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A Review on Grease Lubrication in Rolling Bearings Piet M. Lugt
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SKF, Engineering & Research Center, P.O. Box 2350, Nieuwegein, 3430, DT, The Netherlands Published online: 15 Feb 2011.
To cite this article: Piet M. Lugt (2009) A Review on Grease Lubrication in Rolling Bearings, Tribology Transactions, 52:4, 470-480, DOI: 10.1080/10402000802687940 To link to this article: http://dx.doi.org/10.1080/10402000802687940
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Tribology Transactions, 52: 470-480, 2009 C Society of Tribologists and Lubrication Engineers Copyright ISSN: 1040-2004 print / 1547-397X online DOI: 10.1080/10402000802687940
A Review on Grease Lubrication in Rolling Bearings
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PIET M. LUGT SKF, Engineering & Research Center P.O. Box 2350 Nieuwegein 3430 DT, The Netherlands
Grease lubrication is widely applied to rolling bearings. The consistency of grease prevents it from leaking out of the bearing, makes it easy to use, and will give it good sealing properties. The same consistency prevents an optimal lubrication performance. Most of the grease is pushed out of the bearing during the initial phase of bearing operation and no longer actively participates in the lubrication process, leaving only a limited quantity available, which is stored inside the bearing geometry and on the bearing shoulders (covers or seals). This stored volume strongly determines the remaining lubrication process in the bearing. The distribution of this volume is determined by the grease flow, which is very complex to understand due to the strong nonlinear rheology. There is no consensus on the next phase in the lubrication process. The grease may bleed and provide oil to the raceway; it may be severely sheared in the raceway releasing oil; or small fresh quantities of grease may be sheared off from the volume stored on the shoulder. In addition, the lubrication process may be dynamic. Grease has self-healing properties where fresh grease is supplied in case of film breakdown and self-induced heat development. This article describes the state-of-the-art knowledge on grease lubrication, including grease flow, film formation, film reduction, dynamic behavior, and grease life.
the bearing is properly filled). The main disadvantage of using grease is its limited life. Mechanical work on the grease deteriorates its structure and in cases of high temperature, oxidation takes place (T > 120◦ C) (Ito, et al. (1)). Severe lubricant starvation occurs, causing bearing failures. This implies that the service life of the bearing may be determined by the life of the grease. In that case the bearings may need to be re-lubricated occasionally; i.e., filled with fresh grease. The bearing manufacturers have specified the re-lubrication intervals in their catalogues. These relubrication intervals are calculated from the life of the grease. Unfortunately, there is no absolute value for this. Even if bearings are running under very well-controlled conditions, such as in a laboratory situation, there is the usual significant spread in life. The re-lubrication interval is defined as the L01 of grease life; i.e., the time at which 1% of a population of bearings is expected to have failed (Huiskamp (2)). The challenge in grease research is threefold primarily. The first challenge is to develop greases that will give longer life and/or are able to operate under more severe conditions (extreme low and high temperature and speed). The second challenge is the development of predictive tools, such as numerical models or expert systems. The third challenge is to design bearing systems that would increase grease life by, for example, optimizing the grease flow. All these aspects require a fundamental understanding of the lubrication mechanisms of lubricating greases. The research efforts in grease lubrication have so far been relatively small. The global business for grease does not allow for large research programs. The bearing industry has a particular interest in understanding grease lubrication, though. More than 90% of all rolling element bearings are greased and sealed for life, effectively making grease a bearing component similar to rolling elements and seals. In addition, the internal design of the bearing has an impact on the performance of the grease. This article gives an overview of the existing knowledge on the various aspects of grease lubrication and the state-of-the-art models that exist in the public literature today. The main research on the lubrication mechanisms has been done on single contact configurations where the rolling element ring contact is simulated by a ball on a flat disc. This is generally allowed for studying EHL (elasto-hydrodynamic lubrication), where the contact geometry can easily be reduced to this configuration. The great advantage of this is the possibility to accurately measure film thickness using optical interferometry methods. The main drawback is the large difference in timescale
KEY WORDS EHL with Greases; Grease Application; Greases; Lubricant Degradation; Rolling Element Bearings; General; Starvation in EHL
INTRODUCTION The main role of grease in a rolling bearing is to provide the rolling element ring contact with a lubricant to ensure a separation of the two such that the bearing has a long life and low friction. The main advantages of using grease rather than oil lubrication are the ease of use (it will not easily leak out of the bearing due to its consistency), the inherent sealing action, the protection against corrosion, and low friction (provided that Manuscript received November 2, 2008 Manuscript accepted November 27, 2008 Review led by Andy Jackson
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between successive over-rollings and the impossibility to simulate the true lubricant flow and the feed and loss mechanisms, which will be described later. Moreover, the large centrifugal forces acting on the grease in the bearing cannot be simulated in these instruments. Nevertheless, despite these drawbacks, the observations using these instruments have contributed significantly to understanding the lubrication mechanisms in grease lubrication. Greases are classified primarily on the thickener type. In rolling bearings, lithium/lithium complex soap and polyurea greases are mostly used. A description of the various types can be found in any textbook on lubrication and goes beyond the scope of this article (e.g., Bartz (3) or Harris (4)).
