Grease Lubrication in Rolling Bearings

By Piet M. Lugt

The definitive book on the science of grease lubrication for roller and needle bearings in industrial and vehicle engineering.

Grease Lubrication in Rolling Bearings provides an overview of the existing knowledge on the various aspects of grease lubrication (including lubrication systems) and the state of the art models that exist today. The book reviews the physical and chemical aspects of grease lubrication, primarily directed towards lubrication of rolling bearings.

The first part of the book covers grease composition, properties and rheology, including thermal and dynamics properties. Later chapters cover the dynamics of greased bearings, including grease life, bearing life, reliability and testing. The final chapter covers lubrications systems – the systems that deliver grease to the components requiring lubrication.

Grease Lubrication in Rolling Bearings:

• Describes the underlying physical and chemical properties of grease.
• Discusses the effect of load, speed, temperature, bearing geometry, bearing materials and grease type on bearing wear.
• Covers both bearing and grease performance, including thermo-mechanical ageing and testing methodologies.

It is intended for researchers and engineers in the petro-chemical and bearing industry, industries related to this (e.g. wind turbine industry, automotive industry) and for application engineers. It will also be of interest for teaching in post-graduate courses.

• Language: English
• Publisher: Wiley
• Release date: Dec 10, 2012
• ISBN: 9781118483978

Book preview

1

Introduction

1.1 Why Lubricate Rolling Bearings?

Rolling motion can be used to carry and transmit load while facilitating movement with very low friction and low wear rates, even in the absence of lubrication. The best known example where this is used is the wheel, invented by Mesopotamians in ca. 3500 BC. Lubrication of wheel–road (or later wheel–rail) contact is difficult, but even in the absence of lubrication wear rates are much lower than those of, for example, sledges or sliding shoes. The rolling bearing is based on this principle, although the configuration is more complex since, for carrying a single load, several rolling elements are used, which have a double contact (with the inner-ring and the outer-ring). Unfortunately, even in the apparently rolling contacts, slip occurs. This is partly due to the elastic deformation of the bodies in contact, which flattens the contacts to some extent, and partly due to the kinematics in the bearings. The first effect is usually very small (and can be decreased by using materials with a high elastic modulus). The second effect is more severe. The first effect is dominant in the contacts on tapered and cylindrical roller bearing raceways, which can run at very low friction levels (note that this does not apply to the contacts on the flanges in these bearings). For other bearing types, sliding profiles in the contacts between rolling element and rings typically show just one or two points of pure rolling. Positive slip occurs between these points and negative slip outside these points. This is shown in Figure 1.1 for a thrust spherical roller bearing.

Figure 1.1 Slip in a spherical roller thrust bearing. Reproduced from Olofsson, 1997 © Elsevier.

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In the absence of lubrication, the surfaces will be in intimate contact, resulting in high friction and wear at the areas where slip occurs. This will produce high stresses close to the surface, leading both to reduction of the fatigue life of the bearing and also to wear.

The occurrence of wear in the slip zones and the absence of wear in the points of pure rolling will produce a nonuniform wear profile across the tracks, leading again to high stresses at the zero-sliding points where less wear has taken place, with a corresponding reduction in the life of the bearing. This does not mean that rolling bearings cannot run in the absence of lubrication, but with no lubrication the service life will be impaired.

Full separation of the surfaces in contact, or ‘full film lubrication’, is preferable. In this case there is virtually no wear and the life of the bearing will be determined by fatigue. If full film conditions are not possible the materials should preferably be ‘incompatible’, meaning that adhesion and ‘welding’ can be avoided. This can be done by using ceramic rolling elements, for example, or by applying a suitable coating (or surface treatment) on one or both surfaces in contact. Due to the local sliding conditions a coating will wear and the service life of the bearing is determined by the wear rate and thickness of the coating. Nevertheless, the relatively short unlubricated life can be increased substantially by this solution. The advantage of using fluid lubrication is its ability to repair itself after shear in the contacts due to its ability to replenish the contacts (a self-healing mechanism). If a sufficient quantity of lubricant is available, this will happen through churning, splashing or will be flow-induced by the geometry of the bearing (by the pumping effect and centrifugal forces). In the case of grease lubrication, it occurs primarily through oil-bleeding, spin, cage distribution and to some extent through a centrifugal force inducing flow in thin lubricant layers.

