszela W, Wieczorowski M. Effects of honed cylinder liner surface texture on tribological properties of piston ring-liner assembly in short time tests. Tribol Int 2017;113:137–48. [31] Koszela W, Dzierwa A, Galda L, Pawlus P. Experimental investigation of oil pockets effect on abrasive wear resistance. TribolInt 2012;46:145–53. [32] Koszela W, Galda L, Dzierwa A, Pawlus P. The effect of surface texturing on seizure resistance of a steel-bronze assembly. Tribol Int 2010;43:1933–42. [33] Nakano M, Korenaga A, Miyake K, Murakami T, Ando Y, Usami H, Sasaki S. Applying micro-texture to cast iron surfaces to reduce the friction coefficient under lubricated conditions. Tribol Lett 2007;28:131–8. [34] Wang X, Liu W, Zhou F, Zhu D. Preliminary investigation of the effect of dimple size on friction in line contacts. Tribol Int 2009;42:1118–23. [35] Wos S, Koszela W, Pawlus P. Determination of oil demand for textured surfaces under conformal contact conditions. Tribol Int 2016;93:602–13. [36] Wos S, Koszela W, Pawlus P. The effect of both surfaces textured on improvement of tribological properties of sliding elements. Tribol Int 2017;113:137–48.
[18] Zhang YL, Zhang XG, Matsoukas G. Numerical study of surface texturing for improving tribological properties of ultra-high molecular weight polyethylene. Biosurf Biotribol 2015;1(4):270–7. [19] Rahmani R, Shirvani A, Shirvani H. Optimization of partially textured parallel thrust bearings with square-shaped micro-dimples. Tribol Trans 2007;50(3):401–6. [20] Lu P, Wood RJK, Gee MG, Wang L, Pfleging W. The friction reducing effect of square-shaped surface textures under lubricated line-contacts—an experimental study. Lubricants 2016;4(3):26. [21] Zhang H, Hua M, Dong G, Zhang D, Chen W, Dong G. Optimization of texture shape based on Genetic Algorithm under unidirectional sliding. Tribol Int 2017;115: 222–32. [22] Shen C, Khonsari MM. Numerical optimization of texture shape for parallel surfaces under unidirectional and bidirectional sliding. Tribol Int 2015;82:1–11. [23] Akbarzadeh A, Khonsari MM. Effect of untampered plasma coating and surface texturing on friction and running-in behavior of piston rings. Coatings 2018;8:110. [24] Shen C, Khonsari MM. Texture shape optimization for seal-like parallel surfaces: theory and experiment. Tribol Trans 2016;59:698–706. [25] Costa HL, Hutchings IM. Hydrodynamic lubrication of textured steel surfaces under reciprocating sliding conditions. Tribol Int 2007;40(8):1227–38. [26] Morris N, Leighton M, De la Cruz M, Rahmani R, Rahnejat H, Howell-Smith S. Combined numerical and experimental investigation of the micro-hydrodynamics of chevron-based textured patterns influencing conjunctional friction of sliding contacts. Proc IMechE Part J: J Eng Tribol 2015;229(4):316–35.
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Tribology International journal homepage: http://www.elsevier.com/locate/triboint
Comparing tribological effects of various chevron-based surface textures under lubricated unidirectional sliding Slawomir Wos, Waldemar Koszela, Pawel Pawlus * Department of Manufacturing Technology and Production Engineering, Rzeszow University of Technology, Poland
A R T I C L E I N F O
A B S T R A C T
Keywords: Surface texturing Friction Pin-on-disc
Experiments were carried out using a pin-on-disc tribological tester in a flat-on-flat configuration in unidirec tional sliding. Both counterparts were made from steel with 50 HRC hardness. The normal load was 20 N, the sliding speed was 0.4 m/s, the number of revolutions was 10,000. Before each test, one drop of L-AN-46 oil (0.08 ml) was supplied to the inlet side of the contact zone. The discs were textured using abrasive jet machining. Tribological effects of eight chevron-based textures were compared with that of the surface having circular oil pockets. All textures had the same pit-area ratios and depths. The untextured assembly was also tested for comparison. It was found that chevron-shaped dimples were sensitive to their angular orientation to the direction of sliding. The best tribological performances were achieved for chevrons inclined to the outer sides of the rotating discs.