GREASE PROPERTIES Grease Structure Grease is defined as “a solid to semi-fluid product or dispersion of a thickening agent in a liquid lubricant. Other ingredients imparting special properties may also be included” (NGLI (5)). The base oil is kept inside the thickener structure by a combination of Van der Waals and capillary forces (Bauer, et al. (6)). Interactions between thickener molecules are dipole-dipole including hydrogen bonding (Hurley (7)) or ionic and Van der Waals forces (Forster, et al. (8)). The effectiveness of these forces depends on how these fibers contact each other. The thickener fibers vary in length from about 1 to 100 µm and have a length to diameter ratio of 10 to 100 where this ratio has been correlated with the consistency of the grease for a given concentration of thickener (Scarlett (9)). It is not obvious how to visualize the structure of grease. In general, wet samples cannot be used in a scanning electron microscope. Figure 1 shows the structure of some greases where the oil has been carefully washed away using a non-polar solvent. Since grease contains 80–90% oil, one may argue that the thickener structure may collapse if the oil is washed out and that such a picture may be misleading. Other visualization techniques have been used as well such as a freeze-fracture technique (Magnin and Piau (10); Shuff and Clarke (11)), where a replica is made of a frozen grease sample, which can be observed in the SEM. Also, atomic force microscopy (AFM) has been used (Hurley and Cann (12)). Shin, et al. (13) visualized the grease flow in a shear field with an optical microscope in phase contrast mode. They observed very long fibers ranging from 50 to 100 µm, which are much longer than those observed with the SEM. An environmental SEM may be used but the low contrast between soap and base oil makes visualization difficult. Salomonsson, et al. (14) have visualized the grease structure using the cryo-TEM technique and visualized naphthenic lithium grease and paraffinic lithium grease by increasing the contrast between soap and oil through, e.g., replacing lithium by lithium/cesium soap.
Additives The role of additives in grease has not been explored in much detail. At high temperatures, the antioxidant additives will have the greatest effect. These additives are continuously consumed during bearing operation and, according to van den Kommer (15), totally consumed after 50% of grease life. Extreme pressure/anti-wear (EP/AW) additives are generally applied for
Fig. 1—SEM photographs of different grease soap structures: (a) lithium-12-hydroxy stearate in mineral oil, coarse structure; (b) lithium-12-hydroxy stearate in mineral oil, fine structure; (c) lithium-12-hydroxy stearate in ester oil, very fine structure; (d) modified lithium-12-hydroxy stearate in mineral oil.
low speed and/or high load. The effect on grease life that these additives have is not well understood. According to Gow (16) some 90% of all lubricant additives destroy the thickener structure of grease since they are often based on surface-active materials and this leads to what is commonly called the mayonnaise effect (softening and discoloring). He also mentions that of the remaining 10%, some 90% do not work. He ascribes this to the fact that the thickener material is almost always very polar (metallic soaps) and that the (also polar) EP additives will adhere to the soap structure rather than to the metal surface (Gow (17)). This is in contradiction to the results found by McClintock (18) who tested a number of greases on lubricant life and found an increase in life. A very promising development is the use of bismuth as an EP/AW additive because it is nontoxic and shows very good performance (Rohr (19)). Kaperick (20) shows in an evaluation of the “Timken OK Load Test,” in which identical EP additives give a different response to EP action for different formulated greases and ascribes this to a possible impact of mobility towards the surface through chemical interactions or attractive forces. If Gow is right then it is likely that the impact of EP additives on grease performance
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may be measurable through the mechanical and thermal aging of grease. This direct impact may be primarily on consistency and bleeding rate.