1.2 History of Grease Lubrication

The word ‘grease’ is derived from the Latin word ‘crassus’ meaning fat. As far back as 1400 BC, both mutton fat and beef fat were used as axle greases in chariots (journal bearings). Early forms of grease lubricants before the 19th century were largely based on natural triglycerides, animal fats and oils, commonly known as ‘grease’ (Polishuk [475]).¹ Partial rendering of fats with lime or lye would produce simple greases that were effective as lubricants for wooden axles and simple machinery. Triglycerides are good boundary lubricants that show low coefficients of friction but they show poor oxidative stability at elevated operating temperatures.

After the discovery of oil in the USA (Drake) in 1859, most lubricants were based on mineral oil [450]. The first ‘modern’ greases were lime soaps or calcium soaps, which today are not much used in rolling bearings. They may however be used providing that the temperatures stay low. Later, aluminium and sodium greases were developed, which could accommodate higher temperatures. Until the Second World War only these calcium, sodium and aluminium greases were used.

In the 1930s–1940s new thickeners were discovered for multipurpose greases, based on calcium, lithium and barium [450]. In 1940 the first calcium complex grease and lithium grease patents [182] were issued. Today, over 50% of the market still consists of lithium grease. Aluminium complex greases were developed in the 1950s and lithium complex greases in the 1960s. Polyurea use, started in the 1980s, especially in Japan.

In 1992 a new type of grease was invented by Meijer [414], where the thickener comprises a mixture of a high molecular and low molecular weight polymer of propylene. A grease structure could be obtained through rapid quenching. This type of grease has been successfully tested and is used today in, for example, paper mill bearings [73]. Another example is nanotube grease[271,272].

Grease lubricants are used in a large variety of environments. Operating temperatures for grease lubricated applications range from subzero, −70 °C to temperatures exceeding 300 °C for high temperatures applications. They are also used in vacuum atmospheres encountered by space applications. More often, the operating environment involves wet and humid atmospheres, exposure to salt water and many other types of corrosive agents that affect the performance of rolling bearings and machine elements. The chemical composition of grease lubricants varies considerably to accommodate the large variety of applications and extremes in operating environments. Grease is commonly used for rolling bearing lubrication as a cost-effective and convenient source of lubrication.

1.3 Grease Versus Oil Lubrication

As mentioned above, the longest service life can be obtained if the lubricant film fully separates the contacting surfaces. In a rolling bearing this is achieved through hydrodynamic action where the lubricant is sheared inbetween the roller–ring contacts. Once inside these contacts the viscosity becomes so high, due to the high pressures, that leakage (pressure-driven flow) out of the contact will remain very small. It will be shown later, in Chapter 9, that this film thickness depends on oil viscosity and bearing speed. Obviously, a film can only be maintained if sufficient oil is available. In oil bath lubrication this is not a problem, but in the case of grease lubrication this is more difficult. The lubricating grease will generate a thick film at the beginning of bearing operation, formed by the combination of thickener and base oil. Side flow occurs due to the pressure difference inside the bearing contacts and next to the tracks. There may be very little reflow back into the track and the bearing may suffer from starvation, with thinner films then expected based on EHL (Elasto-Hydrodynamic Lubrication) theory.

Inside the bearing contacts (micro) slip occurs and heat will be generated. In the case of oil lubrication, the oil will act as a coolant for the bearing, reducing the temperature rise and therefore maintaining a sufficiently high viscosity and film thickness. Unfortunately, this is not possible in grease lubrication. There is generally no flow here and therefore no cooling effect by the lubricant.

High temperatures, mechanical work and the build-up of contaminants cause aging of the lubricant. In the case of oil lubrication this will be small due to the cooling and replenishment action. Unfortunately the effect of aging cannot be neglected in grease lubrication. Aging will primarily occur through oxidation of the base oil and thickener and through the breakdown of the structure. A long service life therefore often requires periodic replenishment through active relubrication (systems). Sometimes, the specific rheological behaviour of grease creates difficulties in centralized lubrication systems (pumpability).

Despite the above mentioned drawbacks, there are also clear advantages in using grease as a lubricant. Generally, friction levels are lower than in the case of oil lubrication, primarily due to the absence of churning, apart from the start-up phase. The next advantage is the ease of operation. Sealed and greased-for-life bearings do not require oil baths, which may leak. A well designed bearing with good quality grease requires no maintenance. In addition, the grease will fulfil a sealing function and form a barrier against entry of contaminants onto the raceway, extending the service life of the bearing.

