Comparison of frictional properties of gear oils in boundary and mixed lubricated rolling–sliding and pure sliding contacts

Tribology International 62 (2013) 100–109

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Comparison of frictional properties of gear oils in boundary and mixed lubricated rolling–sliding and pure sliding contacts c ¨ Balasubramaniam Vengudusamy a,n, Alexander Grafl a, Franz Novotny-Farkas b, Werner Schofmann a

AC2T research GmbH, Viktor-Kaplan-Straße 2 D, 2700 Wiener Neustadt, Austria ¨ lhafen Lobau—Uferstrasse 8, A-1220 Vienna, Austria OMV Refining & Marketing GmbH, O c Magna Powertrain AG & Co KG, Industriestrasse 35, A-8502 Lannach, Austria b

a r t i c l e i n f o

a b s t r a c t

Article history: Received 4 October 2012 Received in revised form 20 December 2012 Accepted 1 February 2013 Available online 11 February 2013

The friction responses of five fully formulated gear oils including mineral and synthetic oils were studied. This article examines the impact of contact motion types (rolling–sliding and pure sliding) and contact pressure on boundary and mixed friction properties of the selected gear oils in MTM (minitraction machine) and SRV (Schwing-Reib-Verschleiss tribometer). Mineral oils are found to be less affected by contact pressure compared to synthetic oils. Gear oils that show adsorption appear to be less sensitive to contact motion type in mixed lubrication while behave much more sensitive in boundary lubrication regimes. The ranking of gear oils for mixed friction was similar regardless of contact motion types at low contact pressures while differ at high contact pressures. & 2013 Elsevier Ltd. All rights reserved.

Keywords: Gear oils Friction Rolling–sliding Pure sliding

1. Introduction The demand for producing friction-optimized drivetrain components is an ever increasing challenge both in the area of lubricant design and surface technology. In high pressure contacts such as cams and gears, where mixed/elastohydrodynamic (EHD) lubrication predominates, friction is mainly controlled by the rheological properties of lubricants such as its limiting shear stress and pressure–viscosity coefficient [1]. This means that the molecular structure of both base oils and additives can have great impact on the energy efficiency performance of a lubricant. Thus, optimization of both base oil and additives is important and essential, former focusing on EHD contacts while the latter on mixed/boundary lubricated contacts. The method of determining friction in mixed or boundary lubricated contacts is straightforward while it is not so in EHD contacts, especially when evaluating the friction properties of different oil types. This discrepancy is resolved by work by Gunsel et al., who defined a method to evaluate friction in EHD contacts and provided some insights into what qualities should the ‘‘ideal’’ EHD oil have [1]. Their study also highlighted the existence of correlation between EHD friction coefficient and effective pressure–viscosity, and showed that, for fluids of similar type, EHD friction coefficient increased proportionally with log (viscosity).

n

Corresponding author. Tel.: þ43 2622 816 00 330; fax: þ 43 2622 816 00 99. E-mail address: [email protected] (B. Vengudusamy).

0301-679X/$ - see front matter & 2013 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.triboint.2013.02.001

Concerning the viscosity of lubricants, it is known that less viscous lubricants exhibit lower shear losses, thus result in lower EHD friction than the more viscous ones. But the question of interest is what is the minimum viscosity required for a lubricant to provide minimum energy loss? No definite answer appears to exist since it depends on the application. But, in general, as hypothesized in [1], lubricants that are able to form sufficient lubricant film with the minimum viscosity and provide low EHD friction are considered to be a good EHD lubricant. Devlin et al. proposed a similar conclusion, too [2]. Similar study on optimizing EHD friction with a focus on other important lubricant properties such as high temperature high shear (HTHS) viscosity was conducted by Devlin et al. [2] and Bronshteyn et al. [3]. They showed that low HTHS viscosity and high pressure viscosity coefficient (a) yield low EHD friction. In contrast to Devlin et al., Greaves reported an opposite effect of a on EHD friction when studied across the temperature range of 40–120 1C [4]. Greaves compared two synthetic fluids, one with high a-value (polyalphaolefins, PAO) and the other with low a-value (polyalkylene glycols, PAG/ester blend) and showed that low a-value oils exhibit lower EHD friction than the high a-value oils [4]. Similar conclusion was also reported by Gunsel et al. [1]. Despite forming thicker EHD films, one commonly reported downside of using high a-value oils is their low viscosity indices (VI) [5,6]. The theories discussed so far are related to lubricants and their important rheological properties that control EHD friction. There is yet another complication involved in the process of optimizing the EHD friction of lubricating fluids, which is the evaluation

B. Vengudusamy et al. / Tribology International 62 (2013) 100–109

method/condition itself. It is important to pay attention to how a method in a bench test can simulate the real situation as close as possible. This is a well known question debated for many decades. Hoehn et al. have attempted to find an answer by comparing the behaviour of gear oils using different tribometers including FZG, Timken apparatus, Almen-Wieland apparatus, SAE twin-disc machine, ball-on-disc tester, 4-ball tester, pin-on-disc tester, Reichert frictional wear rig, etc. [7]. Their study highlighted the differences in these tribometers and indicated that most rigs use simple specimens under pure sliding condition, which in most cases do not directly relate to the actual gear contact. Also, Castro et al. reported a similar discrepancy in obtaining a correlation between a twin-disc machine and FZG rig [8]. These studies highlight the lack of correlation between the bench test and the actual gear contact and warn the users to be cautious when using such conventional tribometers since each was designed for certain application or to examine lubricant response to certain phenomena such as scuffing, micropitting, wear, friction, etc. It is equally important to note the difference in the types of contact geometry (circular or elliptical) and contact motion (rolling–sliding or pure sliding) that each tribometer was designed to simulate and choose the right one that possibly simulates closest to real gear contact. While most studies reported in literature focus on pure sliding contacts, few studies can be found concerning rolling–sliding contacts [9–12]. A recent study by Brandao et al. investigated the friction behaviour of gear oils in rolling–sliding contacts and showed that increasing slide–roll ratio (SRR) from 0.1 to 0.5 increases EHD friction [10]. Although several different tribometers that can simulate a wide range of contact conditions exist, FZG is still widely believed to produce results much closer to the real situation since it uses the actual gear contacts [7]. However, it should be noted that FZG studies were limited to gears operating in parallel axis and/or only for sliding bodies having surface velocities oriented in the same direction (e.g. spur gears). One question that arises is whether the conclusions made from such studies/tribometers can be applied to gears operating in non-parallel axis such as hypoid gears (i.e. sliding bodies having surface velocities oriented in different directions) or to much more generic applications. This seems unlikely since the surface velocity direction, (i.e. the angle between the surface velocity vectors designated as skew angle) has been reported to influence the EHD film thickness in a recent study by Hoehn et al. [13]. They showed lubricant film thickness to be increasing with skew angle. However, it could be possible that different oil types (mineral, synthetic, etc.) still exhibit similar EHD friction ranking for both with (surface velocities in different directions) and without (surface velocities in same direction) skew angles, former with thick while the latter with thin lubricant films. Despite all this, tribometers are still widely used and will stay in use until a better solution is found. This means that tribologists

101

will still have to depend on the existing tribometers. Thus, it is critically important to pay careful attention to apply laboratory test results to gear contacts in real applications. With this understanding, the aims of the current article were to measure friction (boundary and mixed) both under pure sliding and rolling–sliding conditions, understand their differences in friction behaviour, and thereby develop a better understanding of the correlation between laboratory experiments and the actual gear contacts. Also of interest is to understand the behaviour of different fully formulated gear oils under a wide range of contact pressures, especially when higher than 3 GPa to focus on applications like hypoid gears.