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Non-Newtonian Rheology A lubricating oil shows non-Newtonian behavior at high shear rates and pressure. Due to the thickener structure and its interaction with the base oil, grease shows this behavior already at very low shear rates and ambient pressure. Measurements from low to high shear rates can be found in Pavlov and Vinogradov (21). The solid-like behavior, or resistance to flow (or leakage), is traditionally characterized through the consistency or penetration, measured using a cone penetrometer (ISO 2137, ASTM D217), which is translated into an NLGI consistency number. An alternative is to measure the yield strength on a rheometer, by means of the computerized evaluation of yield value (CEY; Gow ´ et al. (23); (22)) or through the “cross-over stress” (Couronne, Couronne´ and Vergne (24)). A correlation between yield stress and penetration/consistency can be found in Couronne´ et al. (25). Generally, this is only determined at room temperature. It is clear that the grease will be severely worked in the bearing. This applies to the grease that is being churned between the rolling elements but also applies to the fraction that passed the EHL contacts where shear rates are O(106 s−1 ). This causes a rapid change in rheological properties of the grease during the initial phase of bearing operation. It is therefore relevant to measure the rheology after working the grease. This can be done in a rheometer itself, in a grease worker (26), or in a Shell roll tester (27). The change in consistency loss is quantified by measuring the mechanical stability of the grease before and after working the grease. The yield stress strongly depends on temperature. Measurements for different types of grease can be found in Karis, et al. (28) or Czarny (29). Karis shows that the yield strength of a lithium grease may drop from 500 Pa at 20◦ C to 100 Pa at 60◦ C. As shown by Forster and Kolfenbach (30), greases show viscoelastic behavior. There are a number of models proposed for low and high shear rates. The best known are the power law, RheeEyring, Bingham and Herschel-Bulkley models. A definition of these models can be found in Yousif (31). These models assume solid or very high viscous behavior at low shear rates and viscous (with possible shear thinning) at higher shear rates. An example can be seen in Fig. 2. In addition to this nonlinear shear stress-shear rate behavior, grease is thixotropic, meaning that the measured stress also depends on time. Paszkowski (32) defined thixotropy as an isothermal decrease in structural (apparent) viscosity during shearing (at both constant and variable shear rates) followed by an increase in the viscosity and the re-solidification of the substance once shearing ends. Pavlov and Vinogradov (21) show creep-like behavior at very low shear (106 s−1 ) until the yield strength has been reached. As shown by Hurley and Cann (33), grease rheology approaches that of the base oil at high shear rates. They also show that mechanical work changes the rheological behavior significantly. The grease thickener structure is thus continuously degrading and is transformed from a Bingham plastic or Herschel-Bulkely material towards a more viscous material ´ (Merieux, et al. (34)).
LUBRICATION MECHANISMS Lubricating Conditions There is no consensus on the lubrication mechanisms in grease lubrication. There is an overall agreement though that greaselubricated bearings are generally running under starved lubrication conditions. This has been shown by Poon in 1972 on a disc machine (Poon (35)) and by Wilson in 1979 (36) in full bearings. Wilson measured the film thickness in cylindrical and spherical roller bearings and showed that the lubricant film initially exceeds the value in case of fully flooded oil lubrication by 20–25%. However, already after a few hours the film thickness has decreased below this value. At this point the bearing runs under starved lubrication conditions. Barz (37) measured the film thickness in a cylindrical thrust bearing as a function of bearing speed. His measurements show that the film thickness is relatively large at low speed but decreases with speed up to a speed where the film stays ¨ and Jacobson (38) measured the electrical constant. Wikstrom capacity over the bearing contacts (using a method developed by Heemskerk, et al. (39)) in a grease lubricated spherical roller bearing and showed that metal-to-metal contact occurs very regularly, meaning that the films are very thin, certainly smaller than the values that could be expected assuming fully flooded conditions. As a rule of thumb approximately 30% of the free volume of the bearing should be initially filled with grease. It will be clear that this is much more than required to provide the bearing with a (fully flooded) lubricant film. In the beginning, excessive grease churning, or grease flow, takes place, which is responsible for the high temperature peak caused by the churning component of the friction torque. The initially thick lubricant films in the beginning indicate that at least during this initial bearing operation thickener enters the contact. ˚ om, ¨ Single contact measurements by Astr et al. (40), Williamson, et al. (41), and Kaneta, et al. (42), using a scoop to ensure fully flooded conditions, have shown that the film thickness is indeed higher than the fully flooded oil film thickness. The optical setup also made it possible to show that grease thickener lumps were entering the contact. The literature does not reveal if this is really restricted to the initial phase of bearing operation. It might be that this is very pronounced in the beginning and diminishes slowly over time. The most widely used model to describe grease lubrication is that the grease acts as an oil reservoir where the oil is slowly released into the running track (Booser and Wilcock (43)). Lubrication guidelines are then formulated according to this ability to bleed (Baker (44)). There is definitely no consensus here. Already in 1967 Scarlett (9) referred to an alternative mechanism of a high viscosity layer retained within the rolling track. By means of FTIR (Fourier transform infra-red) spectroscopy, Cann and Spikes (45) and Cann, et al. (46) observed thickener layers on the surfaces of a ball-on-disc machine. They proposed a model where it is assumed that the surfaces are covered by a thin layer of soap and where the film is formed by base oil thickened with broken thickener fibres. Cann and Hurley (47) conclude from their experimental work that this is the result of
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Fig. 2—Shear stress versus shear rate curve for lithium soap bearing grease at room temperature. Measurements and fit using a power law model.