For the selection of oil, the main parameters are: viscosity, boundary lubrication properties (lubricity) and type of additives. In the selection of grease the properties of the thickener dominate, but again the oil base stock properties are important. The main parameters are: consistency, operating temperature range, oil bleeding properties, viscosity of the base oil, corrosion inhibiting properties (additives) and load carrying capacity. This makes grease selection much more complex than oil selection. In this book the various aspects of grease lubrication in rolling bearings will be described, that is the lubrication of the bearing, the lubrication of the seal, lubrication systems, condition monitoring techniques and test methods. In the next chapter (Chapter 2) the lubrication mechanisms will be described. This chapter will touch upon many items that will be described in the following chapters, such as ‘film thickness’ (Chapters 9 and 10), ‘rheology’ (Chapter 5), ‘flow’ (Chapter 6), ‘oil bleeding’ (Chapter 7), ‘aging’ (Chapter 8) and ‘dynamic behaviour’ (Chapter 11). A large chapter in this book is dedicated to grease composition and properties for the various grease types (Chapter 3). A very important topic is bearing service life, which is given by the life of the grease (Chapter 4) and the life of the bearing (Chapter 13) supported by a separate chapter on reliability (Chapter 12).

Finally, separate chapters are dedicated to seal lubrication (Chapter 14), condition moni-toring (Chapter 15), test methods (Chapter 16) and lubrication systems (Chapter 17).

¹Lard was used for the lubrication of traditional windmills in the Netherlands.

2

Lubrication Mechanisms

2.1 Introduction

Compared to oil lubrication, the physics and chemistry of lubricating grease in a rolling bearing is today not well understood. Howevere, it is certain that grease provides the bearing with a lubricating film that is initially thick enough to (at least partly) separate the rolling elements from the raceways. Unfortunately, generally the thickness and/or the ‘lubricity’ of this film changes over time, leading to a limited period in which the grease is able to lubricate the bearing, generally denoted as ‘grease life’. This time is preferably much longer than the fatigue life of the bearing. It is still not fully understood how this film is generated or how it deteriorates over time and leads to bearing damage and ultimately failure.

Although an exact prediction of the film thickness and ‘lubricity’ cannot be made, it is certain that a number of aspects are very important in the prediction of the performance of the grease and/or in selecting the optimum grease for the specific bearing application. Examples are the rheology (flow properties of the grease), the bleeding characteristics, EHL oil film formation, boundary film formation, starvation, track replenishment, thermal aging (such as oxidation) and mechanical aging [374].

Another important aspect in grease lubrication in rolling bearings is that the ‘grease life’ is not deterministic, that is, there is no absolute value for this and it is given by a statistical distribution. Even if bearings are running under very well controlled conditions, such as in a laboratory situation, there is the usual significant spread of failures. The ‘grease’ life is therefore usually defined as L10, that is, the time at which 10% of a population of bearings is expected to have failed [280], similar to bearing life. If a higher reliability is required, a correction is needed. To prevent grease failures, a bearing may be relubricated. If possible, this should be done well before failure is to be expected. Generally, the relubrication interval is defined as L01, that is the time at which 1% of a population of bearings is expected to have failed [280].

All this, and more, will be treated in this book in separate chapters. To give the reader a summary and an introduction to these chapters, the possible mechanisms in combination with the physical aspects of grease lubrication will first be given in this chapter.

2.2 Definition of Grease

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’ [450]. The base oil is kept inside the thickener structure by a combination of Van der Waals and capillary forces [70]. Interactions between thickener molecules are dipole-dipole including hydrogen bonding [282] or ionic and Van der Waals forces [197]. The effectiveness of these forces depends on how these fibres contact each other. The thickener fibres vary in length from about 1–100 microns and have a length diameter ratio of 10–100, where this ratio has been correlated with the consistency of the grease for a given concentration of thickener [518]. Sometimes grease is called a thickened oil (rather than a thick oil) [226, 230]. Generally, a lubricating grease shows visco-elastic semi-plastic flow behaviour giving it a consistency such that it does not easily leak out of the bearing.