2. Materials 2.1. Ball and disc specimens The specimens used in this study were made of steel (AISI 52100) for both tribometers (see Section 3) with a mean roughness Rq of 0.015 mm for balls and 0.2 mm for discs. The roughness chosen for disc specimens correspond to the mean roughness measured on real gear flanks. 2.2. Lubricants The lubricants used in this study were commercial gear oils, representative of different types covering one semi-synthetic, two ester-blended synthetic and two mineral oils. The properties of these oils including their API GL service designation and SAE viscosity grades are listed in Table 1. Viscosity and chemical composition of these five commercial gear oils were, respectively, measured using the Stabinger Viscometer SVM 3000 (Anton Paar, Graz, Austria) according to ASTM D7042 and ICP-OES Vista-MPX (Varian, Melbourne, Australia) while pressure–viscosity coefficients were calculated based on the method described in [9]. Oils are numbered based on their viscosity at 40 1C namely, low viscosity (oils 1 and 2), medium viscosity (oil 3) and high viscosity (oils 4 and 5) oils (i.e. oil numbers 1 through 5 in Table 1 means viscosity in increasing order). This order slightly differs at 100 1C because oil 4 has the lowest VI among the five oils, thus results in greatest viscosity loss for oil 4 and change in the order of oils. Oils investigated in this study have different viscosities, thus lubricant film thicknesses of these oils differ and as a result the contacts would operate in different lubrication regimes. In order to incorporate this difference in the Stribeck curves, as proposed in [9], a modified Stribeck parameter (Sp ¼U. Z. a0.5. W  0.5) that includes pressure–viscosity coefficient a has been used; where U, Z and W, respectively, are mean speed, dynamic viscosity and normal load. The significance of this parameter is that equal values of the parameter indicate the same lubrication regime [9]. Sp and lambda ratio l (i.e. ratio of calculated elastohydrodynamic lubricant film thickness to composite surface roughness) corresponding to the

Table 1 Properties of commercial gear oils investigated. Oil no. Lubricant type

Kinematic viscosity, n (cSt) 40(1C)

1 2 3 4 5

Semi-synthetic (GL-4, SAE 75W-80) 60 Synthetic ester 2 (GL-5, SAE 75W-85) 65 Synthetic ester 3 (GL-4, SAE 75W-90) 95 Mineral 1 (GL-4, SAE 85) 110 Mineral 2 (GL-5, SAE 85W-90) 205 a

Polymethacrylates.

100 (1C)

Pressure–viscosity Viscosity Chemical composition index (VI) coefficient, a at 100 1C (GPa  1) Zn (ppm) P (ppm) S (ppm) Ca (ppm) PMAa (wt%)

10.0 12.0 15.5 11.5 18.0

9.0 9.2 10 14 15

155 188 180 92 95

– – – – –

580 3900 1000 610 2200

2500 16000 18000 16000 36000

720 10 – – –

10–15 – – – little amount

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B. Vengudusamy et al. / Tribology International 62 (2013) 100–109

mean speed of 3.5 m/s and contact pressure of 2.3 GPa at 100 1C was calculated, plotted against each other and shown in Fig. 1. These calculations were based on initial roughness and lubricants properties listed in Table 1. The typical definition of lubrication regimes based on lambda ratio and a-incorporated Stribeck parameter are: l 43, l 40.5 and l o0.5 [14] and Sp Z10  7, 10  9 rSp r10  7 and Sp r 10  9 [9], both corresponding, respectively, to thin film/EHD, mixed and boundary lubrication. However, it should be noted that in [9] relatively high viscosity gear oils were used, so the Sp bands for defining the lubrications regimes would slightly differ in this study. In this context, the initial lubrication condition in this study varied from boundary lubrication for low viscosity oils (oils 1 and 2) to mixed lubrication for high viscosity oil (oil 5) based on l, as shown in Fig. 1. However, from the shape of the Stribeck curves, which will be later shown and discussed, Sp value of 2  10  7 correspond to mixed lubrication condition and since all oils as shown in Fig. 1 have Sp values greater than 2  10  7 at 3.5 m/s mean speed, the initial lubrication condition is mostly mixed lubrication.

3. Experimental details The tribological experiments consisted of four steps namely, mild run-in, initial Stribeck curves, 2 h rubbing in boundary lubrication condition and final Stribeck curves. The allowed runin was done for 10 min at 0.05 m/s; 1 GPa; 40 1C. Stribeck curves (friction versus mean speed curves) were taken by measuring friction while varying mean speed in steps from 0.007 up to 3.5 m/s at 100 1C. Each speed step was dwelled for 30 s and friction coefficient measured during this period was averaged. The

1.0 U = 3.5 m/s

Lambda Ratio (Initial)

0.8

0.6

0.4

2 h rubbing under boundary lubrication condition was carried out at a mean speed of 0.05 m/s; 100 1C. With 0.05 m/s mean speed, the calculated lambda ratio was in the order 0.1 for the oils investigated, so the operating regime was boundary lubrication. The details are listed in Table 2. All contact pressures used in this study are maximum Hertz contact pressure. All experiments were repeated twice and showed less variation in friction (Stribeck curves looked similar), so only friction responses of one representative experiment are presented here. 3.1. Mini-traction machine (MTM) Rolling–sliding friction experiments in MTM (PCS Instruments, London, United Kingdom) were carried out using a rotating ballon-disc configuration, where a 19 mm diameter steel ball was loaded and rubbed against a steel disc immersed in lubricant. Experiments were carried out at three different SRRs namely, 0.3, 0.5 and 2 corresponding, respectively, to 30% rolling–sliding, 50% rolling–sliding and pure sliding contact conditions. The mean or entrainment speed is defined as (ub þud)/2, where ub and ud are, respectively, the speed of the ball and disc with respect to the contacting surfaces, while the slide–roll ratio SRR is defined as the ratio of sliding speed 9ub ud9 to mean speed. In pure sliding experiments, ball was held stationary (ub ¼0) while disc was allowed to rotate, so SRR equals to 2. These experiments were conducted at low contact pressures (1.0 and 1.3 GPa). However, only the initial (before any prolonged rubbing) friction results of 1 GPa are discussed in great detail. Table 2 lists the tribological test conditions used in this study. It should be mentioned that the motor speed in MTM is limited at higher SRRs, particularly at SRR¼2 (pure sliding condition), so the Stribeck curves taken at this SRR will lack the last three measurement points in the high speed region which otherwise would be present for low SRRs (0.3 and 0.5). With the maximum possible speed (2 m/s) under pure sliding condition and 1 GPa contact pressure, the calculated Sp-value for the low viscosity oil (oil 1) in this study is 2.5  10  7 at 100 1C, thus all friction comparison for mixed lubrication regime was made at this Sp-value. This friction will hereafter be called in this study as mixed friction. 3.2. Schwing-Reib-Verschleiss tribometer (SRV)

