General Bearing Design

Bearing Design & Lubrication Introduction The components of industrial machinery requiring grease lubrication include bearings, couplings, open gears and a variety of other moving parts. The widest use of grease is in lubricating bearings, which are critical elements in equipment used in steel mills, mining, construction and transportation — the industries largely determining the nation’s economic stability. A bearing is a housing or support for a rotating part (a shaft rotating within a bearing) or one that moves linearly (moves axially within the bearing). A bearing may also restrain motion in some manner. Bearings are of two basic types: plain and rolling-element. Plain bearings are based on sliding motion between a stationary and a moving member. Rolling-element bearings have either balls or rollers that accommodate motion between the stationary and moving parts. In either case, a film of lubricant separating moving surfaces is essential for long service life. Plain bearings that support loads perpendicular to their axis of rotation are called journal bearings; rollingelement bearings carrying similar loads are called radial bearings. Bearings of either type that support loads parallel to their axis of rotation are called thrust bearings.

Operating Ranges for Different Types of Bearings

There are many different types of anti-friction bearings but when all the factors relating to load spectrum and capacity, size, performance, costs, bearing life and reliability are analyzed, designers in most industries rarely deny the logic and value-analysis benefits of the tapered roller bearing. Of course, other bearing types have their accepted fields of use as discussed in the following review, but the tapered roller bearing is the most versatile of all bearing types.

Fluid Film Bearings The most widely used plain bearing. They rely on lubricant viscosity to separate the bearing surfaces. The rotating shaft drags the lubricant around forming a supporting wedge. These bearings usually need a lubrication system. The continuous lubrication acts to cool the bearing allowing high shaft speeds and heavy loads.

Rolling Element Bearings Rolling element bearings generally consist of two rings (races or raceways) with a set of rolling elements running in their tracks. The rolling elements take the form of balls or various types of rollers. The outer race is located in a housing and the inner race on the shaft. Often one of the races is not fixed axially but is free to move to allow for shaft movement. Due to the inertia forces on the bearing shaft speeds are limited but they have an advantage over fluid film bearings as a lubrication system is generally not needed.

Porous Metal Bearings These bearings are made from a porous metal, typically sintered bronze, impregnated with lubricant. They are cheap and simple and quite sufficient for many applications. Applications include shafts for light power transmission, typically electric motors and small engines.

Dry rubbing bearings These are usually plain plastic bushes which have to be able to run with marginal or no lubrication. The main requirement of dry bearings is that the bearing surfaces have a low coefficient of friction. These bearings are used in undemanding applications such as low speed moving parts in domestic appliances, instrument and electro-mechanical devices. Bearings may be self-lubricated or externally lubricated with oil or grease. Grease is generally preferred under high loads or when good adhesion to bearing surfaces or good sealing properties are necessary. Fluid Film

Plain Bearings A plain bearing is the most basic type and contains no moving parts . In their simplest form, these bearings consist of load-carrying cylindrical inserts, made of a material or alloy that is softer than the part that slides or moves against it. Consequently, the bearing assumes most of the wear. This is an important economic advantage because bearings are more conveniently replaced or adjusted than the relatively inaccessible moving components . Plain bearings are used mainly in applications where the loads are relatively light and the motion is relatively continuous. The prime example is crankshafts and connecting rods in internal combustion engines. To work efficiently, there must be some clearance between the stationary part and the rotating part. Extreme care needs to be taken to ensure that particle contaminants do not collect in this clearance in order that rapid wear is avoided. Regular maintenance is therefore very important with most plain bearing applications. In view of the need for frequent maintenance anti-friction bearings with rolling elements have largely replaced plain bearings in modern applications. Plain bearings can be described by their configuration, by their motion, or by the type of loading they accommodate. Thus, the major categories of plain bearings are journal, guide and thrust bearings.

Journal Bearings A journal or sleeve bearing consists of a cylindrical housing supporting a rotating shaft. The term “journal” refers to the portion of a shaft contained within a bearing; “sleeve” refers to the bearing configuration. The terms are commonly used interchangeably. If the bearing is a full-cylindrical, 360° design it is called a bushing. A shaft that is loaded in a single direction can be supported by a journal bearing in the form of a partial cylinder. Such a bearing supports the shaft in the load zone only. For example, cranes, earth-moving equipment and railroad journals use partial-cylinder bearings to support loads directed against the top portion of an axle.

Configurations of plain journal bearings Journal bearings frequently contain two or more parts to facilitate removal or replacement. Automotive engine main bearings, for instance, contain two half-round sleeves which contain the crankshaft journals. A single-piece bearing may be difficult to remove because of flanges on the shaft, which may be too large to be pulled through the bearing housing, or perhaps the shaft is simply too long or too heavy to be easily freed from the bearing.

Guide Bearings Guide bearings support reciprocating rather than rotating machine parts; loading is generally less than in a journal bearing. Interior surfaces may be grooved to help distribute lubricant and relieve pressure. Equipment using guide bearings include crossheads in steam engines and some air compressors. Plain guide bearing accommodates linear motion

Thrust Bearings Thrust bearings accommodate the axial movement of a rotating They are usually used in conjunction with journal bearings and are lubricated by grease, which leaks from the ends of the journal housings.

shaft.

Rolling-Element Bearings The main function of any rolling Bearing type is the decreasing of friction which could be caused from any elements related rotation, in order to reach the maximum efficiency of the power, which could have been converted from torque power into heat power, so the main function for the bearings is decreasing the "converted to heat" power. A great variety of rolling bearing types and designs are known, the manifolds of the bearings is justified by their various purpose of application, each type of bearing has a certain characteristic technical properties, rolling bearings have the following advantages compared with the plain bearings: • • • • • • • • •

Low starting moment. Low friction at all speeds. Low energy consumption. High reliability. Small width. Low consumption of lubricant. Long relubrication intervals. Easy to mount and dismount. Standardized dimensions.

A rolling-element bearing has balls or rollers situated between a stationary housing and a moving journal. Rollers can be cylindrical, spherical or tapered. Roller bearings with relatively long, small-diameter rollers are called “needle bearings.” Rolling element bearings generally consist of two rings (races or raceways) with a set of rolling elements running in their tracks. The rolling elements take the form of balls or various types of rollers. The outer race is located in a housing and the inner race on the shaft. Often one of the races is not fixed axially but is free to move to allow for shaft movement. Due to the inertia forces on the bearing shaft speeds are limited but they have an advantage over fluid film bearings as a lubrication system is generally not needed. Typically, the outer race is stationary and the inner race is affixed to a rotating shaft. Unlike plain bearings, rolling-element bearings are made of hard steel alloys because the small rolling elements must carry a wide range of loadings and unit stresses at the contact surfaces can be very high. Most rolling-element bearings used in industry are grease lubricated. Rolling-element bearings are frequently called “antifriction” bearings although the frictional torque of a full fluid-film plain bearing can be as low as that of a rolling-element bearing. However, starting friction in a plain bearing is usually higher than that of a rolling-element bearing. The following are some of the very common types of bearings, a certain bearing type is chosen in order to make it particularly suitable for the desired applications, however it is not possible to apply general application rules for the selection of any type of bearings, as several factors must be considered and assessed in relation to each other they are all produced by most of the bearings manufacturers:

Deep Groove ball bearing

Cylindrical Roller Bearing

Taper Roller Bearings

Spherical roller and Needles bearings

Basic types of rolling-element bearings

Ball Bearings Ball bearings are, perhaps, the most familiar type of rolling-element bearing. Radial ball bearings contain the rotating motion of a shaft and are functionally similar to plain journal bearings. Ball thrust bearings are the functional equivalent of plain thrust bearings. All types of ball bearings have a ‘point’ contact between the balls and the races. Because of this point contact for a given load capacity it is generally necessary to specify a larger ball bearing than a tapered roller bearing which distributes the load over the length of the roller (line contact). The most popular type is the ‘deep groove ball bearing’ which is suitable for light radial loads only. Axial or shock load conditions should be avoided since this can lead to rapid failure. It is also important to ensure that the lubricant is ‘clean’ since contaminants tend to get trapped in the race groove, which ultimately limit performance and bearing life.

A ball-bearing assembly includes balls, a retainer, races, the rotating shaft and the supporting housing. The balls are made of hardened steel, ground to a true sphere, and polished to a fine finish. The balls are held in position by a retainer, or spacers, and roll between races which must also be ground and polished to a fine finish. This type of bearings divided into two categories Deep groove ball bearings are supplied as an assembly with inseparable races; they necessarily have a radial clearance which can only be adjusted by the fits of the inner and outer races. They are mostly used where size, load capacity and radial clearance are not important, but where ease of assembly and low cost are.

1- Single row deep groove ball bearings Single row deep groove ball bearings are: simple design; carry considerable axial loads in either direction, even at high speeds; and little attention in service. Shielded deep groove bearings are intended to be used in any applications where the inner ring is to rotate. The steel shield forms a small gap with the inner ring shoulder, oil and rubber wearresistant synthetic rubber, seal against a recess in the inner ring. Single row deep groove ball bearings have limited ability to accept any misalignment. Single row deep groove ball bearings are the most popular of all the ball bearings in use today, because they are various types and are cost effective relative to performance.

Radial ball bearings Radial ball bearings use a versatile design that permits relatively high-speed operation under a range of load conditions. Bearings consist of an inner and outer ring with a cage containing a complement of precision balls. The standard Conrad-type bearing has a deep-groove construction capable of handling radial and axial loads from either direction. The maximum-capacity type supports primarily radial loading. An extremely wide variety of sizes is available in extra-light to heavy series. Various shield and seal configurations help protect internal bearing components and retain lubricants.

Angular contact ball bearings The angular contact ball bearing is designed to overcome some of the limitations of the deep groove ball bearing, by improving its ability to cope with the combined thrust loads as well as radial loads. Although load capacity is increased compared to the deep groove ball bearing, it is significantly less than the equivalent dimensioned tapered roller bearing (see Fig.). Single row angular contact ball bearings

are of non-separable design, usually arranged so that they can be adjusted against a second bearing. They are not very tolerant of any misalignment between shaft and housing and this can have a serious effect on bearing life. Single-row bearings have high thrust capacity in one direction. Some single-row bearings are specifically designed for duplex mounting in sets for maximum performance. Double-row Conrad bearings can accommodate thrust in both directions. The special geometry of angular contact bearing raceways and shoulders creates ball contact angles that support higher axial loads. Different designs offer contact angles ranging from 20° to 40°. Higher angles provide more axial load capacity for longer service life operating under axial and radial loads.

2- Double Row Angular Contact Ball Bearings Double Row Angular Contact Ball Bearings correspond in function to two single row bearings arranged back-to-back. These bearings can accommodate axial loads in both direction, as well as, tilting moments. Bearings with one-piece inner ring have a filling slot at one side. If directional axial loads must be accommodated, the bearings should be arranged so that the axial load acting on the shaft is not directed towards the filling slot. Angular misalignment can only be accommodated between the ball bearing and the raceways by force, however, this produces increased ball bearing loads and may lead to a reduction of bearing life

Roller Bearings A roller-bearing assembly consists of rollers, a retainer, races, a shaft and a bearing housing and seals. As with ball bearings, the contact surfaces of roller bearings must have a fine surface finish that perform with maximum efficiency.