1. Introduction Surface texturing, which is the option of surface engineering, de pends on creating the holes (dimples, oil pockets or cavities) on the sliding surface to reduce friction under full and particularly mixed and boundary lubrications [1–5]. Due to surface texturing seizure and abrasive wear resistances can be improved. In the early works, the dimples were circular [1], because of the common opinion that the effect of the oil pockets shape on an improvement of tribological properties of sliding elements was negli gible, compared to an increase in the cost of creation of more compli cated oil pockets. However, it was found later that the geometric shapes of the dimples had substantial impact on the tribological performance of the sliding assemblies. Oil pockets of various shapes were created on machined elements typically on the basis of previous modeling. The effect of the shape of the dimples on the load-carrying capacity under gas-lubricated sliding was modeled by Qiu et al. [6]. Six shapes were taken into consideration (spherical, ellipsoidal, circular, elliptical, triangular and chevron-shaped). It was found that the ellipsoidal shape of oil pockets oriented perpendicular to the sliding direction led to the highest load-carrying capacity. Yu et al. [7,8] found that the elliptical dimple with the major axis perpendicular to the sliding direction generated the higher load-carrying capacity, compared to the ellipse positioned along
the sliding direction, circle, and triangle. These results are consistent with works [9–12] revealing that grooves positioned perpendicular to the sliding direction caused the higher reduction in the friction force that those oriented along the sliding direction. Vladescu et al. [13] created on a steel surface dimples of various shapes and similar depths. Grooves positioned perpendicularly to the sliding direction led to the best tribological performance followed by the chevron layout in condi tions of mixed and boundary lubrications under the reciprocating mo tion. Lu and Khonsari [14] found experimentally superiority of elliptical dimples over circular oil pockets in terms of reducing the friction coef ficient under mixed lubrication. Uddin et al. [15] after numerical modeling found that dimples of triangular shape gave the smallest coefficient of friction compared to oil pockets of other basis shapes. Zhang et al. [16] and Wang et al. [17] found that orientation of the triangular oil pocket to the sliding direction is important in minimizing friction. The smaller coefficient of friction corresponded to situation when one side border perpendicular to the sliding direction first entered the contact area. Zhang et al. [18] found superiority of the square dimples over triangular, circular and rectangular oil pockets in terms of optimizing tribological properties of sliding pair under lubrication in reciprocating motion. Rahmani et al. [19] optimized numerically partially textured parallel thrust bearings with squared dimples. Lu at al [20]. found that textured samples with squared dimples in line contact led to lower
* Corresponding author. E-mail address: [email protected] (P. Pawlus). https://doi.org/10.1016/j.triboint.2020.106205 Received 25 October 2019; Received in revised form 17 December 2019; Accepted 18 January 2020 Available online 23 January 2020 0301-679X/© 2020 The Authors. Published by Elsevier Ltd. This is an open (http://creativecommons.org/licenses/by-nc-nd/4.0/).
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friction coefficient under boundary and mixed lubrication regimes up to 15%. It was found in numerical and then experimental tests that textures of the bullet and fish shapes led to the lower friction force in unidirectional sliding than dimples of the circular shape [21]. Shen and Khonsari [22] numerically optimized the shapes of dimples for generating the greatest load-carrying capacity. They found that pairs of trapezoids were the best for bidirectional sliding, while chevrons for unidirectional sliding. Akbarzadeh and Khonsari [23] combined coating and surface texturing for the piston ring surface. Dimples had trape zoidal shape. This combination allowed for a reduction in both the friction force and the running-in time. Shen and Khonsari [24] tried to improve tribological properties of seal-like samples by adding dimples of different shapes. The numerical followed by experimental investigations revealed that chevron shapes with flat fronts behaved the best. Costa and Hutchings [25] analysed oil pockets of various shapes (circles, grooves and chevrons) and found experimentally that chevrons gave the greatest improvement in hydro dynamic film thickness in reciprocating motion. A numerical and then experimental studies of the effect of chevron-based textured patterns on the friction reduction were performed [26]. Dimple pattern, depth, the angle between arms (inclusion angle) and pit-area ratio varied. All tested textured patterns showed the significant friction reduction compared to the untextured surface (up to 27%). Morris et al. [27] found numerically that shallow chevrons (with a depth of 1 μm) created on cylinder liner surface at piston reversals caused the friction reduction avoiding large lubricant loss. Wos et al. [28] tested textured disc sur faces with four shapes of dimples (circle, oval, triangle and chevron) in contact with pin under lubricated unidirectional sliding. They found that dimples of chevron shape gave a decrease in the friction force compared to the untextured sample independently of the dimple array (spiral or radial). Wang et al. [29] revealed that dimples with asym metric chevron shape led to the best tribological behavior performance of the gas face seal. One can see from the literature analysis that the chevron is the promising shape of dimple in the unidirectional sliding. Under starved lubrication due to the presence of the centrifugal force in the rotary motion chevrons oriented asymmetrically to the sliding direction should be taken into consideration. It is difficult to find the results of experi mental research in this field. The aim of this work is to compare tribo logical effects of various chevron-based surface textures under lubricated unidirectional sliding. Laser texturing is the common technique of dimples creation [1,20, 23,26]. The other techniques can be also applied, like etching [14,25], honing [30] and burnishing [31,32]. Abrasive jet machining is an interesting alternative to laser texturing [28,33–36].