Oil 1 Oil 2 Oil 3

0.2

Oil 4 Oil 5

0.0 0.E+00

1.E-07

2.E-07

3.E-07

4.E-07

5.E-07

6.E-07

7.E-07

Modified Stribeck Parameter (initial) Fig. 1. Relation between lambda ratio l and modified Stribeck parameter Sp.

Pure sliding friction experiments were carried out using a ballon disc configuration in SRV3 rig (Optimol Instruments, Munich, Germany), where a 10 mm diameter stationary steel ball was loaded and rubbed against a rotating steel disc. The disc was held in a block mounted in a pot that contains lubricant to ensure that the contact is lubricated but not fully immersed compared to MTM. The block has a heater underneath and a control system to conduct

Table 2 Tribological test conditions employed. Step

Description

Mean speed (m/s)

Contact pressurea (GPa)

Slide-roll ratio

SRV

MTM

SRV

MTM

1

Run-in (40 1C)

0.05

1.0

0.6

Pure sliding

0.1 (10% rolling–sliding)

2

Initial Stribeck curve (100 1C)

0.007–3.5

2.3

1.0 1.3

Pure sliding

0.3 (30% rolling–sliding) 0.5 (50% rolling–sliding) 2.0 (pure sliding)

3

2 h boundary lubrication test (100 1C)

0.05

2.3

1.0

Pure sliding

0.5 (50% rolling–sliding)

4

Final stribeck curves (100 1C)

0.007–3.5

2.3 2.7 3.7

1.0 1.3

Pure sliding

0.3 (30% rolling–sliding) 0.5 (50% rolling–sliding) 2.0 (pure sliding)

a

Max. Hertzian contact pressure.

B. Vengudusamy et al. / Tribology International 62 (2013) 100–109

test at any desired temperature. SRV experiments were conducted at high contact pressures (2.3, 2.7 and 3.7 GPa) compared to MTM experiments. The details of the experimental conditions used in this study are listed in Table 2. It should be mentioned that, in contrast to MTM, lubricant pot in SRV rotates along with the disc specimen. This lead to some lubricant loss when tested at high mean speeds (3.5 m/s), but was resolved by constructive measures. At 3.5 m/s mean speed and 2.3 GPa, the calculated Sp at 100 1C for the low viscosity oil (oil 1) is 2.5  10  7, and mixed friction was compared at this Sp-value. It should be mentioned that it is mere a coincident that this Sp-value from SRV (3.5 m/s, 2.3 GPa) is similar to that chosen in MTM (2 m/s, 1.0 GPa) for the comparison of mixed friction.

4. Results 4.1. MTM friction results Fig. 2(a) and (b) shows, respectively, the initial friction for 30% rolling–sliding condition plotted against mean speed and modified Stribeck parameter Sp obtained at 1.0 GPa contact pressure. The initial friction here means that it was obtained before prolonged rubbing. It can be seen in Fig. 2(b) that Stribeck curves of low viscosity oils (oils 1 and 2) shift towards left while high viscosity oils (oils 4 and 5) shift towards right and these shifts are due to inclusion of a in Sp. The difference between these two plots is important since the frictional ranking obtained based on speed U (3.5 m/s, Fig. 2(a)) differ from those obtained based on Sp value

(2.5  10  7, Fig. 2(b)). This is because, low viscosity oils (oils 1 and 2) require higher speeds than high viscosity oils (oil 5) to achieve the same lubrication condition (Fig. 1). In contrast to the method of comparing friction based on speed, friction comparison using Sp is much more accurate since equal values of Sp ensure same lubrication condition. Therefore, Sp versus friction curves (which hereafter will be called as modified Stribeck curves) have been used for evaluating/comparing the friction performance of gear oils in this study. As mentioned in Section 3.1, in this study, Sp-value 2.5  10  7 that corresponds to mixed film lubrication condition has been used as the limiting factor for determining the mixed friction coefficient. This is the maximum value with low viscosity oil, oil 1 (Fig. 1) and also is the minimum Sp-value among five gear oils investigated, using the initial surface roughnesses of ball and disc specimens. As can be seen in Fig. 2(b), with 0.3 SRR, oils split into two groups, high friction (oils 1 and 3) and low friction (oils 2, 4 and 5) groups in the boundary lubrication (BL) regime (Sp 10  9) while no such grouping was seen in mixed lubrication regime. All oils continued to show similar grouping in BL regime even after 2 h of rubbing as shown in Fig. 3(a) and (b) but only low friction oils (oils 2, 4 and 5) showed reduction in boundary friction (cp. Fig. 2(b) and Fig. 3(b)). In contrast to initial Stribeck curves, oils showed significant variation in mixed friction as can be seen by comparing friction coefficients corresponding to the 2.5  10  7 Sp-value in Fig. 2(b) and Fig. 3(b). As normally observed in most previous studies, mixed/EHD friction was controlled by oil type and rheological properties such as pressure–viscosity coefficient and viscosity. In this study, mineral oils exhibited higher mixed friction

0.16

0.16 100°C

100°C

SRR = 0.3

1 GPa

Friction Coefficient

MTM 0.08

0.04

Oil 1

After 2h

0.12

0.12

0.00 0.001

SRR = 0.3

Initial 1 GPa

Friction Coefficient

103

Oil 2 0.010

Oil 3 0.100

Oil 4

MTM 0.08

0.04

Oil 1

Oil 5

1.000

0.00 0.001

10.000

Oil 2 0.010

Oil 3 0.100

Oil 4

Oil 5

1.000

10.000

Mean Speed (m/s)

Mean Speed (m/s)

0.16

0.16 100°C

SRR = 0.3

100°C

Initial

SRR = 0.3

After 2h

MTM 0.08

0.04

Oil 1 0.00 1.E-10

1 GPa 0.12

Friction Coefficient

Friction Coefficient

1 GPa 0.12

Oil 2

1.E-09

1.E-08

Oil 3 1.E-07

Oil 4 1.E-06

MTM 0.08

0.04

Oil 1

Oil 5 1.E-05

Modified Stribeck Parameter, Sp Fig. 2. (a) Stribeck (mean speed versus friction) and (b) modified Stribeck curves (Sp versus friction) obtained initially for 30% rolling–sliding condition at 1.0 GPa in five tested gear oils evaluated in MTM.