Cylindrical roller bearings The most common use of cylindrical roller bearings is at non-locating positions where it is necessary to accommodate thermal expansion effects by allowing axial displacement (floating) of the shaft relative to the housing. Inner and outer races are ‘separable’ (which facilitates mounting and dismounting), and the cylindrical rollers have line contact with the races so they can carry more radial load than the point contact of ball bearings. The cage must align and retain the rollers which necessitates a heavier cage and fewer rollers when compared to a tapered roller bearing of similar size, and hence a lower load capacity. The rollers are not true cylinders but are usually crowned or end relieved to reduce stress concentrations at the ends of the roller-race contact. Designs are also available to carry a limited axial load by integrating additional flanges on the inner or outer race. However, the full axial thrust must be taken between the flanges and the roller ends which can lead to high stress concentrations . Cylindrical roller bearings can carry heavy radial loads and operate at high speeds.

Tapered roller bearings Because of their geometry and design features, tapered roller bearings provide several important and unique performance characteristics enabling them to meet a wide range of application requirements. Tapered roller bearings consist of four basic components. These are the inner race (cone), the outer race (cup), tapered rollers and a cage (roller retainer) (Fig. 1). Under normal operating conditions the inner race, outer race and rollers carry the load while the cage spaces and retains the rollers. The inner race, roller and cage is referred to as the 'inner race assembly' and this is usually separable from the outer race, facilitating equipment build.

Fig 1 Tapered Roller Bearing , the tapered roller bearing can handle both radial and thrust loads The tapered roller bearing combines the benefits of all the other bearing types as well as offering additional advantages : • • • • • • • • • • • • •

Because of its tapered roller/race geometry, a tapered roller bearing can carry both heavy radial and thrust loads or heavy combined loads than angular contact ball bearings(Fig. 1) Excellent load carrying capacity/cross section ratios provide economic bearing arrangements. A longer relative life for a given bearing size Reduced bearing size for a given load capacity. high reliability and low operating temperatures Adjustability : for optimum performance either endplay or preload values can be specified according to the design requirements. Tapered roller bearing bearings are less sensitive to misalignment and offer long life Less sensitive to contaminated environments due to the natural pumping action which forces any particle contaminants out of roller/race contact area. Low friction coefficient and high speed capabilities due to true rolling motion, when compared to other roller bearing types. Simple mounting and dismounting with separable inner and outer races. A choice of mounting arrangements to suit the loading aspects and design constraints of an application. Lower price for a given calculated fatigue life This adds up to optimum bearing system performance for virtually any application, together with cost effective design and manufacture of equipment, as well as reduced maintenance and longer life.

Axial loads are only supported in one direction. Account must be taken of the fact that due to the angle of the rollers radial loads will generate axial forces. To counteract this bearings are usually employed in pairs, either face to face or back to back. The bearings can then be adjusted against each other to provide preload which determines the internal clearance. This versatile bearing is especially popular in the automotive industry.

Spherical roller bearings The self aligning feature of spherical roller bearings allows minor angular displacements between shaft and housing to be accommodated. They have a high radial load carrying capacity, but under heavy load the stress is not evenly distributed and true rolling motion only occurs at two contact points on each roller. This naturally induces skidding along the roller length and therefore the spherical roller bearing has a higher coefficient of friction and lower speed capabilities than other types of roller bearings. This adds up to optimum bearing system performance for virtually any application, together with cost effective design and manufacture of equipment, as well as reduced maintenance and longer life. The cage must be of an extremely robust construction to counteract the roller skewing effect which increases the cage moment of inertia and limits the number of rollers. It is best suited to applications where there is a risk of misalignment at assembly, and where speed and deflection criteria are not exacting. Mounted on adapter or withdrawal sleeves and housed in plummer blocks they present economic bearing arrangements. Also available with seals for maintenance-free operation. . However, they are somewhat speed limited.

Needle roller bearings Needle roller bearings are similar to cylindrical roller bearings but with long, thin rollers, giving them a very compact cross-section. Needle bearings, containing cylindrical rollers with a high length-to-diameter ratio, provide the highest load capacity for a given radial space of any rolling-element bearing. They are very adaptable and have a high radial load capacity in relation to their sectional height, but can cope only with very light axial loads. This type frequently has no inner race and can accommodate oscillating motion. Needle bearings without a retainer, or cage, and a full-complement of rollers provide high load capacity, but are speed limited. Needle bearings with a retainer contain fewer rollers and, therefore, have a lower load capacity, but can operate at higher speeds. Wide variety of designs including combined bearings for radial and axial loads providing simple, compact and economic bearing arrangements. Typical applications are in the synchromesh mechanisms of automotive gearboxes, and as planetary gear bearings in light duty epicyclic hub-reduction units.

Lubrication of Bearings Proper lubrication is essential to successful performance of any bearing and necessarily includes the selection of an adequate type of lubricant, the right amount of lubricant and the correct application of the lubricant on the bearing. The fundamental functions of a lubricant are as follows : • • • •

To prevent metallic contact between the rolling elements, raceways, and cage To prevent the bearing from corrosion and wear To separate mating surfaces and reduce friction. To transfer heat (with oil lubrication).

•

To reduce rolling and sliding friction.

•

To protect the highly finished surfaces of the rolling elements and races against corrosion and pitting. This is extremely critical to service life.

•

Sealing and cooling of the bearing.

•

To act as a sealant.

These functions include consideration of the lubrication and generated film thickness on the raceway (simulated according to elastohydrodynamic effects) and on rib/roller end contact. Bearings may be lubricated with grease or oil-in special cases with a solid lubricant-the choice of which depends primarily on the temperature range, operating speeds, and loading conditions of the bearings concerned. All lubricants must be changed from time to time because their properties deteriorate as a result of aging and contamination of the lubricants. The limited speeds for both grease and oil lubrication are usually given in the bearing tables.

Grease Lubrication Grease lubrication is generally used for bearings operating under normal conditions. Grease has certain advantages over oil because it's easily retained in housing as well as its sealing effect against the entry of moisture and outside impurities. In general, the bearing should only be 1/3 to 1/2 partly filled with grease because over filling will cause rapid temperature rise, particularly if speeds are high.

Oil Lubrication Oil lubrication is generally used when high speeds or operating temperatures are beyond the effective range of greases, or when it is so designed that heat developed in the bearing must be transferred away through a lubricating oil circulation system. Viscosity of the lubricating oil is the main factor to be considered to suit the diverse applications with respect to speed, temperature, and loading conditions. Oil lubrication is effective; however, its oil feeding and sealing devices must be provided. Some of them are recommended as follows:

Lubrication of Plain Bearings The mode of lubrication of a plain bearing depends on the conditions that affect the bearing’s ability to develop a load-carrying fluid film to separate the journal and bearing surfaces. If such a film is not produced (or before it is produced), the lubrication mode is termed boundary or mixed-film — lubrication where the surfaces are not completely separated and some metal-to-metal contact occurs. If a lubricating film is formed with sufficient pressure to separate the journal and bearing surfaces, the lubrication mode is termed hydrodynamic or full fluid-film lubrication.

Hydrodynamic Lubrication Keeping a liquid film intact between surfaces moving with respect to each other is generally done mechanically, as by pumping. In a cylindrical journal and bearing, the rotary shaft acts as a pump to maintain the lubricant film. The journal floats on a film of oil with an equilibrium thickness established between oil input and oil leakage (mostly at the bearing ends). The equilibrium thickness of the oil film can be altered by: • • • •

Increasing load, which squeezes out oil. Increasing temperature, causing more oil leakage. Changing to a lower viscosity oil, which also causes more oil leakage. Reducing journal speed, which generates a thinner oil film.

Lubrication of a journal rotating in a cylindrical bearing offers the classic example of the hydrodynamic theory of bearing friction, as described by Osborne Reynolds in 1886. The theory assumes that under these conditions, friction occurs only within the fluid film, and is a function of fluid viscosity. As speed increases, the wedging action of the lubricant moves in the direction of rotation, and pressure within the film becomes greater so the journal is now riding on a full fluid-film and hydrodynamic lubrication is reached. If loading on the bearings is increased sufficiently, the hydrodynamic film may collapse and the bearing will revert to the boundary-lubrication mode.

Elastohydrodynamic Lubrication As pressure or load increases, viscosity of the oil also increases. As the lubricant is carried into the convergent zone approaching the contact area, the two surfaces deform elastically due to lubricant pressure. In the contact zone, the hydrodynamic pressure developed in the lubricant causes a further increases in viscosity that is sufficient to separate the surfaces at the leading edge of the contact area. Because of this high viscosity and the short time required to carry the lubricant through the contact area, the lubricant cannot escape, and the surfaces will remain separated. Load has little effect on film thickness because at the pressures involved, the oil film is actually more rigid than the metal surfaces. Therefore, the main effect of a load increase is to deform the metal surfaces and increase the contact area, rather than decrease the film thickness. The principles governing the lubrication of rolling-element bearings differ from those of plain bearings. In a full fluid-film plain bearing, the journal load is supported by a continuous hydrodynamic lubricant film which keeps the two contact surfaces separated. In a rolling-element bearing, unit pressures are extremely high between the relatively small rolling elements and their raceways. Lubricants subjected to the high pressure within the contact zone of a rolling-element bearing undergo a dramatic increase in viscosity. This enables the lubricant film to withstand the high contact stresses while preventing contact between the rolling surfaces. Pressures of this magnitude do not exist in a full fluid-film plain bearing and lubricant viscosity is unaffected.

The high contact pressures in a rolling-element bearing also elastically deform the rolling surfaces to enlarge the contact area which supports the load. The combination of surface deformation and hydrodynamic lubricating action produces a thin, elastohydrodynamic (EHD) lubricant film which provides for lubrication within the contact zone of rolling-element bearings. The two major considerations in EHD lubrication are : • •

The elastic deformation of the contacting bodies under load. The hydrodynamic effects forcing the lubricant film to separate the contacting surfaces while the pressure is deforming them.

Elastohydrodynamic (EHD) lubrication

Film thickness on the raceway The importance of the elastohydrodynamic lubrication mechanism lies in the fact that the lubricant film thickness between the two contacts can be related to the bearing performance. The thickness of the generated film depends on the operating conditions such as : • •

Velocity Lubricant viscosity

▪ Loads ▪ Pressure/viscosity relationship.