a progressive destruction of the soap matrix by over-rolling. This destruction releases oil, which provides free lubricant for replenishment. This is an interesting hypothesis: oil is not released by bleeding but by the destruction of the thickener. Farcas and Gafitanu (48) confirmed this. There has been very little work published on grease analysis in full bearings. Cann, et al. (49), (50) have investigated the chemical composition in ball bearings. They investigated the grease in R0F (6204-type ball bearings) and R2F tests (6209-type ball bearings). Surprisingly, the differences between R0F and R2F tests are not only related to bearing size. For the R2F test they write that initially, grease is over-rolled, releasing free oil through degradation. Simultaneously, grease is pushed to the side, onto the seals. In the next phase, grease is sheared from the seal into the raceway where it again degrades into an oil-like lubricant (although patches of grease were also found). This lubricant moves onto the balls into the pocket. In the cage pockets oil was found. Significant amounts of free oil could not be found in the R0F, though. This means that in the R0F test configuration, grease is sheared into the contacts and into the cage pockets where it is over-rolled and sheared and where oil is released. So for this test, the grease on the shields did not serve as an oil reservoir. This agrees with the mechanism suggested in Cann and Spikes (45) and Cann, et al. (46). Scarlett (9) described the flow of grease through a ball bearing with an inner ring guided brass machined cage and mentions the formation of “pads” of grease adhering to the cage bore (under the cage bars). These pads had a higher consistency with a higher soap concentration than the original grease. Scarlett explicitly states that this is due to oil loss (bleeding), which occurs during the first 100 h operation and, according to him, does not contribute to feeding oil to the bearing after
this. This statement is not based on any experiments in this article though. Scarlett describes tests where grease has been removed from the covers after the initial churning period. In that case he found premature bearing wear. This means that the grease on the covers plays an important role in lubrication after the initial phase. He investigated this role further by performing experiments using a tracer in the base oil of various greases on the covers only. Surprisingly, he found no flow of oil or grease from the covers into the bearing. He has done his tests with various grease types. A similar conclusion is made by Milne, et al. (51). Scarlett concludes that, after the churning period, there is no grease or base oil flow from the housing recesses into the bearing and postulates that its function is to form a closely fitting seal to prevent escape of essential lubricant from the bearing. Lansdown and Gupta (52) write that there is considerable evidence that in ball bearings the whole of the grease is involved in the lubrication process, so not only the bled base oil. They have various arguments for this, the most important being that the performance in ball bearings is comparable between the case of grease plating (technique where grease is coated on to the bearing surfaces) and the case of conventional grease lubrication. In their analysis on ball bearings they also write that often grease adjacent to the races is softer and has a higher oil content than the grease near to the outside of the bearing covers. Unfortunately, these statements are not illustrated with any examples, proof, or references. ¨ Full bearing tests (spherical roller bearings) from Wikstrom ¨ and Hoglund (53) using both grease and base oil only showed equal friction torque. This shows the importance of the base oil. They claim that these tests confirm the theorem from Booser and Wilcock (43) where grease releases oil, which then lubricates the
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bearing. A similar conclusion is drawn by Dalmaz and Nantua (54), who tested six lithium greases, varying base oil viscosity, thickener structure, and concentration on both angular contact ball bearing (ACBB) life test rigs and on a single contact film thickness instrument. Similar to Hurley (7), they report that the initial film thickness is proportional to the thickener concentration and larger than those of the base oil. However, their bearing life tests show that bearing life is related to base oil viscosity only and not to thickener type. This suggests that the grease film may last only very briefly and the film will be formed by base oil only for the main part of the life of the bearing. Also, Mas and Magnin (55) have investigated grease before and after running in full bearings. They tested tapered roller bearings and found an increased viscosity of the grease and reduced oil content under the cages. This implies that grease bleeding occurs from grease located under the cage bar. They also show by means of SEM the destruction of fibers in the raceway confirming the Cann et al. theorem (Cann Spikes (45); Cann, et al. (46)) again.