0.00 1.E-10

Oil 2

1.E-09

1.E-08

Oil 3

Oil 4

1.E-07

1.E-06

Oil 5 1.E-05

Modified Stribeck Parameter, Sp Fig. 3. (a) Stribeck (mean speed versus friction) and (b) modified Stribeck curves (Sp versus friction) obtained after 2 h rubbing for 30% rolling–sliding condition at 1.0 GPa in five tested gear oils evaluated in MTM.

104

B. Vengudusamy et al. / Tribology International 62 (2013) 100–109

0.16

0.16 100°C

SRR = 0.5

100°C

Initial

SRR = 2

Initial 1 GPa

0.12

Friction Coefficient

Friction Coefficient

1 GPa

MTM 0.08

0.12 MTM

0.08

0.04

0.04

Oil 1 0.00 1.E-10

Oil 2

1.E-09

1.E-08

Oil 3

Oil 4

1.E-07

Oil 1

Oil 5

1.E-06

0.00 1.E-10

1.E-05

Oil 2

1.E-09

1.E-08

Oil 3

Oil 4

1.E-07

Oil 5

1.E-06

0.16

0.16 100°C

SRR = 0.5

100°C

After 2h

SRR = 2

After 2h 1 GPa

1 GPa 0.12

Friction Coefficient

Friction Coefficient

1.E-05

Modified Stribeck Parameter, Sp

Modified Stribeck Parameter, Sp

MTM 0.08

MTM

0.08

0.04

0.04 Oil 1 0.00 1.E-10

0.12

Oil 2

1.E-09

1.E-08

Oil 3 1.E-07

Oil 4 1.E-06

Oil 1

Oil 5 1.E-05

0.00 1.E-10

than the others and at the Sp-value 2.5  10  7, oil 1 showed lower mixed friction followed, in order, by oils 3, 2, 4 and 5. The initial and final Stribeck curves of all oils examined at 50% rolling–sliding and pure sliding conditions are, respectively, shown in Fig. 4(a) and (b) and Fig. 5(a) and (b). Regardless of SRR, the above described friction responses, both grouping of oils in BL regime and varied mixed friction appear to be an inherent property of the oils studied [cp. Figs. 3–5]. One exception is pure sliding condition (SRR¼2), in which Stribeck curves of all low boundary friction oils (oils 2, 4 and 5) approached high boundary friction oils (oils 1 and 3) but grouping was still seen. Also, all oils under pure sliding condition showed a considerable increase in mixed friction compared to 30% and 50% rolling–sliding conditions. The Stribeck curves of 30% rolling–sliding, 50% rolling– sliding and pure sliding conditions obtained at 1.3 GPa are shown, respectively, in Fig. 6(a)–(c). With 1.3 GPa, all oils responded similar to that of 1.0 GPa except oil 4, the boundary friction coefficients of which approached the boundary friction coefficients of oil 1 regardless of SRR, as shown in Fig. 6(a)–(c). Also, regardless of SRR, all oils except oil 1 showed a marginal decrease in mixed friction with 1.3 GPa compared to 1.0 GPa [cp. Fig. 3(b), Fig. 4(b), Fig. 5(b) and Fig. 6(a)–(c)]. 4.2. SRV friction results Fig. 7(a) and (b) shows the modified Stribeck curves of five gear oils obtained initially and after 2 h rubbing at 2.3 GPa in pure sliding condition. The difference in friction (both boundary and mixed) seen initially between oils was not seen after 2 h of rubbing, i.e. Stribeck curves of all oils merged but showed a

1.E-08

Oil 3

Oil 4

1.E-07

1.E-06

Oil 5

1.E-05

Modified Stribeck Parameter, Sp

Modified Stribeck Parameter, Sp Fig. 4. Modified Stribeck curves obtained (a) initially and (b) after 2 h rubbing for 50% rolling–sliding condition at 1.0 GPa in five tested gear oils evaluated in MTM.

1.E-09

Oil 2

Fig. 5. Modified Stribeck curves obtained (a) initially and (b) after 2 h rubbing for pure sliding condition (SRR ¼2) at 1.0 GPa in five tested gear oils evaluated in MTM.

marginal decrease in both boundary and mixed friction after 2 h of rubbing. The modified Stribeck curves after 2 h of rubbing at 2.7 and 3.7 GPa are, respectively, shown in Figs. 8 and 9. With increase in contact pressure, all oils showed a marginal increase in boundary friction while a marginal decrease in mixed friction [cp. Figs. 7(b), 8 and 9]. In all cases, oil 1 showed lower mixed friction than the others.

5. Discussion In this study, five fully formulated gears oils were investigated for their boundary and mixed friction properties and the results reveal differences in friction behaviour. In most cases, as shown in Fig. 10, it is evident that mixed/EHD friction was primarily controlled by oil type, viscosity/pressure–viscosity coefficient and secondarily by contact conditions defined by contact motion (i.e. rolling–sliding or pure sliding) and contact pressure. 5.1. Effect of oil type, viscosity and pressure–viscosity coefficient on mixed friction In order to highlight the observed differences in friction behaviour, only selective results from MTM and SRV were considered for comparison. The friction values corresponding to 2.5  10  7 Sp-value were compared, as shown in Fig. 10. With oil types, as can be seen from Fig. 10, synthetic oils (oils 1, 2 and 3) showed lower mixed friction than the mineral oils (oils 4 and 5) under rolling–sliding conditions (i.e. SRR of 0.3 and 0.5). Oil 1,

B. Vengudusamy et al. / Tribology International 62 (2013) 100–109

105

0.16

0.16 100°C

SRR = 0.3

100°C

After 2h

2.3 GPa

Initial SRV

0.12

Friction Coefficient

Friction Coefficient

1.3 GPa 0.12 MTM 0.08

0.04 Oil 1 0.00 1.E-10

Oil 2

1.E-09

1.E-08

Oil 3

Oil 4

1.E-07

0.08

0.04

Oil 5 1.E-06

Oil 1 0.00 1.E-10

1.E-05

Oil 2

1.E-09

Modified Stribeck Parameter, Sp

1.E-08

Oil 3

Oil 4

1.E-07

Oil 5 1.E-06

1.E-05

Modified Stribeck Parameter, Sp

0.16

0.16 100°C

SRR = 0.5

100°C

After 2h

2.3 GPa

After 2h SRV

0.12

Friction Coefficient

Friction Coefficient

1.3 GPa 0.12 MTM

0.08

0.04 Oil 1 0.00 1.E-10

Oil 2

1.E-09

1.E-08

Oil 3

Oil 4

1.E-07

0.08

0.04

Oil 5 1.E-06

Oil 1 0.00 1.E-10

1.E-05

Oil 2

1.E-09

Modified Stribeck Parameter, Sp

1.E-08

Oil 3

Oil 4

1.E-07

Oil 5 1.E-06

1.E-05

Modified Stribeck Parameter, Sp Fig. 7. Modified Stribeck curves obtained (a) initially and (b) after 2 h rubbing for pure sliding condition at 2.3 GPa in five tested gear oils evaluated in SRV.