Analytical relationships for calculating the minimum and the average film thickness have been developed : Minimum film thickness (based on Dowson Equation) : hmin = KD (µo V)0.7 a0.54 W-0.13R0.43 where : hmin = minimum lubricant film thickness µo = lubricant viscosity at atmospheric pressure a = lubricant pressure viscosity coefficient

KD = constant containing moduli of elasticity V = relative surface velocity W = load per unit length

R = equivalent radius Average film thickness (based on Grubin Equation) : h = 0.039 (µVa)0.728 (P/ l) -0.091 (S 1/ R) 0.364 where : h = lubricant film thickness (mm) V = surface velocity P = load between inner race and rollers S 1/ R = sum of inverses of contact radii

µ = viscosity of lubricant a = lubricant pressure viscosity coefficient l = effective length contact between rollers and inner race

The major factors influencing the lubricant film thickness are viscosity and speed whereas load has less importance. These thin EHD films are often not much larger than the surface roughness height.

The fatigue life of a bearing is related in a complex way to speed, load, lubrication, temperature, setting and alignment. The lubricant’s role in this interaction is determined primarily by speed, viscosity and temperature; and the effects of these factors on bearing fatigue life can be dramatic. For example, in a test program, table 5-A, two bearing test groups were subjected to conditions of constant speed and load. Differing film thicknesses were achieved by varying operating temperature and oil grade, and thereby, oil operating viscosity. Life was dramatically reduced at higher temperatures, with lower viscosity, and thinner resultant films

Film thickness at rib/roller end contact The contact between the large end of the roller and the inner race rib is described as an elastohydrodynamic contact or a hydrodynamic contact (elastic deformations are negligible). As the roller/rib loads are much lower than the roller/race loads, the film thickness at the rib/roller end contact is usually larger than the film thickness on the roller/race contact (approximately 2 times more). Nevertheless, in severe conditions, scoring and/or welding of the rib/roller asperities can occur. This may be related to speed, oil viscosity, load or inadequate oil supply to the rib/roller end contact. In these conditions, the use of Extreme Pressure (EP) lubricant additives may help prevent scoring damage. Grease lubrication: Greases provide a lubricating film on the surfaces of rolling elements, separators, and raceways. The lubricant is actually a thin film of oil which is released as the three-dimensional fibrous network within the grease is ruptured under shear. Only that portion of the grease in intimate contact with moving surfaces breaks down; the balance remains intact and functions as a sealant. When a freshly charged bearing begins to rotate, grease is thrown from the rolling elements and is rapidly circulated through the housing. After a short time, grease from the rotating outer race is then thrown back onto the rolling elements where shearing takes place. This turbulent environment at the outset of rotation creates frictional heat which attains a maximum and then gradually diminishes as the continual shearing action releases a lubricating oil film. As lubrication takes effect, the temperature of a properly charged bearing will drop and assume equilibrium.

WARNING: Bearing lubrication is critical. Failure to maintain proper lubrication can result in equipment failure, creating a risk of serious bodily harm.

Boundary Lubrication The simple assumptions made in discussing fluid film lubrication are hardly ever valid in practice. Under certain conditions -- such as shock loading, steady heavy load, high temperature, slow speed, and critically low viscosity -- the lubricant system no longer remains in the hydrodynamic regime. A situation arises wherein there is intermittent contact between the surfaces, resulting in a significant rise in temperature and subsequent destruction of the contacting surfaces. Under these circumstances, the fluid film is no longer capable of adequately protecting the surfaces, and other approaches must be employed such as adding filmforming additives. Because of the generation of relatively high levels of friction and heat, and the resulting high rate of surface wear, boundary lubrication is not the most desirable mode of operation. However, at times, it is completely unavoidable. Grease should be introduced to the bearing where fluid pressure is least — at the point of maximum clearance within the bearing. Grooves are often incorporated in the interior surface of a journal bearing to relieve pressure and to store reserve lubricant. When loading is in one direction, axial grooves running lengthwise on the bearing surface and located in areas of low pressure will not disturb the lubricating film and can relieve pressure. When the direction of loading is variable, the location of pressure extremes within the bearing is also variable. Under these conditions, well spaced annular or circumferential grooves will relieve pressure without substantially interrupting lubricating films. Axial grooves should be beveled so lubricating grease is more easily swept from the groove by the rotating shaft.

Lubrication of Rolling-Element Bearings The lubricant in a rolling-element bearing has three functions: 1. To reduce rolling and sliding friction. 2. To protect the highly finished surfaces of the rolling elements and races against corrosion and pitting. This is extremely critical to service life. 3. To act as a sealant. Generally , rolling bearings are lubricated with grease or oil. Grease is generally preferred as it is easier to retain in the housing, provides a better barrier to contaminants and is less likely to drain away from the bearing surfaces, therefore providing more reliable lubrication. Oil, on the other hand, is a better lubricant and is essential for high speed and high temperature operation (over 100 C). Bearings are greased by completely filling the bearing (not the housing). They are then occasionally re-lubricated to maintain maximum efficiency. The simplest form of oil lubrication is an oil bath. The oil level should almost reach the centre of the lowest rolling element. This method is acceptable provided that the operating temperature is not excessive. Another common and simple method is oil splash. This method is frequently used in applications such as gearboxes. If the operating temperature excludes the use of an oil bath/splash then an oil circulation system or oil injection system must be used. This aids in the cooling of the bearing.

Sources of friction Rolling action is the predominant source of friction in a rolling-element bearing. Rolling friction arises from metal deformation when a ball or roller under load moves along the surface of a race. A buildup of deformed material preceding the rolling element offers resistance to motion which, in turn, creates frictional heat. Other

less prominent sources of frictional heat in rolling-element bearing include sliding, slippage, and abrasive action. A limited amount of sliding friction results when the spacers between rolling elements contact the raceway. Friction also arises from slippage of rolling elements. This occurs in the unloaded region of a bearing where clearance between rolling elements and raceway is a maximum. Slippage also increases with decreasing speed because the reduction of centrifugal force on the rolling elements results in greater clearance. Friction can also result from rusting or corrosion of metal surfaces which produces abrasive oxide particles.

Oil Bath Lubrication Oil bath lubrication is widely used method in the case of low or medium speeds. The oil is picked up by the rotating bearing elements. The oil should be at the centre of the lowest rolling element. It is desirable to provide a sight gauge so the proper oil level may be maintained.

Dripping Oil Lubrication This method is suitable for the application where a small quantity of lubricating oil is constantly fed into the bearing without interruption. The excessive amount of oil dripped may cause a rise in the temperature of the bearing.

Oil Jet Lubrication Oil jet provides a very effective lubricating method for high speed applications. It is important to ensure that sufficient amount of oil will reach the bearing components and will be able to dissipate the heat generated by friction. The velocity of the oil jet, usually 15 M/S, must enable some of the oil to penetrate through the turbulent air membrane surrounding the bearing. The position of the oil jet should be placed between the inner race and cage of the bearing.

Oil Mist Lubrication This method is often used for high speed applications, such as grinding spindles. The oil mist is produced in an atomizer. Dry compressed air, after filtered, is used in the oil lubricator. The oil is

then introduced into the bearings. The air current will also serve to cool the bearing, and its slightly higher pressure in the housing will also prevent impurities from entering. The small quantity of oil can be regulated so that the lubricant friction is practically negligible.

Solid Lubrication Sometimes it is found that the addition of a small amount of solid lubricant, such as MoS2, into the grease will greatly improve the lubricating properties. In some special cases where very high temperature or high vacuum prevails, solid lubrication will be the solution.

Oil Circulation System Circulating lubrication is commonly used for high speed operation requiring bearing cooling and for bearings used at high temperatures. Oil is supplied by the pipe at the top, it travels through the bearing, and drains out through the pipe an the left. After being cooled in a reservoir, It returns to the bearing through a pump and a filter. The oil discharge pipe should be larger than the supply pipe so that an excessive amount of oil will not back up in the housing . This method of lubrication is the common practice of spindle design for high speed precision machine tools.

Oil Injection System Jet lubrication is often used for ultra high speed lubrication such as bearings in jet engines with a dm.n value (dm: pitch diameter of rolling element set in mm; n: rotational speed in rpm) exceeding one million. Lubricating oil is sprayed under pressure from one or more nozzles directly into the bearing. In the diagram the lubricating oil is sprayed on the inner ring and cage guide face. In the case of high speed operation, the air surrounding the bearing rotates with it causing the jet to be deflected. The jetting speed of the oil from the nozzle should be more that 20% of the circumferential speed of the inner ring outer surface

Oil Splash Lubrication With this lubrication method, oil is splashed onto the bearings by gears or a simple rotating disc installed near bearings without submerging the bearings in oil. It is commonly used in automobile transmission and final drive gears. The figure shows this lubrication method used on a reduction gear

The high contact pressures in a rolling-element bearing also elastically deform the rolling surfaces to enlarge the contact area which supports the load. The combination of surface deformation and hydrodynamic lubricating action produces a thin, elastohydrodynamic (EHD) lubricant film which provides for lubrication within the contact zone of rolling-element bearings. Grease lubrication: Greases provide a lubricating film on the surfaces of rolling elements, separators, and raceways. The lubricant is actually a thin film of oil which is released as the three-dimensional fibrous network within the grease is ruptured under shear. Only that portion of the grease in intimate contact with moving surfaces breaks down; the balance remains intact and functions as a sealant. When a freshly charged bearing begins to rotate, grease is thrown from the rolling elements and is rapidly circulated through the housing. After a short time, grease from the rotating outer race is then thrown back onto the rolling elements where shearing takes place. This turbulent environment at the outset of rotation creates frictional heat which attains a maximum and then gradually diminishes as the continual shearing action releases a lubricating oil film. As lubrication takes effect, the temperature of a properly charged bearing will drop and assume equilibrium.

Choice of lubricant The choice between oil or grease lubrication depends on the relationship of journal speed to viscosity. Slower journal speeds have higher viscosity requirements while high speeds call for a light-bodied oil. Bearings designed for low speed operation usually have a relatively large clearance between shaft and housing, while high-speed bearings usually have a much smaller clearance.

Rolling Element Bearing Applications Bevel Gear Pinion Requirements = Pinion must be accurately located axially and have a high radial capacity at the gear end

LH bearing = X arranged taper roller bearings locate the shaft and carry the high axial load. Spacers accurately control the axial clearance in the bearing pair. X arrangement is less sensitive to misalignment allowing the roller bearing to locate the shaft radially.

RH bearing = NUP roller bearing has a high radial load capacity. The NUP design requires no locating shoulders that would make assembly difficult. The sliding fit is in the housing

Car Wheel Requirements = Wheel must be accurately located with small clearance in the bearings

LH bearing = The O arrangement allows the bearing clearance to be adjusted through the inner race. Since the load direction is stationary relative to the shaft the inner race can be a sliding fit.

RH bearing = O arranged taper roller bearings resist side forces on the wheel better than X arranged bearings. This is important due to the short shaft.

Electric Motor Requirements = Simple and quiet. Axial forces are low but shaft speed is high

LH bearing = Deep groove ball bearing axially locates the shaft

RH bearing = This is the output end of the shaft and the bearing floats in the housing Bearing housings differ in design depending on the application and serve to support the bearing and contain the lubricant. Suitable seals are usually provided to exclude water, dust, dirt or other external contaminants from the bearing components and to prevent leakage of the lubricant from the housing.