Fig. 1. The scheme of conformal contact: 1 – disc, 2 – small disc, 3 - pin.
volume of the one drop of oil is enough to fill dimples. Tests were carried out in starved lubrication regime to simulate conditions of a poor oil environment occurred in the start-up or stoppage of real sliding as semblies. Full lubricating mechanism can be easily simulated in contrast to mixed friction. Therefore experimental investigations under the conditions of the starved lubrication are needed. Surfaces of counterparts were polished to obtain the roughness height, determined by the Ra parameter, in the following range: 0.05–0.07 μm. Surface topographies of counterparts were measured by a white light interferometer Talysurf CCI Lite with vertical resolution of 0.01 nm. Disc samples were textured using abrasive jet machining (Fig. 2). Table 1 presents parameters of abrasive jet machining. The results of tribological behavior of sliding pairs containing discs with chevron-based textured patterns were compared with those of untextured assemblies and the sliding pair with the most popular cir cular shape. The attempt was done to obtain the same pit-area ratio of 15 � 0.5% and average depth of 6–8 μm to study the tribological effect of only shape of dimples. The designed surface area of a single dimple was 0.195 mm2. The spiral pattern of dimples was used. Dimensions of dimples and the type of pattern were selected on the basis of earlier research [28, 35, 36]. Arms of chevrons first entered the contact area. Fig. 3 presents schemes of sliding pairs. Each experiments was repeated five times. Fig. 4 presents the dimple shape and the contour plot of the disc CIR with circular oil pockets. There were eight types of chevron-based patterns. Fig. 5 presents shapes of chevrons, the direction of textured disc rotation is marked by red arrows, while the centrifugal force – green arrows. Fig. 6 presents contour maps of discs containing chevron-based dimples. Parameters of textured surfaces are presented in Table 2. Differences between the target area of a single dimple and obtained one was caused
2. Experimental details Experiments were conducted using a pin-on-disc tribological tester in a flat-on-flat scheme in unidirectional sliding. Both counterparts were made from 42CrMo4 steel with 50 HRC hardness after the heat treat ment. The special construction was used to achieve the conformal con tact between the co-acted surfaces (Fig. 1). It was used in previous research [28,35,36]. Disc 1 (sample) had diameter of 25.4 mm, while small disc 3 (counter sample) 5 mm. Tests were carried out at the ambient temperature of 21 � C, the normal load was set to 20 N. The friction radius was 8 mm, the sliding speed was set to 0.4 m/s. The test duration, determined by the number of revolutions was 10,000, the sliding distance was 502.6 m. The relative humidity was between 35 and 40%. During each test, the sliding speed, the normal load and the test duration were constant. Before each test, one drop of L-AN-46 oil (0.08 ml) was supplied to the inlet side of the contact area. The lubricant was not further supplied during the tests, therefore the experiments were carried out in starved lubrication regime. It was found in the previous work [35] that the
Fig. 2. Device for disc texturing; 1- compressed air, 2- abrasive, 3 - accelerated abrasive, 4- mask. 2
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[34] and analysis of the literature [32]. Because the experiment was conducted under unidirectional sliding in rotary motion at starved lubrication regime, the conditions of lubrication were different from two sides of the sliding direction due to the presence of the centrifugal force. Chevrons CH 600 and CH 900 were oriented symmetrically to the di rection of sliding, chevrons: CH 600 ( 200) and:CH 900 ( 200) were inclined of 200 to the inner, while chevrons: CH 600 (þ200) and CH 900 (þ200) to the outer side of the disc. The mentioned chevrons had the same lengths of arms. Two other chevrons of 900 inclusion angle were symmetrical to the direction of motion, while had different lengths of chevron arms; lower CH 900 AS1 and higher CH 900 AS2 from the inner sider of the disc. Chevron of different lengths of arms, similar to CH 900 AS2 was recommended in Reference [35] on the basis of the numerical analysis. The arm thickness was 0.2 mm, the radius of the arm end was 0.1 mm. The untextured disc sample was called UNT.