2.3 Operating Conditions

The lubrication process is different for different speeds and temperatures and even for different bearing types. At high temperatures, oxidation and loss of consistency play a major role. At very low temperatures, the high values for consistency and/or viscosity may lead to too high start-up friction torque. The temperature window at which a grease can operate is given by the grease manufacturer or by the bearing manufacturer and is determined by life- and functional tests.

At very low speeds, a bearing may be packed with grease because the churning losses will be minimal. This implies that there will always be sufficient lubricant in the inlets of the contacts and effects such as starvation may be neglected. At very high speeds the centrifugal forces on the grease inside the bearing will be so high that most of the grease will be lost from the contacts very quickly, leading to severe starvation or, in the case of sealed bearings, to an overfilled outer ring–rolling element contact. The definition of speed range is roughly as given in Table 2.1.

Table 2.1 Definition of speed ranges. Here n dm is the product of rotational speed (r/min) and bearing mean diameter (mm).

Table02-1

The lubrication mechanism that will be described below typically applies to bearings running at medium speed. Other conditions will be described elsewhere in this book.

2.4 The Phases in Grease Lubrication

There are roughly two phases for grease lubrication in bearings running under constant conditions, see Figure 2.1. After filling the bearing with grease and starting the rotation of the bearing, the grease will start flowing. As a rule of thumb, approximately 30% of the free volume of the bearing should be filled with grease. The quantity of grease that is available in this phase is therefore large enough to provide the bearing contacts with a fully flooded lubricant film.

Part of the grease flows next to the running tracks, where it will stay due to its consistency and part of the grease finds it way inside the bearing, such as under the cage bars or in the cage pocket. During this ‘churning phase’, the grease flow behaviour is governed by the internal design of the bearing, the design of the housing and the rheological properties of the grease. The friction torque will be large due to the relatively high ‘viscosity’ of the grease and the temperature of the bearing will rise. As more and more grease flows out of the swept volume of the bearing, the friction torque will decrease and so will the temperature, until a quasi steady state temperature has been reached. The ‘churning phase’ typically takes from a few hours to up to 24 hours, depending on the percentage filling and the speed. Examples will be given in Chapter 11. During most of this phase, the contacts will be ‘fully flooded’ with grease and the film will consist of grease materials, that is, both thickener material and oil. Typical shear rates in the contact are in the order of 10⁷ s−1 and 10⁴ s−1 in the cage pocket. The grease is therefore severely ‘worked’ and the fraction of grease that participates for longer times in the flow process will degrade heavily. The grease behaviour during this phase is determined by the rheological properties of the grease. Relatively fresh grease will be located on the cage bars or on the seals/shields. Heavily degraded grease can be found on the running tracks [111]. The film thickness during this phase may change rapidly as a function of the change of the rheological properties of the severely degraded grease on the tracks. Such a change in film thickness was measured by Wilson in 1979 [616] in cylindrical and spherical roller bearings, where he showed that the lubricant film initially exceeds the value that could be expected based on fully flooded base oil lubrication calculation. In his measurements, the film thickness decreased below this value almost instantaneously. The thick lubricant films at the beginning indicate that, at least during the initial bearing operation, thickener material enters the contact.

These film thickness measurements were made by measuring the electrical capacitance of the gap between rolling elements and raceways (see also e.g. Heemskerk et al. [253], Baly et al. [58] and Schrader [519]). This is a rather complex technique where all bearing contacts are measured simultaneously and where only relatively thick films can be measured. Therefore, single contact measurements are often made where a single ball runs on an optically coated glass disc and the film thickness is measured using interferometry techniques. This makes it possible to measure very thin films down to a few nanometers. Such single contact measurements have been made by Åström et al. [35], Williamson et al. [614] and Kaneta et al. [309], using a scoop for the grease to provide fully flooded conditions. They have confirmed with these techniques that the film thickness is higher than the fully flooded base oil film thickness. The optical set-up also made it possible to show that grease thickener lumps were entering the contact.

It is only in the initial churning phase that a fully flooded situation exists. Side flow, both in the inlet of a contact and in the Hertzian contact itself, will reduce the volume of lubricant on the tracks and starvation will occur. This can also be seen in the measurements of Wilson [616], later confirmed by Barz [68], who measured film thickness in cylindrical roller thrust bearings for longer times. The films become so thin that metal-to-metal contact occurs very regularly, which was shown by Wikström and Jacobson [613] who measured the electrical capacity in a grease lubricated spherical roller bearing.