Rolling Element Bearings Selection The table below demonstrates the features of each bearing type enabling a suitable bearing to be chosen for a particular application.

Having chosen the bearing type the bearing size can be calculated. A bearing is selecting based on its quoted value of static load carrying capacity or dynamic load carrying capacity. If the bearing is stationary for long periods or rotates slowly and is subject to shock loads then the bearing selection procedure is based on the Static Load Carrying Capacity. For more continuous operation the bearing is selected on the Dynamic Load Carrying Capacity.

Static load Carrying Capacity The Basic Static Load Rating given for each bearing in bearing catalogues is based on the stationary axial and radial forces acting on the bearing. When bearings are subject to both radial and axial loadings the equivalent static load must be found thus: Po = XoFr + YoFa If only radial forces act then Po=Fr Where Po = The equivalent static bearing load (N) Fa = static axial load on the bearing (N) Yo = static axial factor

Fr = static radial load on the bearing (N) Xo = static radial factor

N.B. The values of Xo & Yo are given in the Bearing Data The Basic Static Load Rating Co can be calculated from Co = SoPo where

Co = basic load rating (N) , o = static safety factor and Po = equivalent static bearing load (N)

Values of So depend on the type of bearing and the requirements regarding quiet running. Guidelines on Static Safety Factor

If the bearing is stationary for long periods or rotates slowly and is subject to shock loads then the bearing selection procedure is based on this basic load rating. Values of Co for each bearing are quoted in the bearing catalogues. Choose a bearing whose quoted value for Co is equal to the required value of Co calculated above.

Dynamic Load Carrying Capacity The dynamic load carrying capacity of a bearing is dependent on the dynamic forces acting on the bearing as well as the basic static forces. The first step is therefore to calculate the Static Load Rating before continuing with the following procedure. If the bearing is subject to radial and axial forces then an equivalent dynamic bearing load must be calculated. P = XFr + YFa P = The equivalent dynamic bearing load (N) Fa = static axial load on the bearing (N) Y= axial factor

Fr = static radial load on the bearing (N) X = radial factor

When Fa = 0 or Fa is relatively small up to a limiting case of Fa/Fr = e (where e is a certain limiting value) then P = Fr N.B. The values of X, Y & e are given in the Bearing Data Once a value for the equivalent dynamic bearing load is obtained it can be used to calculate the dynamic load rating of the bearing. This value is used to select the bearing. Each bearing in the bearing catalogue has a quoted value for dynamic load rating and so a bearing should be chosen that has a higher rating than the one calculated. The dynamic load rating that is quoted in the catalogues for each bearing is dependent on the required life of the bearing and the equivalent dynamic bearing load (P). The ISO equation for basic rating life is

Where L = basic rated life, Millions of revolutions P = equivalent dynamic bearing load p = 3 for all ball bearings

C = basic dynamic load rating p = exponent in the life equation p = 10/3 for all other roller bearings

Basic Rated Life of Bearings The basic rated life (defined as the number of revolutions that 90% of a group of identical bearings would be expected to achieve) is determined from the length of life that is required of the bearing. Typical life expectancies of required of various machines are given below. Machine Usage intermittent- domestic machines short periods- hand tools, construction machines high reliability for short periods- lifts, cranes 8 h/day partial use gears, motors 8 h/day full use machine tools, fans continuous use

Hours 300-3000 3000-8000 8000-12000 10000-25000 20000-30000 40000-50000

N.B. The Following figure shows that the same calculated fatigue life as for ball or cylindrical roller bearings can be achieved under the same combined axial / radial load by a tapered roller bearing of a much smaller bore ; alternatively, a tapered roller bearing of the same bore can achieve much greater fatigue life.

Fig. 2

The calculated fatigue life for ball , cylindrical roller bearings and a tapered roller bearing of a much smaller bore

True rolling motion The extensions of the raceways and rollers of a tapered rollerbearing are designed to converge at a common point on the axis of rotation (fig. 2). This results in true rolling motion of the rollers on the raceways, at every point along the roller body.

Fig. 2 On-apex design results on true rolling motion at all points along the roller body.

Radial & thrust load capability The angled raceways allow the tapered roller bearing to carry combinations of radial and thrust loads. The greater the angle between the outer race and bearing center line the greater is the ratio of thrust to radial load capacity (fig. 3). Long line roller/race contact gives the tapered roller bearing a high load carrying capacity. This and the capability to carry radial loads, thrust loads, or any combination of the two, makes tapered roller bearings the ideal choice for many applications. For a given bore it is possible to select a specially light or heavy section to meet application load/duty requirements (fig. 4)

Fig. 3 Designs to support radial and thrust loads in Combination

Fig. 4 Designs to suit the space available

Bearing damage Analysis Recognize and preventing Damage Recognizing and preventing damage can dramatically increase bearing life and decrease the potential for improper handling, installation and adjustment. Additionally, it can highlight other conditions associated with the system design and/or operation. Our service engineering team can help customers apply proper practices in their plants, helping to decrease downtime Damage to bearings while handling before and during installation and damage caused by improper installation, setting and operating conditions are, by far, the causes of the largest percentage of premature trouble. In the following, examples are shown of the most common types of damage and some of the causes of this damage. In many cases the damage is easily identified by the appearance of the bearing, but it is not easy and sometimes it is impossible, to determine the exact cause of that damage. As an example, a bearing with scored and heat discolored roller ends and rib is easily identified as a burned up bearing and damaged beyond further use. The cause of the burning or damage, however, might be traced to any one of a number of things such as insufficient or improper lubricant. It may be the wrong type of lubricant or the wrong system for supplying lubricant. Perhaps a lighter or a heavier lubricant is needed or an extreme pressure type of lubricant rather than a straight mineral oil and a circulating oil system needed rather than an oil level or splash system. This type of damage could be caused by excessively tight bearing setting or a combination of too tight setting and inadequate lubrication. From this it can be seen that simple examination of a bearing will not reveal the cause of the trouble. It can reveal if the bearing is good for further service, but often it is necessary to make a thorough and complete investigation of the mounting, installation and parts affecting the bearing operation to determine the cause of the damage. Unless the true cause of the damage is found and corrected, the replacement bearing will be damaged in the same manner and again there will be premature trouble. This information is not an attempt to make "trouble shooters" or "bearing experts" of all who read it. It is intended to caution users about possible causes of damage and alert them to take preventive action. With proper precautions during the handling, assembly and operation of bearings, almost all damage can be prevented. It is much easier, and a great deal less expensive, to prevent damage than to determine and correct the cause of damage after the machine or equipment is in operation.

Typical modes of failures Mode of contact fatigue

Geometric stress concentration Geometric stress concentration fatigue results from locally increased stress at the ends of roller/race contact.

Point Surface Origin (PSO) PSO is fatigue damage that has its origin associated with surface asperities, which act as local stress concentrations.

Peeling This type of fatigue is characterized by a shallow < 2.5 m m (0.1 m in) deep, spalling which sometimes occurs locally around bruises, grooves, or ends of roller/race contacts where the EHD film is lost by leakage.

Transverse cracking fatigue

a) Non-propagating spall

Inclusion origin spall b) Spall propagated by hydraulic pressure Damage by mechanisms other than contact fatigue

Abrasive Wear

Spalling Wear from foreign material. Debris bruises on all contact surfaces due to hard particles in the lubricant

Brinelling Brinelling is the plastic deformation of bearing element surfaces due to extreme or repeated shock loads.

False brinelling False brinelling is recognisable by the grooves worn into the raceways by axial movement of the rollers

during transportation. Cage damage

Cage breakage Results of good practices In the preceeding comments, the results of bad handling, improper assemblies, adjustments and operating conditions have been stressed and the resulting damage shown. The following image shows what happens when there is good lubrication, good assembly and maintenance and the proper fitting practice for the bearing application has been followed. This bearing shows that, with reasonable care in machining the parts and in the assembly and maintenance, it is not difficult to get excellent life. This bearing operated for over 400,000 km (250,000 miles) in a bus and is still in excellent condition and probably would run for many more kilometres.

How to determine probable causes

PLAIN BEARING FAILURES PLAIN BEARING DAMAGE The identification of the cause of damage or failure in plain bearings is crucial before remedial action can be taken to prevent further recurrence. The first step is a visual examination of the damaged component. The following photographs can then be used to find a match. "A picture is worth a thousand words." However, do not stop at this stage. More than one mechanism of failure can lead to superficially similar bearing damage. Moreover, once failure has been initiated, the original mechanism may lead to other failure mechanisms and it is essential to identify the original cause. We are primarily concerned with bearings from industrial machines. The vast majority of these are designed to operate with a hydrodynamic lubricating film and are based on white metals (Babbitts) - alloys of tin and lead. Most of the photographs are of damaged white metal bearings, but some examples of other materials are included. White metals are chosen because they are of low hardness and melting point, so that, in the event of failure, damage should be limited to the bearing, a comparatively cheap and readily replaced component, leaving the expensive rotor unharmed. These desirable characteristics, however, place limits on their operation (see below). The effect of temperature on the hardness of white metals, and hence their ability to carry load, is critical:

Suggested Temperature Limits for Whitemetal Bearings

OPERATING LIMITS The Figure below gives a working envelope for steadily-loaded journal bearings in terms of the main operating parameters: load, speed, lubricant viscosity and bearing geometry.

<> Load Capacity Diagram for Steadily-Loaded Plain Journal Bearings (EDSU 66023 )

The Figure below gives a working envelope for steadily-loaded thrust bearings in terms of load capacity and operating speed.

<> Limits of Safe Operation Note. Both these figures are for steadily-loaded operation. Under alternating load conditions, such as occur in the bearings in reciprocating machines or in the bearings of rotating machines subject to vibration (rotor out-of-balance, rotor instability), a separate limitation is imposed by the fatigue strength of the bearing material. In the case of reciprocating machines, higher strength bearing materials can be used, but with the risk of damage to the bearing journal.

Index of Failure and Damage Mechanisms

1. Lubrication Breakdown 2. Wear 3. Fatigue 4. Cavitation Erosion 5. Fretting Damage 6. Chemical Effects 7. Thermal Ratchetting of Tin-Rich White Metal 8. Electrical Damage 9. Incompatible Materials 10.Manufacturing and Assembly Errors

1. Lubrication breakdown 1. Start-up wiping damage 2. Excessive load at start-up 3. Loss of clearance in tilting-pad journal bearing at start-up 4. Typical wiping failure 5. Lubricant breakdown with lead-bronze bearing 6. Summary 1.1 Start-up wiping damage

Minor rub in bottom of bearing subject to a vertical load Main Characteristics Minor wiping - melting and resolidification of white metal at right hand side of damage patch (the black axial mark is a photographic artefact and should be ignored) - in centre of bottom half of bearing subjected to a vertically downward load.

Cause Damage is caused by rubbing of journal before hydrodynamic lubricating film has developed.

Note When hydrodynamic film is present the load-line, and the minimum film thickness, is displaced in the direction of rotation. Hence any damage due to small dirt particles, transient lubrication film breakdown is displaced from the centre of the loaded half of the bearing.