Film Thickness and Starvation As mentioned previously, the initial film thickness is higher than can be expected based on the base oil viscosity only. The initial film must therefore consist of (sheared, degraded) grease. After some time, starvation occurs where the film is reduced in thickness. The initial film thickness has been modeled by Dalmaz and Nantua (54) and Hurley (7) by assuming that the initial film thickness is proportional to the thickener concentration and larger than that of the base oil. Hurley developed an empirical formula for this. Others have used the grease rheology as input for a model. Jonkisz and Krzemiski-Fredihave (56) and Kauzlarich and Greenwood (57) used a Herschel-Bulkley model. Bordenet, et al. (58) used a “four parameter rheology model,” which is quite similar to the Herschel-Bulkley rheology model. They all found slightly higher values of the film thickness compared to those calculated using the base oil viscosity only. Yang and Qian (59) used a Bingham rheology model to predict the film thickness. They showed that the film thickness, again for fully flooded conditions, can be calculated by using the conventional EHL formula whereby the viscosity of the grease at high shear rates should be used, rather than the oil viscosity. Aihara and Dowson (60) performed an experimental study of the factors affecting film thickness in a grease lubricated two-disc machine. They suggest that the grease lubricated film thickness can be estimated by taking 70% of the value of the fully flooded film thickness using the base oil viscosity. This is in accordance to Saman’s (61) theory, who assumed that the contacts will ultimately be so starved that the inlet meniscus will be so close to the Hertzian contact that zero-reverse flow can be assumed. Theoretically this will lead to a reduction to 71% of the fully flooded film thickness. The reduction on film thickness after the initial phase need not totally be ascribed to classical starvation. Kauzlarich and Greenwood (56) show that shear degradation of the grease also leads to a reduction of film thickness in time.
Starvation models for oil-lubricated contacts have been developed by Chevalier, et al. (62), later refined by Damiens, et al. (63) and Damiens (64). In these models, a given layer of oil with a given thickness is supplied to the EHL contact. The main problem here is the input layer thickness, which is not known for a bearing with multiple contacts. This layer needs to be calculated based on the nonuniform layer, which leaves the preceding EHL contact and the feed and loss of lubricant in-between over-rollings (or rather in-between two rolling elements). The first problem is the possible occurrence of replenishment between two consecutive rolling elements. Van Zoelen, et al. (65) has recently developed an innovative approach to this by assuming that starvation is caused by side flow in the EHL contacts only and that track replenishment can be neglected. They showed that the starved film thickness increases with increasing load caused by increasing viscosity inside the contact and therefore a reduced side flow.
Track Replenishment Replenishment has been the topic of many articles starting with the model from Chiu (66). Jacod, et al. (67) showed that surface tension-driven replenishment of the running track is too slow to give any significant effect and shows that only capillary forces may have some effect. Recently, Gershuni, et al. (68) investigated the impact of bearing centrifugal forces on the layers formed behind a contact of cylindrical roller bearings. They showed that these forces increased the replenishment rates enormously in case of outer ring rotation. In case of inner ring rotation, replenishment is seriously retarded by the centrifugal action on the layers. They showed that replenishment times, even in the absence of centrifugal forces, were too long to have any effect. Farcas and Gafitanu (48) developed a model based on the wetting properties of the lubricant only. They calculated the critical speed at which lubricant droplets are no longer able to adhere to the surface due to the centrifugal forces. They have validated their model using electrical resistance measurements over the bearing contacts. ´ Merieux, et al. (34) show that grease shear degradation may also lead to replenishment and film growth. They assume that the grease next to the contact is continuously sheared and therefore degrades. The grease will ultimately lose its grease-like behavior and will then behave as an oil, replenishing the contact. Van Zoelen, et al. (69) investigated the impact of the tangential component of the centrifugal forces on the thin-film flow on tapered and spherical roller bearing inner rings. They showed a significant effect here, which should certainly be incorporated in any replenishment model. The cage design also plays a role in the prediction of the film thickness and track replenishment. Damiens, et al. (70) have done film thickness measurements on a single contact where they mounted a single cage pocket, cut from a full cage, on their ball-on-disc device and where they were able to vary the clearance between cage and ball from 0.05 to 0.5 mm. They show that, in case of oil lubrication, the film thickness decreases with decreasing clearance between cage and ball. They ascribe this to a “scraping” effect. However, in the case of grease lubrication they found an inverse effect—the film thickness was
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Grease Lubrication in Rolling Bearings
increasing with decreasing cage clearance. They concluded that a cage probably redistributes the grease in the cage/ball contacts, preventing starvation. As they mention in their article, both effects, redistribution and scraping, may take place in a bearing. Also, in the case of grease lubrication scraping may dominate if the clearance is too small. In addition to scraping/redistribution, the cage may also operate at much higher temperature than the other bearing components, affecting the flow but also the possibilities to maintain a grease reservoir under the cage. The yield stress is strongly dependent on temperature (Czarny (29)). Generally, for practical reasons, the bearing temperature is measured on the outer ring and the cage temperature is never reported. Joshi, et al. (71) have performed temperature measurements on the cage of a tapered roller bearing (TRB). The bearing was running in an oil bath (75% fill). They recorded the temperature of both housing and cage and showed that the cage temperature response is much higher to changes in lubrication than the housing temperature. In addition to lubricant replenishment, lubricant is supplied by bleeding from the grease, either from under the cage or from the grease attached to the covers. No grease bleeding models have been published as yet. Then, finally, evaporation (Komatsuzaki, et al. (72)) and oxidation need to be taken into account, which will reduce the layer in between the rolling elements again. At low temperatures, these effects may be neglected, though.