2.5 Film Thickness During the Bleeding Phase

During the bleeding phase, there are several possible mechanisms for maintaining a lubricant film. The grease may release oil by bleeding [97] or by breakdown of the thickener structure in the contacts. It may also simply provide a stiff ‘grease film’ as referred to by Scarlett [518], who called this a ‘high viscosity layer retained within the rolling track’. It is very difficult to investigate this since so little oil is necessary for lubrication. For instance, Booser [93] operated a ball bearing on only two initial drops of oil at 36 000 r/min for two weeks at 100 °C before encountering failures! So if ‘grease layers’ are found in experiments, these layers are not necessarily the lubricant layers. Very small additional quantities of (bled) oil may provide sufficient lubrication for relatively long times.

By means of FTIR (Fourier Transform Infra-Red) spectroscopy, Cann et al. [116, 117] observed thickener layers on the surfaces of a ball-on-disc machine and assumed that the film was formed by base oil, thickened with broken thickener fibres. This may well be caused by mechanical work on the grease, which is heavily sheared in the highly loaded thin film contacts, causing breakdown of the thickener structure, adhesion to the surfaces [113] and release of oil, providing free oil for replenishment.

There is a clear consensus in the lubrication and bearing industry that the bleeding properties of a lubricating grease are important. For instance Kühl [347] found that roller bearings need greases with higher bleeding rates than ball bearings. Also, the work of Azuma et al. [39] and Saita [510] confirms that the grease bleeding properties have a direct impact on grease life. It is likely that both effects (bleeding and breakdown of the grease structure) play a role in providing the contacts with the lubricant, where the dominating mechanism depends on the operating conditions and/or bearing design.

2.5.1 Ball Bearings

Cann et al. [111, 112] have investigated the chemical composition of the lubricant in grease lubricated ball bearings taken from R0F (6204-type ball bearings, bore diameter diameter 20 mm)¹ and R2F tests (6209-type ball bearings, bore diameter 45 mm). In both cases, the operating temperature was the same. The small bearings (with a steel cage) were run at n dm = 335 000 mm rev/min and 670 000 mm rev/min, C/P = 65 and the larger bearings (with a polymer cage) at n dm = 97 500 mm rev/min, C/P = 3 and C/P = 10. The observed differences in lubrication conditions were not only related to bearing size, but more likely caused by differences in the operating speeds and load. For the lower speed, higher load R2F test, they write that initially grease is overrolled, 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 back 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. Oil was found in the cage pockets.

In the higher speed and lower load R0F test, no significant amounts of free oil could be found. This means that under these conditions, grease is sheared into the contacts and into the cage pockets where it is overrolled and sheared and where oil is released. Hence, contrary to the lower speed, high load R2F test, the grease on the shields may not serve as an oil reservoir.

Scarlett [518] described the flow of grease in a ball bearing (1 3/4 inch bore, n dm = 176 000 mm rev/min) with an inner-ring guided brass machined cage and mentions the formation of ‘pads’ of grease adhering to the cage bore (‘under’ the cage). These grease pads had a higher consistency with a higher soap concentration than the original grease. Scarlet explicitly states that this is due to oil loss by bleeding, which occurs during the first 100 hours of 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 paper though. In his tests he found heavily degraded thickener on the tracks but not next to the tracks. Scarlett describes tests where grease was removed from the covers after the initial churning period. In this case he found early 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 carried out his tests with various grease types! A similar conclusion was reached by Milne et al. [423]. Scarlett concluded 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. These results are contrary to what was found by Saita [510] who put tracer material in grease next to the bearing and measured tracer flow from this grease reservoir next to the bearing towards the running track, from which he concluded that this grease provides oil by bleeding!

Lansdown and Gupta [354] write that there is clear evidence that in ball bearings the whole of the grease is involved in the lubrication process, not just the bled base oil. They found equal performance in grease plated ball bearings (a technique where grease is coated on to the bearing surfaces) and in the case of conventional grease lubrication. In their analysis of ball bearings they also write that grease adjacent to the raceways is often softer and has a higher oil content then the grease near the outside of the bearing covers. Unfortunately, these statements are not illustrated with any examples, proofs or references.