Table 1 Parameters of abrasive jet machining. Parameter
Value
Work pressure: Nozzle diameter Distance between nozzle and processed surface Abrasive Grain size
0.6 MPa 8 mm 100 mm Aluminium oxide F150 75–106 μm
by the heat affecting zone on the polypropylene mask during abrasive jet machining. This area due to overheating is sensitive to the action of abrasive particles. This was the reason why the real area of a single dimple was larger than the presumed (designed) area. Fig. 7 shows examples of extracted profiles from textured surfaces. The inclusion angles were 600: CH 600 ( 200), CH 600 (þ200) and CH 600 and 900: CH 90 0 ( 200), CH 900 (þ200), CH 900, CH 900 AS1 and CH 900 AS2. These angles were selected based on the initial research
Fig. 3. Scheme of sliding parts containing the untextured disc (a), textured discs with circular (a) and chevron-based oil pockets (c).
Fig. 4. Dimple shape (a) and contour plot of disc with circular oil pockets (CIR) (b). 3
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Fig. 5. Shapes of chevrons with their assumed dimensions.
3. Results and discussions
Addition of circular oil pockets caused both reductions in values and fluctuations of the resistance to motion (Fig. 9). After few revolution the coefficient of friction obtained the stable value a little higher than 0.02, on average for all test repetitions. The coefficient of friction marginally increased as test progressed with the low fluctuation; only in few cases abrupt increases occurred. These growths were the results of presence of the adhesive junctions caused probably by penetration of the abrasive particles to the contact zone. The reduction in the coefficient of friction was more than twice compared to the assembly with the untextured disc sample. Fig. 10 shows the friction coefficient runs in time for sliding pair
Fig. 8 shows the variation of the friction coefficient with the number of revolutions for the sliding pair with the untextured disc sample. Each curve corresponds to the test repetition. The coefficient of friction μ after initial sharp growth obtained the stable value after about 300 s. In four test runs this value was a little smaller than 0.04. The friction coefficient was stable as test progressed with a small fluctuation; the final value was about 0.04. In the fifth run, the stabilized friction coefficient was a little smaller than 0.06. This value was constant during test and increased in the final 2000 revolutions to about 0.063. 4
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Fig. 6. Contour maps of discs containing chevron-based dimples.
containing disc sample with chevron-type dimples for the inclusion angle of 600. For symmetric orientation of chevrons to the measurement direction (CH 600 sample - Fig. 10a) the coefficient of friction after initial fluctuations obtained low value (about 0.01) after 2000 revolu tions. Then, the coefficient of friction increased in time with various
rates. The final values of the coefficient of friction were in the range: 0.012–0.03. In the fifth run the coefficient of friction was similar to 0.02 after 200 revolutions, then it increased as test progressed and obtained the final value of about 0.06 – this behavior was probably caused by not proper pattern of chevrons – the low distance between dimples in the 5
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2000 revolutions and then they increased with time, obtaining the final values between 0.02 and 0.035. However, in the remaining runs, the friction coefficients were comparatively high in initial (0.04–0.057 after 2000 revolution) and final parts of tests (0.065–0.083). In contrast, the inclination of chevrons to the outer side of the disc CH 600 (þ200) sample (Fig. 10c), led to the improvement of the tribo logical behaviors of tested assembly, particularly in the initial part of the test. For all test runs the coefficients of friction obtained values about 0.01 after 10,000 resolutions, so in this case both values and variations of the friction force was small. Unfortunately, similar to other chevronbased disc surface textures, the coefficients of friction increased as tests progressed with various rates and obtained the final values in the ranges: 0.013–0.033. Fig. 11 shows the results of investigations for sliding pairs containing disc samples with chevron-shaped dimples for the inclusion angle of 900 and equal lengths of chevron arms. Various results were obtained when chevrons were located sym metrically to the direction of motion (disc CH 900 – Fig. 11a). For two test runs the coefficient of friction after obtaining shortly the stable value increased to the range: 0.023–0.035. However, in three other runs the coefficients of frictions were much higher; the final values were between 0.045 and 0.082 – higher frictional resistances were obtained compared to the analogic disc with chevrons of the smaller inclusion angle CH 600 and even the untextured disc. Inclinations of chevrons to the inner disc side; disc CH 900 ( 200) Fig. 11b, did not change significantly tribological performance of the sliding pair. In two cases the coefficients of friction were low (the final values were about 0.015) with small variation. However, for three other runs much higher resistances to motion were obtained; the final co efficients of friction were between 0.06 and 0.085. After inclinations of chevrons to the outer disc side; CH 900 (þ200) Fig. 11b, low and comparatively stable resistance to motion was ach ieved. The coefficient of friction obtained stable value after 1000 revo lutions, and then slowly increased. In four cases the final coefficients of friction were between 0.013 and 0.017; only for one, non-stable run the