Dynamic Behavior In 1996, Mas and Magnin (55) speculated on the release of fresh “grease” after heat development caused by film breakdown. They wrote that a grease lubricated bearing will fail as soon as this can no longer take place. This would imply a dynamic behavior of subsequent film breakdown and repair. Such dynamic behavior ¨ and Jacobson can be observed in the measurements of Wikstrom (38), who tested spherical roller bearings, measuring the electrical capacitance across the contacts. Unlike Barz (37) and Wilson (36), who used a similar technique, they did not translate this signal to film thickness values. Their intent was only to illustrate the occurrence of metal-to-metal contact. Their measurements show that the capacitance signal was not constant but had a dynamic character. Cann and Lubrecht (73) illustrated this in their ball-on-disc machine and showed that severe starvation can be “repaired” by adding additional lubricant to the contact. Very pronounced dynamic behavior was observed in 2008 in Lugt, et al. (74), where temperature signals from cylindrical roller bearing tests were analyzed and where it was shown that these signals show “events” characterized by periods where the temperature rises significantly and falls back to the steady state. Moreover, additional analysis of the electrical resistance measurements proved that these events are caused by film breakdown followed by replenishment and that grease lubrication exhibits “deterministic chaotic” behavior. This had not been observed so explicitly before for two reasons. In the first place, a sufficiently long test time is required to measure this, which is usually not applied because of high costs for testing or lack of patience. The second reason is that most grease life tests
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are done under controlled temperature. It is expected, though, that the temperature controllers will show the same dynamics as observed in the temperature signal obtained in tests where bearings are running under self-induced temperature conditions.
Grease Flow During the churning phase in a grease lubricated bearing, most of the grease is pushed onto the covers/seals/shields of the bearing arrangement and some grease stays inside the bearing where some of it may be mobile and continuously flow and part of it will form patches of grease that can be found under the cage. The amount of grease that can be stored here obviously depends on the cage geometry and the flow properties of the grease; i.e., its rheological behavior. Also, in case of vertical shaft arrangements or in case of vibrations, the amount of grease available for lubrication depends on these properties. Under these circumstances generally a high consistency grease is used to prevent grease falling back into the track again and to maintain a lubricant reservoir adjacent to the rolling elements row(s). According to Cobb (75), there is no difference in grease performance (start-up torque, temperature, leakage through seals) if ball bearings are filled from one side only, provided the same total amount of grease is placed in the bearing under either placement condition. This indicates that most of the grease in the bearing participates in the flow. The flow of grease inside the bearing is very complex. Visualization techniques have shown flow patterns, (9). However, a quantitative prediction and therefore prediction of the formation of the grease reservoir is not possible today. There are several reasons for this. The first reason is the complex rheology of the grease. In addition, thickener-oil separation could occur with entrapped air. A multi-phase flow may occur, which is most difficult to describe using today’s computational tools. Moreover, there is an enormous variation in scale and shear rates inside the bearing configuration. Between the rolling elements, clearly churning takes place with relatively low shear rates. In the inlet of the contacts, phase separation may occur (similar to what happens with water in emulsions in the inlet of EHL contacts), a jet flow may occur, or even droplets may be formed (Larsson, et al. (76)).