Possible Confusing Damage Fretting damage caused by vibration from an external source when machine is not operating with the journal resting in the bearing also occurs on the bottom centre line.

Comments This type of damage is limited to heavily loaded bearings, particularly in machines subject to frequent starts. The main effect is optical with the damage no more than a witness mark; it has no effect on the performance of the bearing and is only found when the bearing is examined at overhaul. Minor wiping or plastic deformation can be a desirable feature, assisting run-in, the bearing generating its own profile to suit the operating conditions. Similar witness marks may also be found at the edges of a bearing where there is misalignment or manufacturing error.

1.2 Wiping damage caused by excessive load at start-up

Wiping of white metal thrust pad caused by successive starts under escessive load

Main Characteristics Obvious melting of bearing metal that has been carried over the trailing edge of the pad and solidified in layers where subjected by cool oil flowing between the pads.

Cause Excessive thrust load at starting has meant that the bearing metal has been subjected to rubbing and frictional overheating before the development of hydrodynamic lubricating film. The molten white metal has resolidified on coming into contact with the cool oil flowing between the pads. The photograph shows successive layers of resolidified white metal following a number of starts.

Comment If load at start up cannot be reduced, the only cure is to use a larger bearing.

1.3 Loss of clearance in tilting-pad journal bearing at start-up

Loss of clearance in five pad, tilting-pad journal bearing at high-speed start-up causing wipe at the centres of all the pads

Main Characteristics Small patch of wiping in centre of all of the pads.

Cause Establishment of thermal equilibrium in a tilting-pad journal bearing takes some time because of the restricted thermal path through the pad pivots. With rapid start-up, sufficient radial thermal

expansion can occur in the pads to take up the radial clearance, preventing the formation of a hydrodynamic lubricating film and causing wiping at the centres of the pads.

Comment This type of failure can be avoided by extending the run up time to full speed or by increasing the radial clearance in the bearing if this is acceptable for the performance of the bearing for the normal operating conditions.

1.4 Typical wiping failure

Wiping of journal bearing showing resolidified white metal in the central circumferential oil groove

Main Characteristics The photograph shows a typical wiping failure with the molten white metal from the bearing surface solidifying in a cooler part of the bearing, in this case the central oil supply groove.

Cause Wiping is caused by loss of the hydrodynamic lubricating oil film during running. There are a number of reasons why this can occur. These are discussed in the summary sheet at the end of this Section.

Possible Confusion Damage Wiping can be a secondary event initiated by fatigue damage or electrical erosion reducing the effective load carrying area of the bearing. The signatures of these failure mechanisms are described in the appropriate Sections and should be looked for, particularly in the nominally unloaded parts of the bearing.

Comment Do not change to a bearing material with higher load-carrying capacity unless it can be established that failure has been the result of inadequate strength in the white metal. Treating the symptoms is not the way to provide a cure. The cause of failure must be identified

1.5 Lubricant breakdown with lead-bronze bearing

Crack in crankpin of reciprocating compressor caused by lubrication breakdown failure of lead-bronze big-end bearing

Main Characteristics The photograph shows a section through one 10 inch diameter crankpin from a 6-throw reciprocating compressor after a lubrication failure. The lead-bronze bearing had wiped, but, in contrast to the case with a white metal bearing that limits the temperature rise to about 240°C, much higher temperatures have been reached and this has resulted in quench cracking of the steel crankpin and total loss of a large crankshaft.

Cause This is a lubrication breakdown failure (see summary sheet at end of section for possible reasons).

Note When wiping occurs with a lead containing copper alloy (lead bronze, copper lead), only a small proportion of the bearing metal melts (the lead phase); high enough temperatures are generated by rubbing against the solid material leading to cracking of the journal by repeated quenching as it leaves the load zone.

Comments White metal bearings should be the first choice where loading conditions permit their use. Leadbronze can be used for little-end bearings where loads tend to be high, but the lower speed oscillating conditions do not tend to give rise to quench cracking. Note that hardened steel journals have to be used with these harder bearing alloys. As a rule of thumb journal hardness should be at least three times that of the bearing metal.

Lubrication Breakdown: Summary Characteristics Melting of white metal, often resolidifying in cooler parts of the bearing (e.g. in the oil grooves, or in the upper half of the bearing, though in the case of minor wipes it will be confined to the trailing edge of the damaged region. Re-deposited bearing material does not form a metallurgical bond with the underlying material and can often be easily peeled off (particularly if whitemetal). Circumferential scoring of wiped bearing usually occurs.

Possible Causes Wiping is caused by a failure to form a hydrodynamic oil film at start up or reduction in lubricant film thickness (sometimes complete breakdown of the lubricating film), resulting in overheating and melting of the bearing material (partial melting in the case of the higher strength lead-bronzes and copper-leads). It should be noted that in high-speed machines wiping can occur without complete film breakdown if the temperature in the oil film rises above the melting point of the bearing material (about 240°C in the case of whitemetals). A number of possible causes are listed below. It is unlikely that failure occurs as a consequence of poor bearing manufacturing quality. 1. Inadequate supply of lubricant. This can occur at start up in low temperature conditions if the lubricant viscosity in the reservoir is so high that flow rate is too low. (Viscosity at start up should not exceed 2000 cSt.) 2. Use of lubricant of too low viscosity grade for the application 3. Lubricant supply is too hot leading to thin lubricant film

4. Misalignment leading to edge loading. 5. Excessive load causing high temperature and thin lubricant film. (Excessive start-up load, preventing or delaying the generation of a hydrodynamic lubricating film can be a problem with hydrodynamic bearings.) 6. Excessive shaft orbiting due to bearing instability (e.g. sub-synchronous whirl), but this can also give rise to fatigue failure. 7. Shaft operating load line close to an oil groove causing breakdown of the oil film or oil starvation. The particular case of gear bearings, where the load line is a combination of gravity and gear thrust, should be noted. 8. Loss of bearing area by some other failure mechanism, e.g. fatigue, electrical erosion damage.

Actions The first check is whether the failure has occurred for the first time or whether there is a history of similar failures on the particular machine, or other similar machines. The operating conditions of the equipment leading up to failure are very important in understanding the cause. For example: has the failure occurred at start-up; was there an increase in vibration levels; has there been a problem with the lubricant pump or was the oil at the correct operating level in a self-contained system; was the filter element blocked? The profile of the wiping patch may indicate the occurrence of misalignment. Measuring the shell thickness can in some instances indicate the exact position of the shaft when the failure occurred. Check that the position of wiping corresponds to the predicted minimum film thickness in the bearing (this can be calculated using the techniques listed below). If the actual and predicted positions do not correspond then look for conditions that may result in the applied load being in a different direction to that expected. Note again that changes in gear loading will alter the direction of the resultant load. Check damage to the mating shaft. Surface scoring and/or thermal cracking can occur; this risk is greater with harder bearing materials (e.g. copper and aluminium alloys).

Useful Analysis Techniques Calculation of bearing operating conditions (lubricant flow rate, temperature, lubricant film thickness, attitude angle) can provide good supporting evidence. See, for example, ESDU Data Items 84031, 90027 (journal bearings), 82029, 83004 (thrust bearings).

2. Wear Bearing materials are chosen to be soft, not only does this allow them to conform to the journal or thrust collar to accommodate slight misalignment (conformability), but also to allow hard contaminant particles to embed so that they do not score the counterface (embeddability). There is, however, a limit to the latter and with severely contaminated lubricants particles small enough to enter the oil film at the supply groove, but larger than the minimum oil-film thickness can remain proud in the soft bearing metal and score the shaft or thrust collar - scoring. Scoring results in an increase in clearance that eventually requires the bearing to be replaced. More seriously, however, wear in non-cylindrical profile bearings (e.g. lemon, multi-lobe), used to give enhanced stability, tends to produce a more cylindrical bore that can result in the bearing developing instability. Hydrodynamically-lubricated bearings can operate satisfactorily with a certain amount of scoring; however, a stage is eventually reached when it is no longer possible to generate an oil film at start-up and wiping failure occurs (see Lubrication breakdown). Particles in high concentration, but small enough to pass through the oil film can erode the soft bearing material in the direction of oil flow in the bearing - erosion. Larger particles trapped in the oil grooves can also cause erosion damage. Other wear mechanisms such as fatigue, cavitation erosion, fretting, chemical attack and electric erosion are dealt with separately.

2.1 Scoring damage of white metal thrust pad by dirt

Scoring of thrust pad by dirt in oil Main Characteristics Score marks follow the direction of motion in the bearing, like the grooves in a gramophone record. The score marks start where the dirt particle bridges the film and form a continuous mark up to the end of the bearing surface, the trailing edge on the right of the pad in the Photograph, or stop with an embedded particle (see following example). Scoring is an example of abrasive wear; adhesive wear gives discontinuous tears rather than clear uninterupted scores.

Cause Contamination of the lubricant with hard particles small enough to enter the oil film at the point of maximum film thickness, but too large to pass right through without bridging the film. See also in the photograph how the soft white metal has been dragged over the trailing edge of the pad

Comments Scoring damage indicates inadequate filtration of the oil or failure to change the oil filter when necessary.

2.2 Scoring of white metal journal bearing

a. Dirt particles embedded in white metal bearing

b. Iron print from bearing surface

Main Characteristics Photograph ‘a’ shows typical score marks in the white metal surface ending in an embedded foreign particle.

Cause Contamination of the lubricant with hard particles small enough to enter the oil film at the point of maximum film thickness, but too large to pass right through without bridging the film.

Note The embedded particles may be analysed to identify the source of the contamination. For example, iron or steel particles can be identified by iron printing using absorbent paper soaked in potassium ferrocyanide solution. Photograph ‘b’ shows an iron print from a bearing surface.

Comment Scoring damage indicates inadequate filtration of the oil or failure to change the oil filter when necessary.

2.3 Scoring of lead-bronze journal bearing

Gross scoring damage to lead-bronze journal bearing Main Characteristics Deep circumferential scoring of bearing surface in the direction of rotation.

Cause Gross contamination of lubricating oil.

Note Note the continuous score marks and the dragging over of the bearing metal at the downstream end of the half bearing (right hand side in Photograph).

Comment This was a main bearing of a three-throw reciprocating pump. The damage occurred at commissioning because the crankcase was not cleaned out before starting.

2.4 Erosion damage of white metal journal bearing

Erosion of surface of bearing caused by the presence of large amounts of small size dirt particles in the oil

Main Characteristics The arrow in the photograph shows the direction of rotation. The erosion damage is clearly caused by dirt particles entering via the oil feed holes in the central circumferential oil groove. Erosive wear is shown by roughening and dulling of the white metal surface.

Cause Contamination of the lubricant by fine particles.

Possible Confusion with Other Types of Damage Cavitation of the oil film gives similar looking damage (see cavitation erosion), but in the present case the damage is in the centre of a bearing land away from the oil feed.

2.5 Erosion damage to axial oil groove in journal bearing

Erosion damage to inlet oil groove caused by particles in oil too large to enter oil film Main Characteristics Edges of oil groove have been eroded away by particles trapped in oil groove as they were too large to enter bearing oil film.