Table 2 Parameters of obtained textured surfaces. Chevron shape
Total number of dimples on sliding track
Number of dimples contacted between the disc and counter sample
Average area of single dimple (based on contour map analysis), mm2
Pit-area ratio on the sliding track, %
Maximum depth of dimple, μm
CIR CH 60� ( 20� ) CH 60� (0� ) CH 60� (þ20� ) CH 90� ( 20� ) CH 90� (0� ) CH 90� (þ20� ) CH 90� AS1 CH 90� AS2
144 144
8–11 9–11
0.196 0.228
14.04 16.33
9.8 10.6
144
9–11
0.211
15.11
7.42
144
9–11
0.202
14.47
9.4
144
9–11
0.209
14.97
8.88
144
9–11
0.227
16.26
9.4
144
9–11
0.210
15.04
8.26
144
9–11
0.214
15.33
9.3
144
9–11
0.221
15.83
8.65
circumferential direction and the large in the radial direction. This could lead to the fluctuation of the hydrodynamic lift due to the irregular placement of dimples. Owing to the higher lengths of chevrons in circumferential direction the pressure generated area was larger in comparison to the circular shape of dimples. This could lead to vibra tions that disrupted the oil film generation. The inclination of chevrons to the inner side of the disc - CH 600 ( 200) sample (Fig. 10b), didn’t cause the improvement of tribological properties of the analysed sliding pair. In three runs after initial changes the coefficients of friction reached values between 0.017 and 0.02 after
Fig. 7. Extracted profiles of textured surfaces: with circular dimples CIRC (a), with chevron-based dimples CH 900 - profile of arm (b).
Fig. 8. Variation of the coefficient of friction with the number of revolutions for the sliding pair containing the untextured disc sample. 6
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Fig. 9. The coefficient of friction versus the number of revolutions for the sliding pair containing the disc sample with circular oil pockets CIR.
Fig. 10. Variation of the coefficient of friction with the number of revolutions for sliding pairs containing disc samples with chevron-shaped oil pockets CH 600 (a), CH 600 ( 200) (b) and CH 600 (þ200) (c). 7
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Fig. 11. The coefficient of friction versus the number of revolutions for sliding pairs containing disc samples with chevron-shaped oil pockets CH 900 (a), CH 900 ( 200) (b) and CH 900 (þ200) (c).
final coefficient of friction was 0.025. Unfortunately, sharps increases of the friction force occurred in some cases during tests, resulting from the presence of adhesive junctions. Fig. 12 presents the results of investigations for sliding pairs con taining disc samples with chevron-type dimples for the inclusion angle of 900 and different lengths of chevron arms. When the arm was shorter in the inner disc part (disc CH 900 AS1 Fig. 12a) the coefficient of friction increased, reached the local maximal value (typically after 100–200 revolutions) and then decreased. In two cases the coefficient of friction next increased after 2000 revolution. However, for three remaining test runs the coefficient of friction after obtaining the maximal value decreased or stabilized. This performance was different to that obtained for other sliding parts having chevronshaped dimples. Perhaps a reduction in the length of the inner arm caused the lower oil flow to the vertex point of the chevron which led to smaller oil loss. A decrease in friction resulted from matching of the
contacting surfaces. The final coefficients of friction were between 0.012 and 0.05. When the arm length was larger in the inner disc part (disc CH 900 AS2 – Fig. 12b) for all test runs the coefficient of friction after initial increasing decreased and slowly increased – this behavior was common for sliding pairs with chevron-shaped dimples. In three cases the final coefficients of friction were comparatively small (about 0.02), in two other cases they were higher (0.04). Fig. 13 presents average values and scatters of the coefficient of friction at the beginning of tests (2000–4000 revolutions), at the end of tests (8000–10000 revolutions) and at whole tests (2000–10000 revo lutions). Due to initial perturbations the first 2000 revolutions were not taken into consideration. In most cases, disc surface texturing resulted in reductions of the coefficients of friction. Circular dimples led to a reduction in the resis tance to motion about 2 times compared to the sliding pair with 8
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Fig. 12. Variation of the coefficient of friction with the number of revolutions for sliding pairs containing disc samples with chevron-type oil pockets CH 900 AS1 (a), and CH 900 AS2 (b).