Contrary to this, Dalmaz and Nantua [155] indicate again that the base oil provides the film. They tested six lithium greases in angular contact ball bearings, varying base oil viscosity and thickener structure and concentration. In addition, they performed single-contact film thickness measurements. Similarly to Hurley [282] they report that the initial film thickness is proportional to the thickener concentration and larger than that 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 after that the film will be formed by the base oil only for the main part of the life of the bearing. This mechanism is confirmed by Saita [510] who developed a new grease for deep groove ball bearings and cylindrical roller bearings in a traction motor for a high speed train in Japan, where additional grease reservoirs were made next to the bearing. Their measurements indicate a flow of base oil from grease (called oil bleeding) adjacent to the track, feeding the contacts.

2.5.2 Roller Bearings

It seems that greater consensus exists on the lubrication mechanisms in roller bearings. Here clearly oil bleeding is considered to be a main mechanism providing the lubricant for the rolling contacts (e.g. Booster and Wilcock [97]). This was also clearly found in the spherical roller bearings test from Wilström and Höglund [611, 612] where both grease and base oil were used and where equal friction torque was measured. Mas and Magnin [402] investigated grease before and after running in 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. However, they also show by means of SEM the destruction of fibres in the raceway, confirming once more Cann et al. [116, 117]’s conclusions, that is that the grease film consists of base oil thickened with broken grease fibres.

2.6 Feed and Loss Mechanisms During the Bleeding Phase

According to Wikström and Jacobsson [613], the film thickness during the ‘bleeding phase’ is determined by a mass balance of oil feed and loss to the contacts. Such a balance is schematically drawn in Figure 2.2. In their paper they assume a lubricant feed by oil bleeding, by shear, centrifugal forces and capillary forces. The feed by shear takes place through, for example, the cage shearing action on the volume of grease located on the bearing shoulders or seals (see e.g. Cann et al. [111, 112]). Inside the EHL contact, the contact pressure will drive lubricant out from the contact. The fraction that is driven out in the running direction can be used to lubricate the following contacts and is therefore not lost. The fraction that flows out of the track, however, does not easily flow back and may be considered as lost. Oxidation and polymerization may not necessarily be considered as part of the ‘loss mechanisms’. These processes may also change the properties of the lubricant and have an indirect effect on film formation. Evaporation will be relevant in the case of air going through the bearing. In some bearing types the centrifugal forces on lubricant layers may be so high that these forces induce flow either towards or away from the contact. Capillary forces, surface tension driven forces or ‘Marangoni’ effects (thin layer flow due to temperature gradients), may replenish the contact [327]. This may be especially relevant in the case of low starvation or occasional starts and stops. Finally, the cage may be scraping off or redistributing the lubricant on the tracks [157]. The above described mechanism does not always apply. In the case of slow rotation, outer-ring rotation, large bearings, vibrations, shock loads and so on other mechanisms also play a role.

Figure 2.1 The phases in grease lubrication of rolling bearings.

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Figure 2.2 Balance between feed and loss of lubricant ultimately determining the lubricant film thickness.

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2.7 Film Thickness and Starvation (Side Flow)

During the ‘churning phase’ and at the beginning of the ‘bleeding phase’ the bearing contacts will be fully flooded with grease where the initial film thickness is higher than can be expected based on the base oil viscosity alone. Both base oil and thickener material will be dragged into the gap between rolling element and ring raceway. The fully flooded film thickness has been modelled by Dalmaz and Nantua [155] and Hurley [282] by assuming that the initial film thickness is proportional to the thickener concentration and larger than that of the base oil. Others have used grease rheology as input for a model. Jonkisz and Krzemiński-Fredihave [306] and Kauzlarich and Greenwood [315] used a Herschel–Bulkley model. Bordenet et al. [99] 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 alone. Yang and Qian [625] 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 [21] performed an experimental study of the factors affecting film thickness in a grease lubricated two-disc machine. They suggest that the grease lubricated starved 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 with Saman’s [513] theory, who assumed that the contacts will ultimately be so starved that the inlet meniscus will move close to the Hertzian contact, such that zero-reverse flow can be assumed. Theoretically this will lead to a reduction of 71% of the fully flooded film thickness.

The reduction in film thickness after the initial phase may not only arise from classical starvation. Kauzlarich and Greenwood [315] show that shear degradation of the grease also leads to a reduction of film thickness over time and that the fully flooded film thickness in grease lubricated bearings after some time can simply be calculated using the base oil viscosity.