Wall Slip The flow is not only determined by the rheological properties of the grease. The bearing (cage) material and roughness are also important. The roughness of cage material and type of cage material are important due to the occurrence of wall slip. This wall slip effect has been studied in rheometers and in pipe flows. Forster, et al. (8) claim that the flow close to the wall is restricted by the breaking fiber contacts. They report that at high slip rates internal slip in the fibers would be responsible for wall slip. Bramhall and Hutton (77) ascribe wall slip to a lower concentration of the thickener particles at the wall, so slip over a layer of oil. Czarny’s experiments (Czarny (78); Czarny and Moes (79)) show that the wall slip depends on the wall material and on the thickener type. Since this does not apply to oil lubrication, he claims that the formation of a wall layer is a result of interactions between the particles of the grease thickener and the wall material resulting in a concentration gradient of thickener close to the wall. He based
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this explanation not only on his own work but also on the experimental work of Vinogradov, et al. (80). Delgado, et al. (81) showed the impact of roughness on wall slip though measurements of pressure drop and flow through pipes. They showed that smoothening of the wall gave a significant reduction in pressure drop. This can only be ascribed to wall slip again. The wall slip phenomenon is not only of importance to understand grease flow, it may also be of importance for the ability of the bearing to maintain its grease reservoir after the churning phase. In case of line contact bearings grease may slide away from under the cage bar into the track.
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GREASE LIFE Existing Grease Life Models Grease life models have been developed mainly by the bearing manufacturers. The scientific developments are still very limited and there is still much to be done for the development of a true physical grease life model. Therefore, all published grease life models are empirical; i.e., based on grease life testing. Such tests are done on R0F, R2F (machine described in DIN 51806) or FE9 machines (described in DIN 51562). It is generally believed that grease life follows the Weibull probability density distribution and therefore a sufficiently large number of bearings need to be tested. The main parameters determining grease life are ndm, (the product of rotational speed and mean bearing diameter) and temperature. This ndm parameter translates into a peripheral speed. All models assume a temperature-Arrhenius behavior. Ito, et al. (1) showed this in their extensive grease life test program for small deep-groove ball bearings at temperatures exceeding 100◦ C. The effect of load is generally less pronounced, which may be due to the weak relation between load and EHL film thickness. Deep-groove ball bearings (DGBBs) are “easier” to lubricate than other bearing types. This is related to more pronounced replenishment (Damiens, et al. (63)) or is sometimes ascribed to ball spin. Most models are normalized for this bearing type and correction factors are applied for the other bearing types. In addition to the starvation effect, some bearings, such as tapered and spherical roller bearings, and also angular contact ball bearings, show an inherent pumping effect, which reduces the available lubricant in the running track significantly (1) van Zoelen, et al. (82). The models are only applicable if the bearing operates in the temperature domain for which the grease has been designed. The lower temperature limit is usually determined by the bleeding properties of the grease or the base oil’s pour point, whereas the high temperature limit is determined by the dropping point of the grease; i.e., the point at which a droplet falls from a standardized cup. This point is accepted as the maximum temperature at which the grease can be exposed without losing its structure. For safety reasons this is reduced by 15–20◦ C. The models can be found in the catalogues of the various bearing manufacturers and are applicable for a certain type of grease despite the fact that significant quality differences can be found. Huiskamp (2) incorporated the grease quality into his model for
deep-groove ball bearings by means of grease performance factors (GPFs), which is shown in Fig. 4. The GPF is defined as the ratio of real life and the grease life predicted by his model. In order to determine its values grease life tests are necessary, such as the R0F/R0F+ test. As mentioned above, the initial filling rate is important. Generally, the models apply to an optimally filled bearing. To the author’s knowledge only Farcas and Gafitanu (49) have included the initial volume of grease inside the bearing as a parameter in a life mode. Their empirical model is based on a limited data set, though. The effects of shock loads and vibrations are also often incorporated through penalty factors. These effects cause grease lumps from the covers/seals to fall into the bearing, resulting in high temperatures and loss of grease reservoir. The same applies to the effect of air flow through the bearing. Lubricant droplets formed behind the rolling elements (Larsson, et al. (76)) will be dragged out of the bearing and will no longer replenish the inlet of the next rolling element. Air flow will also have an impact on the evaporation rate, especially at higher temperatures. Lansdown and Gupta (52) show that evaporation of base oil not only happens on thin films bled out of the grease but may even happen to oil contained in grease directly.