untextured disc; due to texturing, the variation of the friction coefficient also much decreased. A further decrease of the coefficient of friction was achieved for sliding pairs containing discs with chevron-type dimples inclined to the outer side of discs CH 600 (þ200) and CH 900 (þ200). Discs CH 600 (þ200) behaved better at the beginning (lower values and scatters of the friction coefficients), while disc CH 900 (þ200) at the end of the tests and at whole tests. Inclinations of chevrons to the outer side of the disc did not give beneficial results; the friction forces and their scatters even increased for the lower inclusion angle CH 600 ( 200) compared to the disc with symmetrical chevrons to the direction of motion CH 600. For symmetrically oriented chevrons to the sliding di rection CH 600 and CH 900, lower values and variations of the co efficients of friction were achieved for the disc with lower inclusion angle CH 600. Various lengths of chevron arms CH 900 AS1 and CH 900 AS1 resulted in reductions in the friction coefficient values and scatters compared to the disc with chevrons of the same length of arms CH 900. A characteristic feature of sliding pairs with discs containing chevron-type dimples was growths of the coefficients of friction, at tests progressed. Due to high hardness levels of both co-acting parts, wear levels of discs were difficult to characterize. The highest peaks of disc surfaces were removed. In addition the grooves were created. In order to increase wear levels, the test duration increased 3 times. Fig. 14 presents contour
plots and extracted profiles free of dimples of selected discs after tribological tests. The height of these grooves were higher (typically between 2.5 and 5 μm) when the inclusion angles of chevrons were higher (900) compared to the smaller inclusion angle (600) – in this case heights of grooves were smaller than 2 μm. Higher wear occurred in the inner parts of the textured disc, probably due to the lack of the lubricant in the final test parts. Typically the highest frictional resistance was obtained for the untextured assembly. In this case due to the lack of dimples mixed or mainly boundary friction occurred. Circular dimples led to a reduction in the friction coefficient about two times and its stabilization compared to the sliding pair with untextured disc. The beneficial effect of adding chevron-shaped dimples instead of circular oil pockets was visible particularly in the initial test part, when chevrons led to much smaller coefficient of friction. This superiority is the consequence of the oil flow inside two chevron arms, which causes an increase in the oil pressure at the vertex leading to the higher hydrodynamic lift. However, this oil flow caused also the oil loss due to the presence of the centrifugal force leading to an increase in the friction force with time. The smaller length of the inner arm is one of method taken against the friction increase as test progresses. In contrast, much smaller oil flow inside circular dimples resulted in the stable coefficient of friction in time. Therefore the 9
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Fig. 13. Average values and scatters of the coefficient of friction at the beginning (a), at the end (b) and at whole tests (c).
beneficial effect of chevron-shaped dimples is visible especially for the larger amount of supplied oil. Tribological behavior of the sliding elements depends on orientation of chevrons to the sliding direction. The inclination of chevrons to the outer sides of discs resulted in the best tribological performance (low value and variation of the coefficient of friction). Fig. 15 will be helpful to explain this behavior. When the chevron is aligned parallel to sliding (Fig. 15a) because of similar oil flows inside two chevron arms the centrifugal force caused the slow oil loss. When chevron is inclined to the inner side of the disc (Fig. 15b) the larger amount of oil is displaced in the inner arm (closer to the disc center) to the vertex point of the chevron compared to the outer arm, leading to the higher oil loss,
compared to the action of only the centrifugal force. However, when the chevron is inclined to the outer disc side (Fig. 15c) thanks to displace ment of larger amount of the lubricant in the outer chevron arms, the oil loss due to the action of the centrifugal force is reduced and conse quently the film thickness is increased and the coefficient of friction is reduced. This analysis explains tribological performance of sliding pairs with discs having chevrons of smaller inclusion angle (600) and partially behavior of assemblies with chevron-based textures of larger inclusion angle (1200). However, the chevron lubricating mechanism can be simulated by CFD or analysed by fluid hydrodynamic lubrication. The lower values and scatters of the coefficient of friction were achieved for discs with the lower angle between chevron arms (600), 10
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Fig. 14. Contour plots (a, c, e, g) and extracted profiles free of dimples (b, d, f, h) of disc samples: UNT (a, b), CH 600 (c, d), CH 900 ( 200) (e, f) and CH 900 AS1 (g, h) after tribological tests.
thanks to better oil flow to the chevron vertex point, leading to gener ation of the higher oil pressure, compared to discs with the larger in clusion angle. Further investigation in this area would be carried out for larger amount of supplied oil to the contact zone. The inclination of chevrons to the outer side of the disc would be combined with a shorter arm in the inner side. It would be also good to study the tribological effects of
chevrons with the sloped bottoms and of different chevron-based pattern. Surface texturing should be dedicated to a special tribological sys tem. The obtained finding are valid to conformal contact in unidirec tional sliding under starved lubrication regime.