0.16 100°C

SRR = 2

After 2h 0.16 100°C

2.7 GPa

After 2h

MTM

SRV 0.08

Friction Coefficient

Friction Coefficient

1.3 GPa 0.12

0.04 Oil 1 0.00 1.E-10

Oil 2

1.E-09

1.E-08

Oil 3

Oil 4

1.E-07

Oil 5 1.E-06

0.12

0.08

0.04

1.E-05

Modified Stribeck Parameter, Sp Oil 1

Fig. 6. Modified Stribeck curves obtained after 2 h rubbing at 1.3 GPa for (a) 30% rolling–sliding, (b) 50% rolling–sliding and (c) pure sliding condition (SRR¼ 2) in five tested gear oils evaluated in MTM.

being the only one blended with a polymethacrylate (PMA) type viscosity modifier exhibited lower mixed friction than oils that contained no PMAs (oils 2 and 3), but ester based-fluids. Oils blended with high PMA concentrations (here 10–15 wt%) are known to provide low mixed/EHD friction [15]. However, at high contact pressures under pure sliding contact conditions, the benefit of PMAs was lost (shown by increase in mixed friction), probably because limiting shear stress was reached. However, further increase in contact pressure to 3.7 GPa decreased mixed friction as shown in Fig. 10, probably because limiting shear stress decreased due to thermal effects. Among mineral oils (oils 4 and 5), low viscosity mineral oil (oil 4) showed lower mixed friction than the high viscosity mineral oil (oil 5)

0.00 1.E-10

1.E-09

Oil 2 1.E-08

Oil 3

Oil 4

1.E-07

Oil 5 1.E-06

1.E-05

Modified Stribeck Parameter, Sp Fig. 8. Modified Stribeck curves obtained after 2 h rubbing for pure sliding condition at 2.7 GPa in five tested gear oils evaluated in SRV.

under rolling–sliding condition while both oils showed similar friction values under pure sliding condition. Oil 5, containing small amounts of PMA [in function of pour point depressants (PPD)], did not show any reduction in mixed friction in most cases, probably the concentration here appeared to be too low for significant action in mixed/ EHD region. Also, oils 1 to 5, respectively correspond to low through to high viscosity at 40 1C but this order differs slightly at 100 1C for oils 3 and 4. However, mixed friction of most oils investigated in this study was controlled by viscosity at 40 1C despite being examined at 100 1C. This is evident from Fig. 10, which shows mixed friction to increase with viscosity for all oils except oil 5, which showed a slight

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0.16 100°C

3.7 GPa

After 2h

Friction Coefficient

SRV 0.12

0.08

0.04 Oil 1 0.00 1.E-10

Oil 2

1.E-09

Oil 3

1.E-08

Oil 4

1.E-07

Oil 5 1.E-06

1.E-05

Modified Stribeck Parameter, Sp Fig. 9. Modified Stribeck curves obtained after 2 h rubbing for pure sliding condition at 3.7 GPa in five tested gear oils evaluated in SRV.

0.12 100°C

Mixed Friction

0.10

0.08

0.06

SRR = 0.3, 1 GPa, MTM SRR = 0.5, 1 GPa, MTM Pure sliding, 1 GPa, MTM Pure sliding, 2.3 GPa, SRV Pure sliding, 2.7 GPa, SRV Pure sliding, 3.7 GPa, SRV

0.04

0.02

that no polymer-type friction behaviour was observed with oil 1 at low speeds. This indicates that no thick PMA boundary films were formed with this oil after 2 h. For PMA polymers to form thick boundary films, it is necessary for them to have functional groups [15]. Here it is suggested that PMA in oil 1 may not have such functional groups, thus not promoting significant adsorption to the surface and not effectively reducing boundary friction. In contrast to this, oil 5 that also had PMA but at low concentration seems to promote adsorption (like expected by PMA boundary film formation) both at low and high SRRs but only up to 2.7 GPa (Fig. 11(e)). But, when looking at the interesting shape of the Stribeck curves for this oil (oil 5), which are similar to oils 2, 4 in the Sp range 10  9–10  8, where boundary friction decreased with decrease in mean speed or Sp, a more convincing explanation can be conceived for this behaviour. This type of low-speed friction behaviour has been reported to be the typical response of friction modifiers (FM) [16,17]. Here it is suggested that a FM film formed rapidly even before any prolonged rubbing (shown by similar low friction grouping for oils 2, 4 and 5 in Fig. 2), continued to retain on surface even after 2 h of rubbing (Fig. 11) and exhibited such a low-speed boundary friction characteristics. This results from the physical adsorption of FM molecules on to the sliding surfaces [16–18]. In this study, it appears that these adsorbed FM molecules from oils 2, 4 and 5 slightly lose their original low boundary friction properties when SRR was increased, in particular at high contact pressures and pure sliding condition. However, the boundary friction values of these oils under pure sliding condition (SRR ¼2) are still lower than that of oils 1 and 3. In contrast to oils 2, 4 and 5, additives in oils 1 and 3 were neither effective nor showed any adsorption so their boundary friction is slightly less sensitive to SRR.

0.00 Oil 1

Oil 2

Oil 3

Oil 4

Oil 5

Fig. 10. Comparison of mixed friction at different SRRs and contact pressures (lines are drawn to guide the eyes).

decrease in friction at higher contact pressures and pure sliding condition. Similar behaviour was observed with pressure–viscosity coefficient where low a-value oils tended to exhibit lower mixed/EHD friction than high a-value oils, conforming to the results reported in [1]. In general, low viscosity oil (oil 1) showed lower mixed friction than the medium (oils 2 to 4) and high (oils 4 and 5) viscosity oils at 100 1C, especially under rolling–sliding condition while under pure sliding and high contact pressure (42 GPa) condition mixed friction was similar. 5.2. Influence of SRR or contact motion type (rolling–sliding or pure sliding) on friction It is interesting to know how different oil types having varying viscosity and a-values respond to different contact conditions. The comparison of Stribeck curves obtained at different SRR and contact pressure is shown in Fig. 11. The results shown in Fig. 11 include both MTM (30% and 50% rolling–sliding and pure sliding at 1.0 GPa) and SRV (pure sliding at 2.3, 2.7 and 3.7 GPa) results. 5.2.1. Boundary friction By comparing MTM results in boundary lubrication regime in Fig. 11, it can be observed that oils 1 and 3 responded quite similarly and showed negligible change in boundary friction with SRR while oils 2, 4 and 5 responded quite similarly but showed an increase in boundary friction with SRR. The increase was marginal between 30% and 50% rolling–sliding conditions while significant when SRR was increased to 2. It should be noted