During the ‘bleeding phase’, the observed decrease of film thickness over time is primarily caused by side flow of base oil from out of the Hertzian contacts. If the film thickness is assumed to consist mainly of base oil, starvation models for oil lubricated contacts are relevant. Such models have been developed by, for example, Chevalier et al. [124], Damiens [156, 158] and Van Zoelen et al. [584, 586]. The oil will be driven out of the running track by flow in front of the contacts and inside the Hertzian contacts. The first effect will be relevant at the onset of starvation. Later, the oil layers on the running tracks will be so thin that the film thickness inside the contacts will be almost equal to the combined oil layers on the track, however reduced, due to the compression of the oil layers by the contact pressures with a maximum of approximately 30%. At this point side leakage is primarily caused by side flow from oil inside the contacts. However, this will be relatively small, due to the very high viscosities caused by high contact pressures and the piezo-viscous behaviour of lubricating oils.

At higher temperatures the thin lubricant layers feeding the starved lubricated contacts may be deteriorated by effects such as evaporation [339] and/or oxidation [494–496].

2.8 Track Replenishment

In the absence of track replenishment, the film thickness decreases very rapidly [586]. Replenishment of running tracks has been investigated since the early 1970s when Chiu [125] showed, using a viscous flow model, that replenishment of oil that is pushed to the side by the roller/ball–raceway contact can flow back into the track if a sufficiently thick layer is present next to the track and if the bearing speed is not too high. Surface tension driven replenishment is generally too slow to be relevant in bearings, but capillary forces may have some effect [298]. Even at moderate speeds, centrifugal forces may drive the flow in free oil layers in rolling bearings. Gershuni et al. [218] calculated the flow from ridges of oil next to the tracks on the inner-ring and the outer-ring in cylindrical roller bearings and showed that in the case of outer-ring rotation significant replenishment of the tracks takes place, whereas replenishment from these ridges will be very slow in the case of inner-ring rotation. Oil may actually be thrown off the rings! Farcas and Gafitanu [192] developed a model based on the wetting properties of the lubricant only for inner-ring rotation, they calculated the critical speed at which lubricant droplets are no longer able to adhere to the surface due to the centrifugal forces (in their tests at about n dm = 700 000 mm rev/min). They validated their model using electrical resistance measurements over the bearing contacts and showed that metal-to-metal contact occurs above a critical speed.

Another possible replenishment mechanism is described by Merieux et al. [417] who show that grease shear degradation in the vicinity of the contacts may cause softening of the grease until the grease has been transformed into plain base oil, with a enough quantity to replenish the running track and cause film growth. This softening was confirmed in the work of Landsdown and Gupta [354].

Van Zoelen et al. [583] 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. Such a flow may either replenish the track or shear oil further away from the contacts.

The cage also plays a very important role in film replenishment. The cage may store grease from which oil bleeding will take place. It will also direct the flow in the cage pocket. The cage may scrape off the lubricant from the running tracks but it may also redistribute the lubricant and thereby repair the lubricant layers which have become critically thin locally. This was shown by Damiens et al. [157] who made 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 were then able to vary the clearance between the cage and ball from 0.05–0.5 mm. They show that the behaviour with oil is very different from that with grease and that the clearance in the cage–ball contact is critical here.

Lubricant replenishment by oil bleeding plays an important role. This bleeding action may take place from grease which is stored on the cage and is heavily pressurized by centrifugal forces [39, 45] or from stationary grease stored next to the swept volume (e.g. [510]).

2.9 Grease Flow

After the initial filling of the system, the grease is forced to flow by the moving rolling elements and cage. Most of the grease is pushed sideways but part of it stays close to the contacts or ends up on the cage. For grease lubricated systems, the initial filling is of crucial importance. Too much grease will lead to excessive churning and therefore, because of the high consistency of grease, too high friction levels, which produce increased operating temperatures [612]. This will cause breakdown of the network structure and oxidation of oil and thickener leading to a short grease life and leakage out of the system. Too little grease reduces the efficiency of replenishing the running tracks and therefore also leads to a short lifetime [192, 332]. In addition to the amount of grease, the initial position of the grease in the bearing or gearbox before churning is also important [379]. Relatively small differences in initial filling may lead to large differences in performance. However, according to Cobb [129] there is no difference in grease performance with respect to start-up torque, temperature and 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 often most of the grease in a ball bearing participates in the initial flow phase.