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Fig. 15. The effect of orientation of chevrons in relation to the sliding direction. Bold arrows A indicate directions of oil flow caused by disc rota tions (in addition sliding speeds are various), yel low arrows B indicate direction of small oil flow caused by the presence of the centrifugal force and red allows C indicate directions of oil flows inside the chevron: chevron aligned parallel to sliding (a), chevron inclined to the inner side of the disc (b), chevron inclined to the outer side to the disc (c). (For interpretation of the references to colour in this figure legend, the reader is referred to the Web version of this article.)
4. Conclusions
Acknowledgments
In most cases, surface texturing resulted in a reduction of frictional resistance and led to stabilization of the co-action between contacting surfaces compared to untextured samples. Tribological behaviors of chevron-shaped oil pockets are sensitive to their angular position to the direction of sliding. This orientation affects the oil flow during rotary unidirectional motion. The best tribological performances were achieved for the chevrons inclined to the outer sides of the discs. Tribological performance of the sliding pair with chevron-based disc surface texture was better for larger amount of supplied oil. During test a constant loss of lubricant due to the presence of the centrifugal force led to an increase or a fluctuation of the friction force. For chevron vertex aligned parallel to sliding the lower values and scatters of the frictional resistance were achieved for the discs with the lower angle between chevron arms. Circular dimples led to the low and stable resistance to motion.
Part of this work is was supported by the National Science Centre (Decision No. 2018/31/B/ST8/02946) “The effect of disc surface texturing on tribological properties of pin-on-disc assembly”. References [1] Etsion I. State of the art in laser surface texturing. ASME J Tribol 2005;125:248–53. [2] Nilsson B, Rosen B-G, Thomas TR, Wiklund D. Oil pockets and surface topography: mechanism of friction reduction. In: Proceedings of the XI international colloquium on surfaces, chemnitz, Germany; 2004. [3] Erdemir A. Review of engineered tribological interfaces for improved boundary lubrication. Tribol Int 2005;38:249–56. [4] Gachot C, Rosenkranz A, Hsu SM, Costa HL. A critical assessment of surface texturing for friction and wear improvement. Wear 2017;372–373:21–41. [5] Rosenkranz A, Grützmacher PG, Gachot C, Costa HL. Surface texturing in machine elements - a critical discussion for rolling and sliding contacts. Adv Eng Mater 2019:1900194. [6] Qiu M, Deli A, Raeymaekers B. The effect of texture shape on the load-carrying capacity of gas-lubricated parallel slider bearings. Tribol Lett 2012;48(3):315–32. December. [7] Yu H, Wang X, Zhou F. Geometric shape effects of surface texture on the generation of hydrodynamic pressure between conformal contacting surfaces. Tribol Lett 2010;37. 123–3. [8] Yu H, Deng H, Huang W, Wang X. The effect of dimple shapes on friction of parallel surfaces. Proc Inst Mech Eng Part J Eng Tribol 2011;225(8):693–703. [9] Pettersson U, Jacobson S. Friction and wear properties of micro textured DLC coated surfaces in boundary lubricated sliding. Tribol Lett 2004;17:553–9. [10] Moronuki N, Furukawa Y. Frictional properties of the micro-textured surface of anisotropically etched silicon. CIRP Ann - Manuf Technol 2003;52:471–4. [11] Yuan S, Huang W, Wang X. Orientation effects of micro-grooves on sliding surfaces. Tribol Int 2011;44:1047–54. [12] Zum Gahr KH, Wahl R, Wauthier K. Experimental study of the effect of microtexturing on oil lubricated ceramic/steel friction pairs. Wear 2009;267: 1241–51. [13] Vladescu SC, Olver AV, Pegg G, Reddyhoff T. The effects of surface texture in reciprocating contacts - an experimental study. Tribol Int 2015;82. 28–4. [14] Lu X, Khonsari MM. An experimental investigation of dimple effect on the stribeck curve of journal bearings. Tribol Lett 2007;27:169–76. [15] Uddin MS, Liu YW. Design and optimization of a new geometric texture shape for the enhancement of hydrodynamic lubrication performance of parallel slider surfaces. Biosurf Biotribol 2016;2:59–69. [16] Zhang H, Hua M, Dong GN, Zhang DY, Chin KS. A mixed lubrication model for studying tribological behaviors of surface texturing. Tribol Int 2016;93:583–92. [17] Wang W, Huang Z, Shen D, Kong L, Li S. The effect of triangle-shaped surface textures on the performance of the lubricated point-contacts. J Tribol 2013;135. 021503–1-11.