5.2.2. Mixed friction It can be seen from Fig. 10 that all oils showed increase in mixed friction with increase in SRR. This friction trend is in agreement with results reported by Brandao et al. [9]. However, only oils 1 and 3 showed considerable increase in mixed friction, former when SRR ¼2; 2.3 GPa and the latter when SRR¼2; 1.0 GPa. Mixed friction increases with SRR as probably the shear stress increases. Martini et al. reported similar effect of SRR on friction [19] but their study was limited to 1.0 GPa. In this study, at low contact pressures, oil 1 (ca. 0.045) showed lower mixed friction than the others (ca. 0.07–0.09). At high contact pressure (3.7 GPa), only oils 1 and 5 were able to provide low mixed friction (ca. 0.07) in pure sliding condition, probably because of PMA present in these two oils. However, this value was much higher than mixed friction value of oil 1 observed at low contact pressures. It appears that SRR has less effect on mixed friction for oils 1, 2, 4 and 5 at 1.0 GPa while notable effect for oils 1 and 3, in particular at high contact pressures (e.g. oil 1 at 2.3 GPa, Fig. 10). It is concluded that oils 2, 4 and 5 showed additive (i.e. FM) adsorption while oils 1 and 3 showed none due to lack of FM. The results indicate that when adsorption is present mixed friction is not controlled by contact motion type, especially at low contact pressures. Furthermore, it is suggested that mixed friction and contact motion type are to some extent decoupled when additive adsorption occurs, in particular at low contact pressures. The question of interest is how adsorption makes mixed friction insensitive to contact motion type? One commonly reported phenomenon concerning the effect of FM adsorption on EHD friction is liquid slip at solid–liquid interface. In this context, now, the question of interest is can adsorption cause slip, thereby cause friction to be insensitive to contact motion type?

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0.16

0.16 After 2h 100°C

0.12

0.08 SRR = 0.3, 1 GPa, MTM SRR = 0.5, 1 GPa, MTM Pure sliding, 1 GPa, MTM Pure sliding, 2.3 GPa, SRV Pure sliding, 2.7 GPa, SRV Pure sliding, 3.7 GPa, SRV

0.04

0.00 1.E-10

1.E-09

1.E-08

1.E-07

1.E-06

Oil 4

Friction Coefficient

Friction Coefficient

Oil 1

100°C

0.08 SRR = 0.3, 1 GPa, MTM SRR = 0.5, 1 GPa, MTM Pure sliding, 1 GPa, MTM Pure sliding, 2.3 GPa, SRV Pure sliding, 2.7 GPa, SRV Pure sliding, 3.7 GPa, SRV

0.04

0.00 1.E-10

1.E-05

After 2h

0.12

Modified Stribeck Parameter, Sp

1.E-09

1.E-08

1.E-07

1.E-06

1.E-05

Modified Stribeck Parameter, Sp

0.16

0.16 100°C

0.12

0.08 SRR = 0.3, 1 GPa, MTM SRR = 0.5, 1 GPa, MTM Pure sliding, 1 GPa, MTM Pure sliding, 2.3 GPa, SRV Pure sliding, 2.7 GPa, SRV Pure sliding, 3.7 GPa, SRV

0.04

0.00 1.E-10

1.E-09

1.E-08

1.E-07

1.E-06

1.E-05

Modified Stribeck Parameter, Sp

After 2h

Oil 5

After 2h

Friction Coefficient

Oil 2

Friction Coefficient

107

100°C

0.12

0.08

0.04

0.00 1.E-10

SRR = 0.3, 1 GPa, MTM SRR = 0.5, 1 GPa, MTM Pure sliding, 1 GPa, MTM Pure sliding, 2.3 GPa, SRV Pure sliding, 2.7 GPa, SRV Pure sliding, 3.7 GPa, SRV

1.E-09

1.E-08

1.E-07

1.E-06

1.E-05

Modified Stribeck Parameter, Sp

0.16

Friction Coefficient

Oil 3

After 2h 100°C

0.12

0.08

0.04

0.00 1.E-10

SRR = 0.3, 1 GPa, MTM SRR = 0.5, 1 GPa, MTM Pure sliding, 1 GPa, MTM Pure sliding, 2.3 GPa, SRV Pure sliding, 2.7 GPa, SRV Pure sliding, 3.7 GPa, SRV

1.E-09

1.E-08

1.E-07

1.E-06

1.E-05

Modified Stribeck Parameter, Sp Fig. 11. Comparison of modified Stribeck curves at different SRRs for (a) oil 1, (b) oil 2, (c) oil 3, (d) oil 4 and (e) oil 5.

This seems improbable because to enable slip, it is necessary to have smooth surfaces, Rq less than 5 nm [20,21] and/or one of the surfaces to be non-wetting [22]. Neither prerequisite was achieved in the current study. Also, no reduction in mixed friction was noted for the oils that showed additive adsorption (oils 2, 4 and 5), which otherwise should have reduced if slip had occurred. This indicated that adsorption did not cause any slip. It should be mentioned that this study did not conduct any explicit study on slip instead only discusses some possible causes for the unusual behaviour observed with oils 2, 4 and 5 based on friction measurements. So, the cause for insensitiveness of friction to contact motion type observed with these three oils is not clear yet and needs further investigation. In general, as shown in Fig. 12(a) and (b), when rolling was present, mixed friction showed a linear relationship with viscosity/ pressure–viscosity coefficient (Fig. 12(a)) while this relationship was not pronounced when rolling was not present (i.e. in pure sliding) or when contact pressure was higher (Fig. 12(b)). However, the ranking of oils was similar regardless of whether the contact was rolling– sliding or pure sliding but only at low contact pressures (1.0 GPa). At high contact pressures and under pure sliding condition, all oils showed similar mixed friction [see friction results of contact pressure 42 GPa in Fig. 12(b)].