Cause Oil contamination.

Note Axial oil grooves should have dirt escape gutters cut at 45° at ends of groove. These gutters should be deep enough to discharge the dirt particles, but not so deep that they cause excessive sideways drainage of oil from the groove.

Comment This damage indicates inadequate filtration of the oil.

2.6 Summary Characteristics Wear by hard particles larger than the minimum oil film thickness in the bearing give characteristic continuous scores extending to a discontinuity in the bearing surface (oil groove, trailing edge of thrust pad) or ending in an embedded foreign particle. Erosive wear is a fatigue process causing small scale removal of material that results in a characteristic roughening of the surface. Superficially this can resemble cavitation damage (see cavitation erosion), and electric erosion/spark erosion, but the former occurs in the centre of bearing lands away from oil feed holes and the latter results in small rounded shiny melt pits that are quite different in appearance.

Possible Causes Two causes of wear damage are discussed in this section: abrasive wear caused by the cutting action of hard particles that bridge the bearing oil film; erosive wear, the small scale removal of material by fatigue resulting from the mechanical action of small particles repeatedly striking the surface.

Actions Where damage is caused by solid particles in the lubricant, the only remedy is improved filtration of the oil. This should be no problem where there are duplicate filters, but single filters with overpressure relief can release particles into the lubricant when the filter becomes blocked and there is a risk of particle release when filter changes are made. Magnetic plugs should be fitted to reduce damage by ferrous wear particles.

Analytical Methods Analysis of filter debris may be useful in identifying the source of solid contaminants, with iron printing using absorbent paper soaked in potassium ferrocyanide a useful technique for the identification of ferrous particles embedded in soft bearing materials.

3. Fatigue White metal bearing alloys have low strength and readily suffer fatigue damage when subject to reversing loads, such as occur in reciprocating machines (engines, reciprocating compressors, ram pumps), and also when subject to vibration caused by out-of-balance or dynamic instabilities. White metals are always used as a lining on a backing of a harder material, usually mild steel, but more rarely bronze. The fatigue strength of the white metal can be increased by reducing the thickness of the lining, but with the penalty of reducing the conformability and embeddability of the bearing material. White metal thicknesses of 1 to 3 mm (thick wall bearings) are most commonly used in industrial machines; this not only gives good conformability and embeddability, but may also provide a molten layer that allows a machine to be run down safely in the event of failure without damage to the rotor. Thick wall bearings can be repaired by re-metalling. Thin wall bearings with white metal linings in the range 0.08 to 0.12 mm are only used in reciprocating applications, in particular highspeed engines, to give enhanced load-carrying capacity; these bearings cannot be repaired. See Section 7 for a different fatigue mechanism, thermal fatigue. Mechanical erosion (Section 2) and cavitation erosion (Section 4), which are other forms of fatigue damage, are also treated separately.

3.1 Fatigue of White Metal Journal Bearings

Initial stages of fatigue showing ‘bruising’ of the white metal surface

Both photographs show the characteristic crazy cracking of the white metal Main Characteristics surface. Where loose pieces are formed these may remain in the bearing

(photograph ‘b’) or come away and get trapped so that they melt and wipe, but this is a secondary effect, not the primary damage. The loose pieces tend to have an aspect ratio (length / thickness) of about 5:1, so that on thick wall bearings the fatigued areas are large

Cause Note

Excessive alternating loading. Photograph ‘b’, where the fatigue damage is on one side of the bearing suggests that there may have been slight misalignment.

3.2 Fatigue of White Metal Journal Bearing Caused by Vibration

White metal journal bearing showing fatigue in both bearing halves

Main Characteristics

Cause

Note

Fatigue of journal bearing removed from a high-speed centrifugal compressor (16,000 rev/min) following a 12-hour vibration incident. Synchronous vibration causing excessive loading on the white metal. Note that the fatigue patches are diametrically opposed, showing that the vibration mode was elliptical. The circumferential bands of fatigue correspond to the dovetail grooves that were machined in the bearing shell. This bearing remained in service for six weeks after the vibration had been suppressed. This shows the remarkable load-carrying capacity of the oil film formed by hydrodynamic action. The loss of bearing material would of course have prevented the bearing from starting up satisfactorily.

Comment

This type of machine was prone to a peculiar type of intermittent synchronous vibration instability. In one machine, with a rotational speed of 26,000 rev/min, the manufacturer supplied lead-bronze bearings with a lead overlay in an attempt to prevent fatigue failure, though this was treating the symptoms rather than addressing the real problem. Bearing fatigue still occurred,

leading to lubricant film breakdown and cracking of the expensive rotor by thermal shock (see Section 1: Lubrication Breakdown with Lead-Bronze Bearing) that incorporated an integral gear. The problem was successfully solved by changing to a modified design of bearing (elliptical bore) that inhibited the instability.

3.3 Premature Fatigue of White Metal Journal Bearing

Premature fatigue caused by poor bonding of white metal to bearing shell

Exposed dovetails in machined oil groove showing lack of bonding between the edges of the dovetail and the white metal (the black lines at the edges of the dovetails)

Main Characteristics

This has the appearance of typical fatigue cracking. The tell-tale ‘bubbles’ in the central oil groove that have been caused by vaporisation of oil trapped in the cast iron bearing shell point to poor bonding. This suggests that the fatigue damage has also been the result of poor bonding.

Cause

Note

Premature fatigue from inadequate bonding of the white metal to the bearing shell resulting in decreased fatigue strength of the white metal lining. te the presence of the dovetails in the bearing ell. This has been used in the mistaken pression that the strength of the white metal is reased by giving it a mechanical bond. The rmful effect of dovetails in preventing the mation of a metallic bond because of the different rmal conductivities and rates of expansion in el and white metal is shown in photograph ‘b’ ere the machined oil groove has exposed the vetails revealing lack of bonding between the ite metal and the sides of the dovetails.

Possible Confusion with Other Types of Damage The appearance of premature fatigue is very similar to that or normal fatigue. It can be confirmed by removing loose pieces and examining the bearing shell. In the case of normal fatigue the tin lining on the shell will still be present.

Comment Cast iron should not be used for bearing shells. Cast iron is porous and absorbs lubricating oil in service. This oil is almost impossible to remove and prevents satisfactory re-metalling of the shell. The bonding should be checked by ultrasonic testing before re-metalled bearings are put into service.

3.4 Fatigue Failures: Summary

Section through fatigued white metal showing crack penetrating to the backing layer, but not to the backing metal

Fatigued white metal showing presence of tin layer on steel backing

Characteristics Fatigue damage is readily distinguished from other failure mechanisms. It manifests itself by crazy cracking on the surface; the cracks propagate through the white metal to the tin lining on the bearing shell where they continue until joining up with another crack and form a loose piece that detaches from the backing. These loose pieces have an aspect ration of about 5:1 (Fig. ‘a’). The full fatigue strength of the white metal lining depends on a sound metallurgical bond between the backing material and the white metal that is obtained by tinning the backing before pouring on the molten white metal. If a sound metallurgical bond is not created, the fatigue strength is markedly reduced, making the bearing susceptible to premature fatigue failure. Normal and premature fatigue can be distinguished by examining the bottom of the fatigue pit; this should show a coating of tin on the backing, not steel or a rusty surface (Fig. ‘b’). Fatigue failure is normally confined to journal bearings and rarely occurs with thrust bearings. The loose pieces of white metal can remain in place, held in by the closing fitting journal, or disappear by getting trapped so that they rub and wipe’.

Possible Causes Fatigue in journal bearings can result from inadequate design, but is usually the result of excessive alternating loading.

Actions Where fatigue is the result of excessive load, the only remedy is to change to a more fatigue resistant bearing material. Where it is a consequence of vibration, the vibration level has to be reduced.

4. Cavitation Erosion This is a particular form of fatigue caused by rapid fluctuation of pressure in the bearing oil film. When the pressure is low, bubbles of vapour or dissolved gas are formed and then collapse as they go into a high pressure region. Vaporous cavitation, where the bubble collapse is much more violent, results in shock waves in the lubricant film that cause fatigue failure in the white metal surface. This differs from normal fatigue in that small pits are formed rather than loose pieces. Cavitation damage occurs where there are reciprocating loads, either as part of the normal loading cycle or because of high-frequency vibration. It can also occur in bearings where there are sharp discontinuities in the thickness of the bearing oil film Gaseous cavitation, in which the bubbles are of gas from solution in the lubricant, is much less energetic than vaporous cavitation as bubble dispersion depends on re-absorption of the gas by diffusion rather than instantaneous collapse. Gaseous cavitation can still, however, cause damage to soft white metal bearings.

4.1 Cavitation Erosion Damage in Reciprocating Engine Bearing

Cavitation erosion damage to engine bearing caused by rapid approach of shaft to bearing, collapsing vapour bubbles in the oil film. (Photograph: Glacier Metal Co.)

Main Characteristics

Cavitation damage results in roughening of the bearing lands. It takes the form of arrows with the arrow head pointing against the direction of motion.

Cause

Rapid fluctuations of pressure in the oil film in reciprocating machine causing the formation of bubbles as the shaft moves away from the bearing followed by rapid collapse of the bubbles as the load is applied and the shaft approaches the bearing.

Note

The damage is in the middle of the bearing lands, away from the oil

groove and the oil feed holes. Possible Confusion with Other Types of Damage

Cavitation damage is very similar to the damage resulting from dirt erosion (see Section 2.4: Erosion damage of white metal journal bearing). In the latter case, however, the arrow head starts at the oil feed hole

Comment

This is a case of vaporous cavitation. For an example where the less violent gaseous cavitation was involved, see Section 4.3: Cavitation Damage Caused by Bearing Instability below and Section 6 "Copper deposit on thrust bearing".

4.2 Cavitation Erosion of White Metal Thrust Collar

Cavitation erosion of white metal thrust face

Photograph ‘a’

Steel thrust collar with lubricant feed grooves on end of gear shaft

Photograph ‘b’ Main Characteristics

Photograph ‘a’ shows typical roughening of white metal caused by erosion.

Cause

Photograph ‘b’ shows the thrust collar associated with the cavitation damage shown in Photograph ‘a’. This is an unusual arrangement with the oil grooves machined in the hard steel thrust collar, the reverse of the more normal arrangement with the grooves in the soft white metal.

Comment

The problem was caused by the sharp fall in pressure at the beginning of the grooves in the steel collar, followed by bubble collapse as the oil re-entered the high pressure region downstream of the groove. The problem was solved by changing to the more normal arrangement with the grooves in the white metal face.

4.3 Cavitation Damage Caused by Bearing Instability

Cavitation erosion damage to white metal journal bearing in high-speed refrigeration compressor subject to instability

Photograph 'a'

Part of Photograph 'a' at higher magnification

Photograph ‘b’ Main Characteristics

This is another aspect of cavitation erosion damage, in this case the bearing of a high-speed (25,000 rev/min) refrigeration compressor that was subject to a synchronous vibration. The higher magnification Photograph ‘b’ shows erosion damage resembling the effect of wave action on a sandy shore.