The ultimate aim is to provide the bearing with a grease distribution that is optimal for the system performance – not too much, preventing the grease from continuously churning/ flowing, and not too little, ensuring an optimal supply of oil by bleeding or shear. The amount of grease that can be stored close to the running tracks obviously depends on the internal design of the bearing and the flow properties of the grease, that is its rheological behaviour. The temperature distribution in the bearing is also important here. Generally, for practical reasons, the bearing temperature is measured on the outer-ring and the cage temperature is hardly ever reported. Joshi et al. [307] have performed temperature measurements on the cage of a tapered roller bearing. The bearing was running in an oil bath (75% full). They recorded the temperature of both housing and cage and showed that the cage temperature response is much more sensitive to changes in lubrication than the housing temperature. This has an impact on the ‘fluidity’ of the grease and therefore again on the flow.

The operating conditions and also the design of the equipment, have an impact on the flow of grease. For instance, in a case of vertical shaft arrangements or where vibrations are present, the amount of grease available for lubrication will be different from ‘standard conditions’. Under such circumstances, generally a high consistency grease is used to prevent grease falling back into the track and to maintain a lubricant reservoir adjacent to the row(s) of rolling elements.

Until now the flow of grease in bearings has only been studied experimentally. Visualization techniques have revealed flow patterns [518] but most of the work has been indirect, relating flow to friction torque or temperature measurements. A quantitative model enabling prediction of the formation of the grease reservoir is not available today.

The flow of grease in a bearing is a two-phase system: a mixture of air and grease. The crucial free surface effect is missing in all studies that have been done on grease flow so far. Strictly, oil separation takes place as well, adding another phase to the system. In addition, thermal and mechanical aging takes place, continuously changing the mechanical properties of the grease. Another complicating factor in the study of grease flow in rolling bearings is the large 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 there may be phase separation (similar to what happens with water in emulsions in the inlet of EHL contacts), so that a jet flow may occur or even droplets formed [356].

2.9.1 Non-Newtonian Rheology

The flow behaviour, but also the volume of grease that can be maintained inside the bearing close to the tracks, is determined by the flow properties of the grease, that is the grease rheology. This topic has been studied quite extensively. For instance, the possible visco-plastic behaviour, that is the existence of a yield stress for grease, has been the topic of many papers [34, 61, 64, 311, 403, 630], where the main conclusion is that this behaviour may be assumed if high accuracy at low shear rates is not required. Actually, creep occurs and the grease has a very high viscosity at such low shear rates. The solid-like behaviour, 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 a NLGI consistency number. A correlation between yield stress and penetration/consistency can be found in Chapter 5 or, for example Couronné et al. [139]. Generally, this is only determined at room temperature, which makes it a good general stiffness classification number but which also makes it useless as a measure for the stiffness of the grease at the bearing operating temperature.

The grease will be severely worked in the bearing. This applies to the grease that is being churned between the rolling elements but also to the fraction that passes the EHL contacts where shear rates are O(10⁶ s−1). This causes a rapid change in the 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 [9] (Figure 2.1) or in a Shell roll stability tester [10] (Figure 2.2). The ability to maintain its consistency when worked is called ‘shear stability’ or ‘mechanical stability’. The yield stress depends strongly on temperature. Measurements for different types of grease can be found in Karis et al. [311] and Czarny [148]. For example Karis shows that the yield strength of a lithium grease may drop from 500 Pa at 20 °C to 100 Pa at 60 °C.

At higher shear rates, visco-elastic behaviour is observed (e.g. Forster and Kolfenbach [198]), often in combination with shear thinning. In general the well known non-Newtonian rheology models such as the Cross model, power law, Herschel–Bulkley or Sisko models can be used to describe the rheological behaviour of grease [630]. Measurements from low to high shear rates can be found in Pavlov and Vinogradov [468].

There are a number of models proposed for low and high shear rates. The best known are the power law, Rhee–Eyring, Bingham and Herschel–Bulkley models. A definition of these models can be found in Yousif [630]. These models assume solid or very high viscous behaviour at low shear rates and viscous behaviour (with possible shear thinning) at higher shear rates. A model that fits the measurements well in a wide range of shear stress is Palacios and Palacios’s [461]:

(2.1) numbered Display Equation

Here is the shear stress, the yield stress, K the grease consistency