Overall, additive adsorption influences both boundary and mixed friction coefficients, former with significant decrease while the latter with negligible change. Additive adsorption controls the boundary friction but was found to be degraded when SRR was increased to 2 (pure sliding condition). Additives in oils 1 and 3 appeared not to be as effective as those in other oils regardless of SRR because of not having functional groups in PMA or FM, thus always exhibited high boundary friction. Also, oils 2, 4 and 5 that showed adsorption appear to be less sensitive to contact motion type in mixed lubrication (Fig. 10) while much sensitive in boundary lubrication (BL) regimes. Thus significant difference in friction was seen in BL while not in mixed lubrication when SRR was increased. Oil 1 (having low viscosity and high concentration of PMA) always exhibited lower mixed friction than the others regardless of SRR and contact pressures investigated in this study. 5.3. Effect of contact pressure on friction 5.3.1. Comparison of rolling–sliding contacts in MTM By comparing Fig. 3(b) and Fig. 6(a) for 30% rolling–sliding and Fig. 4b and Fig. 6b for 50% rolling–sliding, it can be observed that most oils investigated in this study showed increase in boundary friction with increase in contact pressure. The increase was

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0.12 100°C

Mixed Friction

0.10

0.08

0.06 SRR = 0.3, 1 GPa, MTM SRR = 0.3, 1.3 GPa, MTM

0.04

SRR = 0.5, 1 GPa, MTM SRR = 0.5, 1.3 GPa, MTM 0.02 Oil 1

Oil 2

Oil 3

Oil 4

Oil 5

0.12 100°C

Mixed Friction

0.10

0.08

0.06

Pure sliding, 1 GPa, MTM Pure sliding, 1.3 GPa, MTM Pure sliding, 2.3 GPa, SRV

0.04

Pure sliding, 2.7 GPa, SRV Pure sliding, 3.7 GPa, SRV 0.02 Oil 1

Oil 2

Oil 3

Oil 4

Oil 5

Fig. 12. Comparison of mixed friction at different contact pressures for (a) rolling– sliding and (b) pure sliding contacts (lines are drawn to guide the eyes).

significant with oils 2, 4 and 5 while negligible with oils 1 and 3. However, the boundary friction values of oils 2, 4 and 5 were much lower than that of oils 1 and 3. One interesting observation with oil 4 at 1.3 GPa as shown in Fig. 6(a)–(c) is the friction response at intermediate and very low speed or Sp regime. Oil 4 markedly reduced the friction in the Sp range 5  10  9–5  10  8 while significantly increased the boundary friction coefficient at very low Sp values. This may be because oil 4 formed an adsorbed film that survived and reduced friction when majority of load was shared by fluid film in the intermediate Sp range (or speed range). But, at very low Sp (very low speed) range, where there was no fluid film, the load has to be solely supported by adsorbed films and it appears that the pressure (1.3 GPa) exceeded the level that the adsorbed films can withstand, so the adsorbed films might have collapsed, thereby causing increase in boundary friction. Similar adsorption behaviour at intermediate speeds was reported by Topolovec-Miklozic et al. [16] but for DLC surfaces. In terms of mixed friction, as shown in Fig. 12(a), all oils showed decrease in mixed friction with increase in contact pressure except oil 1. A similar study by Brandao et al. reported a similar effect of contact pressure on mixed friction [9]. However, it is important to note the difference in operating conditions employed in this and Brandao’s studies. This study employed SRR up to 2 while their study up to 0.5 and a 2-h rubbing in boundary lubrication condition was included in this study while not in the work by Brandao et al. [9]. Despite the differences in operating conditions and additive packages used in both the studies, contact pressure appears to show similar effects on mixed friction, i.e. contact pressure decreases the mixed friction.

5.3.2. Comparison of pure sliding contacts in MTM and SRV By comparing Fig. 5(b) and Fig. 6(c), it can be observed that all oils tested in pure sliding condition in MTM showed decrease in boundary friction with increase in contact pressure, although only marginal. However, when tested in SRV all oils showed increase in boundary friction with contact pressure as shown by SRV results in Fig. 11. With oils 1 and 3, as can be seen from Fig. 11, boundary friction obtained in SRV at 2.3 or 2.7 GPa was much lower than that obtained in MTM and boundary friction values of 3.7 GPa in SRV approached the boundary friction values of 1.0 or 1.3 GPa in MTM. While other oils showed no such response, one interesting observation with oil 2 is its similar boundary friction behaviour at all contact pressures investigated, under pure sliding condition, both in MTM and SRV [see Fig. 11(b)]. This indicates that additives present in this oil appeared to be much stable and effective at all contact pressures investigated and in general less sensitive to contact pressure compared to SRR. The results show that while increase in contact pressure caused decrease in boundary friction in MTM, it caused increase in boundary friction in SRV. It appears that at low contact pressures ( o2 GPa; MTM) boundary friction decreases with contact pressure while at high contact pressures (42 GPa; SRV) boundary friction increases with contact pressure. In terms of mixed friction as shown in Fig. 12(b), most oils investigated in this study, generally showed decrease in mixed friction with increase in contact pressure under pure sliding condition, both in MTM and SRV. One exception is oil 1, which showed slight increase in mixed friction with contact pressure in MTM while reverse effect in SRV. In general, the friction response to contact pressure was different with respect to lubrication regime, both in rolling–sliding and pure sliding contacts. While increase in contact pressure caused increase in friction in boundary lubrication regime, it caused opposite effect in mixed lubrication regime. This indicates that boundary and lubricant films responded quite differently to contact pressure. One common response between rolling–sliding and pure sliding contacts is that, in general, both showed decrease in mixed friction with an increase in contact pressure, except for oil 5. One possibility for this behaviour could be due to increased liquid slip. This phenomenon was reported by Guo et al. [23], who experimentally showed that boundary slip increases with an increase in contact pressure. However, their study was limited to oil/glass interface in contrast to oil/steel interface in this study, which needs further investigation. Despite the difference in contact pressure, one noteworthy difference between MTM and SRV friction results is the shape of Stribeck curves in pure sliding condition. Although both did not show a classical shape in pure sliding condition as otherwise normally seen under rolling–sliding conditions (Fig. 11), Stribeck curves obtained from SRV appeared much more flatter than those from MTM [cp. Figs. 6(c) and 7(b)]. This probably results from the difference in contact pressure and lubrication condition employed in MTM and SRV. Friction tests in MTM were conducted at much lower contact pressure compared to SRV, and the ball and disc contacts in MTM were fully immersed while not so in SRV. These differences along with the change in surface condition during the experiment could be attributed for the observed difference in the shape of Stribeck curve and also the difference in boundary friction behaviour. This indicates that additives were activated quite differently in these two experiments and to some extent based on contact pressure. At higher contact pressures such as 3.7 GPa, all oils showed similar boundary friction values of about ca. 0.12 except oil 2 while showed varied boundary friction behaviour at low contact pressures (1.0–2.7 GPa). This suggests that additives become ineffective beyond certain contact pressures and here in this study at 3.7 GPa. Thus conducting experiments at low contact pressures appear to be much reliable for

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ranking oils in terms of their additive behaviour, based on the experimental conditions and oils used in this study. Another interesting difference in the experimental conditions of MTM and SRV is the ball size. As mentioned earlier, 19 mm and 10 mm diameter balls were, respectively, used in MTM and SRV. One might question whether size of the ball could influence the lubrication regime or the limits of Sp or the shape of Stribeck curve itself. It appears to have only negligible influence as, for example, the Stribeck curves of oil 2 shown in Fig. 11(b) taken under pure sliding condition at 1.0 GPa in MTM (19 mm diameter ball) and pure sliding condition at 2.3 GPa in SRV (10 mm diameter ball) are quite similar. Also, these two Stribeck curves (i.e. 1.0 GPa in MTM and 2.3 GPa, both under pure sliding condition) are similar for oils that showed additive adsorption (oils 2, 4 and 5) while dissimilar for oils that showed no adsorption (oils 1 and 3). These results suggest that ball size only has negligible influence on lubrication regimes and the observed differences in the shape of Stribeck curves were mainly due to action of lubricants/additives.