Cause

The refrigerant was Refrigerant R12 that was in contact with the lubricating oil. Because of the vibration, the refrigerant came out of solution in the oil as the shaft moved away from the bearing, only to go back into solution as the pressure increased as the shaft approached the bearing.

Note

This was a case of gaseous cavitation, giving a less energetic bubble collapse, but sufficiently energetic to cause cavitation damage to the bearings.

Comment

For another example where gaseous cavitation was involved, see Section 6 "Copper Deposit on Thrust bearing".

CAVITATION DAMAGE: SUMMARY Characteristics Cavitation erosion gives a fine scale roughened texture to the damaged surface. This could possible be confused with wear by fine particles (see Section 2.4: Erosion damage of white metal journal bearing) or even electrical erosion damage (see Section 8: Electrical damage), but can readily be distinguished by closer examination (electrical damage, for example, shows rounded shiny pits from which molten metal has been removed) and the circumstances in which the damage occurred. Possible Causes Cavitation damage is associated with high-speed reciprocating machines, or machines in which the bearings are subject to high frequency vibration. Gaseous cavitation should be suspected in machines handling lubricating oils with gas in solution, most particularly refrigeration compressors in which the refrigerant is in contact with the lubricating oil. Actions In many cases a modification of the bearing design can prevent bubble formation in a sensitive part of the bearing. Where cavitation is associated with vibration, it is a secondary effect and the cure is to remove the cause of the vibration.

5. Fretting Damage Fretting is a form of adhesive wear; it occurs as the result of small scale oscillatory movement. There are three situations in which this can result in damage in plain bearings. Fretting can occur at the contact of a journal in the bearing in a machine that is subject to external vibration. This causes a change in the bearing profile that can affect its performance. Secondly, fretting can also occur on the backs of bearing shells that are not given an adequate interference fit in the housing (see Section 10: Manufacturing and assembly errors). The third situation is at the contact between the pivots of tilting pads and the carrier, where the pads are subject to ‘fluttering’ through loading by a vibrating shaft

5.1 Fretting in a Journal Bearing

Fretting damage to journal bearing subject to vibration when stationary

Main Characteristics

Fretting when the machine is stationary occurs in line with the gravity load; in the Photograph this is in the centre of the bearing, not displaced to one side as it would be if the damage had occurred at the point of minimum oil film thickness during running.

Cause

This is caused by external vibration that causes small oscillatory movement at the point of contact between journal and bearing.

Comment

The damage resembles a start-up witness mark or wipe (see Section 1.1: Startup wiping damage), but is much more severe.

5.2 Fretting at Tilting-Pad Pivot

Thrust bearing pad (back side) with fretted pivot bar resulting from axial rotor vibration (Photograph: Glacier Metal Co.)

Main Characteristics

With steel contacts fretting produces a characteristic reddish-brown stain at the point of contact

Cause

This has been brought about by axial vibration which causes the pad to flutter with the resulting oscillating load.

FRETTING DAMAGE: SUMMARY Characteristics Fretting on white metal results in a blackening of the surface. With a journal bearing, it occurs when the shaft is stationary and hence the damage is in the direction of the gravity load. In this way it resembles start-up damage (see Section 1.1: Start-up wiping damage), but can be distinguished as a deposit rather than a witness mark or wipe. Fretting between steel surfaces (for example between the backs of bearing shells fitted with inadequate interference fit in the housing, or at the points of contact between the pivots of tiltingpads and the carrier in tilting-pad bearings) leaves a reddish-brown stain on the steel surfaces. Possible Causes Fretting is caused by small scale oscillatory movements. With journal bearings this can occur through external vibration transmitted to a stationary machine. Fretting damage can also occur on the backs of journal bearing shells that are not given an adequate interference fit in the housing. The latter is a fitting error and is dealt with in Section 10: Manufacturing and assembly errors. Fretting at the contact between the carrier and the pivot of a tilting-pad occurs if the pad is subjected to radial vibration (journal bearing) or axial vibration (thrust bearing) causing a ‘fluttering’ instability through the slight changes in load induced by the vibration. It can also occur with a lightly-loaded thrust pad, particularly the pads of the reverse thrust in a double thrust bearing if the axial float is such that the reverse pads are loaded by hydrodynamic action.

Actions Fretting with journal bearings is uncommon. If it is a problem the cure is to isolate the machine from the source of the vibration or to rotate the shaft at weekly intervals so that the point of contact is changed. When fretting occurs with tilting pads subject to vibration, the only remedy is to reduce the level of vibration. Double thrust bearings should be given an adequate axial float to prevent ‘fluttering’. As a guide the float should not be less than 0.1% of the mean diameter of the thrust collar with a minimum value of 0.1 mm. With point (ball) contact pivots, a hardened carrier has to be used. Analysis Fretting damage is normally easy to identify by visual examination. Fretting between steel surfaces results in the formation of the reddish-brown ferric oxide, alpha-Fe2O3. If confirmation of fretting is required, alpha-Fe2O3 can be identified by x-ray diffraction

6. Chemical Effects A number of different chemical effects can cause problems in bearings. Although these effects may be different, there are similarities both in the appearance of the damage and the resulting failure. It is thus convenient to treat them together. Three separate phenomena are described: 1. Chemical reactions taking place in the lubricant that result in the formation of deposits on the bearing surface. In most cases the rate of deposit formation tends to be temperature dependent so that deposit formation follows the temperature profile of the bearing surface. In general, the underlying bearing material is unaffected. 2. Chemical reactions between the bearing metal and the lubricant, lubricant degradation products or contaminants in the lubricant. This can have two effects: selective removal of different phases from the bearing metal or the formation of deposits on the surface of the bearing metal. Chemical reactions are temperature dependent and, once again, the corrosive removal of material or the formation of the reaction deposit reflects the temperature profile of the bearing surface. 3. Electrochemical reaction between materials in the lubricant, normally contaminants of some sort, and the bearing metal producing deposits on the bearing surface. The difference in this case is that electrochemical reactions are relatively independent of temperature so that the deposit forms uniformly over the bearing surface and is not restricted to the high temperature areas. The problems created by these effects are similar: a modification of the bearing surface profile that leads to a degradation in its hydrodynamic film generating capacity and ultimately to a breakdown in the lubricant film. As thrust bearings are generally more sensitive in this respect than journal bearings, failure by these chemical mechanisms tend to be more frequent in the former, though the effects also manifest themselves in the latter. This can be an advantage in failure investigation, providing evidence that may be lost in the complete destruction of a thrust bearing.

6.1 Lubricant Oxidation Deposits

Lubricant oxidation deposits on the trailing edge of a thrust pads

Lubricant oxidation deposits on trailing edges of thrust pads

Main Characteristics

Deposit forms in the loaded (high temperature) part of the bearing: the load line in a journal bearing, or the trailing edge of the pad in a thrust bearing (Photographs above). The deposit ranges in colour from reddish-brown to almost black depending on the degree of oxidation (a function of the temperature) and is characteristically patchy rather than continuous.

Cause

Excessive temperature. The rate of chemical reaction is a function of temperature. Although the time of exposure to the high temperature in the bearing is very short (milliseconds), oxidation of mineral hydrocarbon lubricating oils, even those containing anti-oxidants, can occur if the temperature exceeds about 150°C.

Note

Lubricant oxidation deposits are a clear sign of excessive temperature and are frequently associated with other effects such as wiping and plastic deformation. In the case of thrust bearings the deposits may be restricted to only a few of the pads, showing differences in the pad heights with the affected pads carrying most of the load.

Possible Confusion with other Types of Damage Comments

Other deposits described in this Section tend to look superficially similar. The final diagnosis requires chemical analysis. Oil degradation deposits are usually soluble or partly soluble in aromatic solvents and can be removed by wiping with a swab soaked in the solvent, but see Summary Sheet at end of this Section for possible analytical techniques.

6.2 Lubricant Degradation Deposits

Deposit of ammonium succinate in bearing of ammonia synthesis gas compressor

Main Characteristics

This is clearly a deposit from the lubricant that has been laid over the whole bearing surface apart from the loaded lands on the bottom half (right), where the high shear rate in the oil film has prevented settling. Diagnosis depends on chemical identification of the deposit.

Cause

This occurred in the bearings of the synthesis gas compressor on an ammonia manufacturing plant; this was the first such plant using a centrifugal compressor rather than reciprocating ones. The bearings and shaft seals on this machine were lubricated by a turbine oil from a common system. Chemical analysis showed the deposit to be ammonium succinate. Succinic acid is widely used as the corrosion inhibitor in turbine oils. Ammonium succinate was formed by reaction between ammonia leaking across the oil barrier shaft seals into the lubrication systems and the corrosion inhibitor in the oil. Ammonium succinate is soluble in water, but not in mineral oil. The lubricant that was also used for the steam turbine driver was slightly wet, the water dissolved the ammonium succinate which was then deposited in the hot bearing as the water boiled off.

Note

Note also the small patch of fatigue damage (see Section 3: Fatigue) on the right hand land of the upper half (left} and the signs of wiping (see Section 1: Lubrication breakdown) on both lands on the bottom half (right); the latter is probably a consequence of the deposit formation.

Possible Confusion with other Types of Damage Comments

Other deposits described in this Section tend to look superficially the same. The final diagnosis requires chemical analysis It was not possible with this system to prevent minor contamination of the lubricating oil with ammonia. The solution was to reformulate the oil with a corrosion inhibitor that did not react with ammonia. Such oils are now standard on ammonia synthesis plants.

6.3 Copper Deposit on Thrust bearing

Deposit of metallic copper on the surface of a fixed, inclined pad thrust bearing

Main Characteristics

Note the uniform appearance of the golden-brown deposit on the white metal bearing surface, in the oil grooves as well as on the bearing surface. This clearly has an electrolytic origin. Diagnosis depends on chemical identification of the deposit.

Cause

Deposit was identified by x-ray diffraction as pure copper. Copper deposits preferentially from solution on other metal surfaces that have a more negative Standard Electrode Potential (SEP). The SEP of copper = +0.345, tin = -0.136; hence copper will plate out from solution on to tin-rich white metal, just as it does on a steel penknife blade (SEP for iron = -0.441) dipped into a solution of copper sulphate. The lubrication system contained copper pipes and a copper alloy cooler, but the problem was to find a mechanism that would produce sufficient dissolved copper in the oil to plate out on the bearing. The bearing was from a centrifugal refrigeration compressor that used R12 as the refrigerant. R12 contains chlorine that could attack the copper, but R12 is stable to over 250°C well above any expected temperature in the system. The only hypothesis was that gaseous cavitation was taking place in the bearing (R12 is soluble in mineral oil) and that sufficiently high temperatures occurred on bubble collapse to decompose the R12, producing free chlorine that reacted with water in solution in the oil to give the required electrolyte. The bearing was changed to a tilting-pad one, where it was thought that pressure gradients in the oil in the bearing would be reduced. Although the hypothesis was never fully proved, the change in bearing type solved the problem.