Author contributions Conceptualization: S.W., W.K., P.P.; methodology, investigation and formal analysis: S.W., W.K., P.P.; writing—original draft preparation: S. W., W.K., P.P.; writing—review and editing: S.W., W.K., P.P. Ethical statement Author states that the research was conducted according to ethical standards. Funding body National Science Centre. Declaration of competing interest None declared.
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Tribology International 146 (2020) 106205 [27] Morris N, Rahmani R, Rahnejat H, King PD, Howell-Smith S. A numerical model to study the role of surface textures at top dead center reversal in the piston ring to cylinder liner contact. J Tribol Apr 2016;138(2):021703. [28] Wos S, Koszela W, Pawlus P. The effect of oil pockets shape on the friction force. Tribologia 2018;4. [29] Wang X, Shi L, Dai Q, Huang W, Wang X. Multi-objective optimization on dimple shapes for gas face seals. Tribol Int 2018;123:216–23. [30] Grabon W, Pawlus P, Wos S, Koszela W, Wieczorowski M. Effects of honed cylinder liner surface texture on tribological properties of piston ring-liner assembly in short time tests. Tribol Int 2017;113:137–48. [31] Koszela W, Dzierwa A, Galda L, Pawlus P. Experimental investigation of oil pockets effect on abrasive wear resistance. TribolInt 2012;46:145–53. [32] Koszela W, Galda L, Dzierwa A, Pawlus P. The effect of surface texturing on seizure resistance of a steel-bronze assembly. Tribol Int 2010;43:1933–42. [33] Nakano M, Korenaga A, Miyake K, Murakami T, Ando Y, Usami H, Sasaki S. Applying micro-texture to cast iron surfaces to reduce the friction coefficient under lubricated conditions. Tribol Lett 2007;28:131–8. [34] Wang X, Liu W, Zhou F, Zhu D. Preliminary investigation of the effect of dimple size on friction in line contacts. Tribol Int 2009;42:1118–23. [35] Wos S, Koszela W, Pawlus P. Determination of oil demand for textured surfaces under conformal contact conditions. Tribol Int 2016;93:602–13. [36] Wos S, Koszela W, Pawlus P. The effect of both surfaces textured on improvement of tribological properties of sliding elements. Tribol Int 2017;113:137–48.
[18] Zhang YL, Zhang XG, Matsoukas G. Numerical study of surface texturing for improving tribological properties of ultra-high molecular weight polyethylene. Biosurf Biotribol 2015;1(4):270–7. [19] Rahmani R, Shirvani A, Shirvani H. Optimization of partially textured parallel thrust bearings with square-shaped micro-dimples. Tribol Trans 2007;50(3):401–6. [20] Lu P, Wood RJK, Gee MG, Wang L, Pfleging W. The friction reducing effect of square-shaped surface textures under lubricated line-contacts—an experimental study. Lubricants 2016;4(3):26. [21] Zhang H, Hua M, Dong G, Zhang D, Chen W, Dong G. Optimization of texture shape based on Genetic Algorithm under unidirectional sliding. Tribol Int 2017;115: 222–32. [22] Shen C, Khonsari MM. Numerical optimization of texture shape for parallel surfaces under unidirectional and bidirectional sliding. Tribol Int 2015;82:1–11. [23] Akbarzadeh A, Khonsari MM. Effect of untampered plasma coating and surface texturing on friction and running-in behavior of piston rings. Coatings 2018;8:110. [24] Shen C, Khonsari MM. Texture shape optimization for seal-like parallel surfaces: theory and experiment. Tribol Trans 2016;59:698–706. [25] Costa HL, Hutchings IM. Hydrodynamic lubrication of textured steel surfaces under reciprocating sliding conditions. Tribol Int 2007;40(8):1227–38. [26] Morris N, Leighton M, De la Cruz M, Rahmani R, Rahnejat H, Howell-Smith S. Combined numerical and experimental investigation of the micro-hydrodynamics of chevron-based textured patterns influencing conjunctional friction of sliding contacts. Proc IMechE Part J: J Eng Tribol 2015;229(4):316–35.
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