6. Conclusions Boundary and mixed friction properties of five fully formulated gear oils including three synthetic oils and two mineral oils at different SRRs and contact pressures have been investigated. The results showed that mixed friction is primarily controlled by oil type, viscosity and pressure–viscosity coefficient. It is not surprising to see synthetic oils or oils having low viscosity/ a-value exhibiting lower mixed/EHD friction than mineral oils or oils having high viscosity/a-value. But, it is important to understand the friction behaviour of these different oil types, having different additive packages and viscosity/pressure-viscosity coefficient, under different contact motion types (rolling–sliding or pure sliding) and contact pressures. The results can be summarised as follows.

 Some oils (oils 1 and 3) are sensitive to contact motion type





while some are not (oils 2, 4 and 5) and this phenomenon, to some extent is controlled by adsorption properties but respond disparately to lubrication regimes, i.e. oils that showed adsorption tended to behave less sensitive to contact motion type in mixed lubrication while much sensitive in boundary lubrication compared to oils that showed no adsorption. It is clearly important to carry out adsorption studies when formulating friction modifier-containing lubricants and not to focus solely on friction measurements at high speeds. The ranking of gear oils in terms of their mixed friction appeared to be similar regardless of whether the contact is rolling–sliding or pure sliding at low contact pressures but slightly differed when the contact pressure was increased, especially for pure sliding contacts. This further ascertains that pure sliding tests are good predictors of rolling–sliding tests, for the gear oils investigated in this study. Contact pressure affects both boundary and mixed friction but oppositely, i.e. an increase in contact pressure increases boundary friction while decreases mixed friction. It is important to conduct experiments in a wide range of conditions covering both low and high contact pressures and not solely make decision based on results obtained at one particular contact pressure, especially for gear oils that are used at high contact pressures.

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Acknowledgements This work was funded by the Austrian COMET Programme (Project K2 XTribology, no. 824187) and carried out at the ‘‘Excellence Centre of Tribology’’. The authors wish to thank Collini Applied Surface Intelligence, Evonik Industries AG, Oil Additives, High Tech Coatings GmbH, Magna Powertrain AG & Co KG, OMV Refining & Marketing GmbH for their financial support and active research cooperation. Also, the authors would like to thank Michael Schweitzer of AC2T research GmbH, Austria for conducting SRV experiments, and Boris Krzˇan and Prof. Mitjan Kalin of Laboratory of Tribology and Interface Nanotechnology, University of Ljubljana, Slovenia for their assistance in conducting MTM experiments. References [1] Gunsel S, Korcek S, Smeeth M, Spikes HA. The elastohydrodynamic friction and film forming properties of lubricant base oils. Tribology Transactions 1999;42(3):559–69. [2] Devlin MT, Senn J, Turner TL, Milner J, Jao TC. Reduction in axle oil operating temperature by fluids with optimized torque transfer efficiencies. Lubrication Science 2006;18:7–23. [3] Bronshteyn LV, Kreiner JH. Energy efficiency of industrial oils. Tribology Transactions 1999;42(4):771–6. [4] Greaves M. Pressure viscosity coefficients and traction properties of synthetic lubricants for wind turbine gear systems. Lubrication Science 2012;24:75–83. [5] Pressure–viscosity report: viscosity and density of over 40 lubricating fluids of known composition at pressure to 150,000 psi and temperatures to 425F, ASME, New York; 1953. [6] Spikes HA. Thermodynamic approach to viscosity. Tribology Transactions 1990;33:140–8. [7] Hoehn B-R, Michaelis K, Doleschel A. Limitations of bench testing for gear lubricants. In: Totten G.E., Wedeven L.D., Dickey J.R., Anderson M. (editors.). Bench testing of industrial fluid lubrication and wear properties used in machinery applications, ASTM: STP 1404; (2001). p. 15–32. [8] Castro J, Seabra J. Coefficient of friction in mixed film lubrication:gears versus twin discs. Proceedings of the Institution of Mechanical Engineers, Part J: Journal of Engineering Tribology 2007;221:399–411. [9] Brandao JA, Meheux M, Ville F, Seabra JHO, Castro J. Comparative overview of five gear oils in mixed and boundary film lubrication. Tribology International 2012;47:50–61. [10] Brandao JA, Meheux M, Seabra JHO, Ville F, Castro MJD. Traction curves and rheological parameters of fully formulated gear oils. Proceedings of the Institution of Mechanical Engineers, Part J: Journal of Engineering Tribology 2011;225:577–93. [11] Vergne P. Super low traction under EHD and mixed lubrication regimes. In: Erdemir JM, Martin JM, editors. Superlubricity. Amsterdam: Elsevier B.V; 2007. p. 429–45. [12] Costello MT. Effects of basestock and additive chemistry on traction testing. Tribology Letters 2005;18(1):91–7. [13] Hoehn B-R, Michaelis K, Mayer J, Weigl A. Influence of surface velocity directions on lubricant film formation in EHL point contacts. Tribology International 2012;47:9–15. [14] Spikes HA. Mixed lubrication—an overview. Lubrication Science 1997;9:221–53. [15] Muller M, Topolovec-Miklozic K, Dardin A, Spikes HA. The design of boundary film-forming PMA viscosity modifiers. Tribology Transactions 2006;49:225–32. [16] Topolovec-Miklozic K, Lockwood F, Spikes HA. Behaviour of boundary lubricating additives on DLC coatings. Wear 2008;265:1893–901. [17] Ingram M, Noles J, Watts R, Harris S, Spikes HA. Frictional properties of automatic transmission fluids: Part II-Origins of friction-sliding speed behaviour. Tribology Transactions 2011;54:154–67. [18] Reddyhoff T, Ku ISY, Holmes AS. Friction modifier behaviour in lubricated MEMS devices. Tribology Letters 2011;41:239–46. [19] Martini A, Zhu D, Wang Q. Friction reduction in mixed lubrication. Tribology Letters 2007;28:139–47. [20] Choo JH, Forrest AK, Spikes HA. Influence of organic friction modifier on liquid slip: a new mechanism of organic friction modifier action. Tribology Letters 2007;27:239–44. [21] Choo JH, Spikes HA, Ratoi M, Glovnea R, Forrest A. Friction reduction in lowload hydrodynamic lubrication with a hydrophobic surface. Tribology International 2007;40:154–9. [22] Kalin M, Velkavrh I, Vizintin J. The Stribeck curve and lubrication design for non-fully wetted surfaces. Wear 2009;267:1232–40. [23] Guo F, Wong PL, Geng M, Kaneta M. Occurrence of wall slip in elastohydrodynamic lubrication contacts. Tribology Letters 2009;34:103–11.