Reliability Bearing life is usually expressed in L10 life (10% of the population will have failed), and grease life is usually expressed in L50 life (50% of the population will have failed). This is the result of test practice where grease life is measured using a limited number of tests representing a large quantity of bearings. A two-parameter Weibull distribution is assumed, since a minimum
Fig. 3—Balance between feed and loss of lubricant ultimately determin¨ and Jacobson (38) ). ing the lubricant film thickness (Wikstrom
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Fig. 4—Grease life model from Huiskamp (2) including grease performance factors.
of guaranteed life, requiring a third parameter, is doubtful for grease lubrication. The distribution can be obtained from the R0F test where the life of five pairs of bearings is evaluated using the “sudden death principle.” Here a pair of bearings is stopped as soon as one bearing has failed. So the test will give five failed and five suspended bearings. Weibull statistics are
applied to these results, giving the full failure distribution and corresponding confidence interval. As illustrated in Fig. 5, by taking L50, the smallest confidence interval is obtained (Andersson (83)). The confidence intervals—i.e., the precision of life estimates—for L1 are very large, which makes it difficult to discriminate between test results. Grease life is therefore
Fig. 5—Failure distribution for a grease life test using the R0F test rig using a population of 10 grease failures.
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generally measured through the L50 value. Re-lubrication intervals are derived from that using a fixed Weibull slope. ROF testing gives an average Weibull slope β = 2.3, which gives L10 ≈ 2.7L01 and L50 ≈ 5L01. Note that for ball bearing fatigue life, typically β = 1.1 (Harris (4); Lundberg and Palmgren (84); Ioannides, et al. (85)). This means that the spread in grease life is less pronounced than that in bearing life.
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DISCUSSION AND CONCLUSION The various hypotheses on the mechanisms of grease lubrication, all based on observations/measurements, indicate that there may be no unique mechanism. As an example, at low temperatures oxidation and evaporation will not give a significant contribution to “grease aging.” At high temperatures oxidation will dominate. Some metals catalyze oxidation (brass cages!). This may be one explanation why there is no consensus on the mechanism. It is certain that initial filling plays a major role. Too much grease leads to excessive churning, high temperatures, and severe grease degradation. If the bearing is properly filled, two phases can be distinguished; i.e., a churning phase where excessive grease will be pushed to the shoulders of the bearing onto the seals/covers. This process is determined by the flow properties of the grease (rheology). Prediction of flow is very complex due to the nonlinear grease rheology and the two/three phases involved (thickener, oil, air). The remaining grease inside the bearing will be over-rolled, where the thickener structure will be broken down, releasing oil, and where the thickener material could form a thin layer (Cann and Spikes (45); Cann, et al. (46)) or a high viscous layer (Scarlett (9)). After this phase the “bleeding” phase takes place. This phase is characterized by starvation. The lubricant film thickness is initially larger than calculated using the base oil viscosity. Side flow reduces the film thickness. Electrical resistance measurements in full bearings confirm the occurrence of starvation. The grease reservoir may be formed by grease under the cage (Scarlett (9)) and/or on the bearing shoulders (Cann, et al. (49)). However, there are also hypotheses that the bearing simply runs on the initial layer throughout its life-time and will not be replenished at all Scarlett (9). Scarlett (9) states that the grease on the shoulders plays a major role in providing a long grease life. He ascribes its role to provide an excellent sealing. Such sealing may indeed prevent side flow and reduce starvation. This would also explain why Lansdown and Gupta (52) found an increase in oil content of grease close to the raceway. At least for line contact bearings it has been clearly proven that grease located under the cage bars is bleeding oil (Mas and Magnin (55)). It is not clear, though, to what extent this bleeding contributes really. The 100 h that Scarlett mentions is very short. Prediction of film thickness in grease lubricated bearings is very complex. The film is a result of feed and loss mechanisms where bleeding, or grease creep flow, is the feed mechanism and where side flow (starvation), oxidation, polymerization, and ¨ and Jaevaporation are the loss mechanisms (Figure 4, Wikstrom cobson (38)). Side flow may be hindered by the excellent sealing from grease located on the shoulders of the bearing. For some bearing types an additional loss mechanism is formed by pumping, which takes place due to the tangential component of the
centrifugal forces on roller and rings (van Zoelen, et al. (65); van Zoelen, et al. (69)). At low shear rates grease creeps, so it may be that grease very slowly flows into the track. An additional complexity is the dynamic behavior of grease lubrication. The grease lubricated bearing shows an inherent “self-healing” mechanism where replenishment may happen due to film breakdown resulting in metal-to-metal contact, local heat development, and release of grease into the raceways (Mas and Magnin (55); Cann and Lubrecht (73); Lugt, et al. (74)). This means that the life of the grease in bearings cannot easily be predicted based on film thickness calculations only. Ultimately, knowledge on multi-phase flow, nonlinear rheology, EHL theory, and chemistry need to be combined to develop predictive models for grease lubrication in rolling bearings.
ACKNOWLEDGEMENT The author would like to thank Prof. E. Ioannides, Technical Director Product R&D for SKF, for his kind permission to publish this article.
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