Possible Confusion with other Types of Damage Comments

Other deposits described in this Section tend to look superficially the same. The final diagnosis requires chemical analysis The identification of the deposit in this case was simple. Once that fact had been established it became possible to look for an explanation. The

hypothesis suggested seemed improbable, but in the absence of any alternative suggestion, it seemed worth trying to modify the conditions in the bearing by changing the geometry. This was a standard package refrigeration unit and it subsequently turned out that, despite assurances from the manufacturer that it was limited to the company involved, in fact it was a common problem with these units.

6.4 Tin Oxide Formation on White Metal

Tin oxide deposit on trailing edge of thrust pad

Tin oxide formation on thrust bearing, the deposit covering the complete surface of the pads. Bearing failure has caused melting of the underlying white metal causing cracking of the brittle tin oxide layer

Main Characteristics

Deposit was identified by x-ray diffraction as tin oxide. The deposit of tin oxide is black and brittle. This is shown by the patchy nature of the deposit in Photograph ‘a’ and the way the deposit has broken up when not supported by solid white metal, Photograph ‘b’. (Rather similar to a thin layer of ice on mud, that fractures when stood on.} The particles of the deposit that have broken away in the case of the single pad (Photograph ‘a’) have scored the softer white metal in the direction of motion, though these scores do not extend to the trailing edge, probably because they have broken down to small particles of size below that of the oil film; this is suggested by the small radial ‘ticks’ at the ends of some of the score marks.

Cause

Tine oxide was a particular problem in marine steam turbines. Tin oxide forms by an electrochemical reaction of a tin-rich white metal and an electrolyte. This occurred in marine turbine sets as steps were taken to remove steam condensate from the oil. In the marine environment there was also some contamination by salt (sodium chloride) and as the oil was dried the concentration of the electrolyte increased until reaction occurred.

Note

The problem has become much less common now that the condition of the oil in marine steam turbine lubrication systems has been improved, but tin oxide formation still occurs occasionally in systems with sea water cooling and machines operating in marine environments.

Possible Confusion with other Types of Damage

Other deposits described in this Section tend to look superficially the same. The final diagnosis requires chemical analysis 6.5 Sulphur Attack of White Metal

Copper sulphide deposits on white metal thrust pads

Copper sulphide deposit on white metal journal bearing

Main Characteristics

Black deposit on the trailing edge of thrust pad, Photograph ‘a’ and in the load area in the bottom of a journal bearing, Photograph ‘b’ (right). The location of the deposits suggests chemical attack. Diagnosis depends on the chemical identification of the deposit.

Cause

The deposit was identified by x-ray diffraction as copper sulphide. Attack of the copper present in tin-rich white metal by reaction with active sulphur compounds in the process gas contaminating the lubricant forming deposit of copper sulphides (Cu2S, CuS} on the bearing surface.

Note

The reaction is a temperature dependent one. The fact that there is hardly any deposit present on the second pad from the right in Photograph ‘a’ indicates that this pad has been carrying little of the load, presumably as it is thinner than the others. Copper-free lead-rich white metal is more resistant to attack by sulphur, though under severe conditions (high temperature, high concentration of active sulphur compounds) a deposit of lead sulphide (PbS) may be formed.

Possible Confusion with other Types of Damage Comment

Other deposits described in this Section tend to look superficially the same. The final diagnosis requires chemical analysis Sulphur attack is more of a problem on floating-bush shaft seals lined with tin-rich white metal, where it can take up the clearance and cause seal failure. It can also cause breakdown of the lubricant film in thrust bearings, leading to wiping and loss of the evidence. Failure of journal bearings is much less likely, so that, in the case of wiping failures with thrust bearings, it always pays to look at the associated journal bearings to see if there is any evidence of deposits.

6.6 Sulphur Attack of Copper-Based Bearing Alloy

Sulphur attack of lead-bronze tiltingpad journal bearing

Main Characteristics

Black deposit in loaded area of pad, suggesting chemical attack. Diagnosis depends on chemical identification of the deposit.

Cause

Deposit was identified by x-ray diffraction as copper sulphide. Copper alloys are more prone to attack that tin-rich white metal. In this case the attack has come from the presence of the sulphur-containing load-carrying additive zinc dialkyl dithiophosphate commonly used in High Duty (HD) Hydraulic Oils.

Note

This example comes from a high speed centrifugal pump with an integral gear. An HD Hydraulic Oil was used to provide additional protection for the gears. These were turbine quality gears that can operate perfectly satisfactorily with oils without load-carrying additives. A change to a turbine oil solved the problem. Note this bearing had not actually failed. The machine was shut down because of failure of the thrust bearing (see next example Sulphur Attack of Silver Bearing).

Possible Confusion with other Types of Damage Comment

Other deposits described in this Section tend to look superficially the same. The final diagnosis requires chemical analysis Extreme pressure (EP) and High Duty hydraulic oils that contain active sulphur additives should be avoided in systems where there are copperbased bearing alloys.

6.7 Sulphur Attack of Silver Bearing

Silver sulphide deposit on thrust pads

Main Characteristics

Black deposit on trailing edge of thrust pads, suggesting chemical attack. Diagnosis depends on chemical identification of the deposit.

Cause

The deposit was identified by x-ray diffraction as silver sulphide. Silver is even more prone to attack than copper. These pads come from a heavily-loaded, high speed thrust bearing. Silver was chosen for the bearing material because of the expected high bearing temperature. (Note the oil feed groove at the inlet edge of the pad, giving individual pad lubrication to assist cooling.} In this case the attack has come from the presence of the sulphur-containing the load-carrying additive zinc dialkyl dithiophosphate commonly used in High Duty (HD) Hydraulic Oils.

Note

This example comes from a high speed centrifugal pump with an integral gear. A HD Hydraulic Oil was used to provide additional protection for the gears. These were turbine quality gears that can operate perfectly satisfactorily with oils without load-carrying additives. A change to a turbine oil solved the problem. (This was the same machine that provided the previous example Sulphur Attack of CopperBased Bearing Alloy that involved the journal bearing. Note that the pads are showing signs of wiping at the trailing edge where the deposit has effectively reduced the load-carrying capacity.

Possible Confusion with other Types of Damage Comment

Other deposits described in this Section tend to look superficially the same. The final diagnosis requires chemical analysis Extreme pressure (EP) and High Duty hydraulic oils that contain active sulphur additives should be avoided in systems where there are silver bearing alloys. Silver can be used for the bearing surface on the swash plate of multiple piston hydraulic pumps.

CHEMICAL EFFECTS: SUMMARY Characteristics Chemical effects can result in changes to the profile of the bearing surface, either by corrosion of the bearing material or by the formation of deposits on the bearing surfaces. Attack may also occur on the journal or thrust collar and these should be examined when the damage to the bearing has been so severe that the evidence for the cause of the failure has been removed. Possible Causes

Chemical attack can be caused by the use of unsuitable lubricants containing active chemical additives (e.g. anti-wear and extreme pressure load-carrying additives) or by contamination from external sources, most frequently the process material being handled. Actions Diagnosis of chemical effects depend identification of the reactants or the deposits forming on the bearing surfaces. Where chemical effects are suspected or where there is no clear evidence for the cause of a failure, a careful examination should be made of the other bearing surfaces that have not failed. Useful Analytical Techniques X-ray diffraction analysis can be used for the specific identification of inorganic deposits, provided the deposit has a crystalline structure. Where the deposit is non-crystalline, ESCA may provide useful evidence by identifying the chemical elements present. Identification of organic deposits is much more difficult. Infrared spectroscopy can be used to identify hydrocarbon mineral oil degradation deposits, which are also usually soluble in aromatic solvents, but the identification of other organic deposits require more sophisticated techniques such as mass spectroscopy or nuclear magnetic resonance (NMR) and the services of a skilled analyst. Again, where there is any suspicion of chemical effects, the lubricating oil should be analysed for contaminants.

PLAIN BEARING FAILURES 7. Thermal Ratcheting of Tin-Rich White Metals and Thermal Fatigue Tin has an anisotropic crystal structure; this results in different physical properties in different directions in the crystals. The crystal size in tin-rich white metals is large (ca. 5 mm) and in thrust pads subject to thermal cycling, the difference in the coefficient of thermal expansion can cause angular roughening corresponding to the grain size of the tin-rich crystals in the white metal.. Thermal ratcheting does not occur with lead-rich white metals. The differential thermal coefficient of expansion between white metals and steel can result in thermal fatigue if thermal cycling takes place. This is most commonly found on thrust pads with dovetailed joints where differences in the thermal conductivities of the two metals are also a factor. Thermal Ratcheting of Tin-Rich White Metal Roughening of tin-rich white metal on trailing edge of thrust pads caused by thermal cycling Points to Note Main Characteristics

Highlighting of tin-rich crystals in tin-rich white metal on the trailing edges of the pads where the temperature is highest.

Cause

Thermal cycling causing differential rotation of the anisotropic tin-rich crystals in the white metal. It occurs at the trailing edges of the pads where thermal changes are greatest.

Note

Possible Confusion with Other Types of Damage

Comment

The effect is produced by differences in optical reflection by differences in rotation of the different crystals. The amount of rotation is small and does not have a significant effect on performance.

Thermal Fatigue of White Metal Thermal fatigue of white metal thrust pad. Note the absence of a tinned layer on the steel backing Thermal fatigue of white metal thrust pad. Note how the fatigue lines up with the dovetail Points to Note Main Characteristics

The Photographs show typical fatigue of white metal. Normally, however, there is insufficient axial vibration in industrial machines to cause fatigue failure of thrust bearings. This is much more likely to be caused by thermal cycling.

Cause

Thermal cycling can lead to cracking because of the difference in the thermal coefficient of expansion between steel and white metal. Differences in thermal conductivity may also be a factor, particularly where dovetails are used because the different thermal paths accentuate the temperature difference between the white metal and the steel. Note the way the fatigue damage lines up with the dovetail in Photograph ‘b’. Poor bonding accentuates the situation. The absence of a tinned layer on the exposed steel in Photograph ‘a’ shows poor bonding. The use of dovetails tends to make it more difficult to achieve a sound metallurgical bond between the white metal and the steel backing..

Note

Notice particularly the coincidence in location of the fatigue with the dovetails in the steel backing. Possible Confusion with Other Types of Damage Normal fatigue (see Section 3) can occur in thrust bearings, but is very uncommon.

Comment

Dovetails should not be used in white metals bearings. The differences in the thermal paths tend to disrupt the bond when cooling after pouring on the molten white metal because of the different coefficients of thermal expansion between the two metals. (See also Section 3: Premature Fatigue)

Thermal Ratcheting and Thermal Fatigue: Summary Characteristics Thermal ratcheting of tin-rich white metal produces a chequer board appearance on the trailing edges of thrust pads subject to thermal cycling. The effect is similar to that produced by chemical etching used for metallographic examination. It is largely an optical effect resulting from variations in the light reflected from slight differences in the orientation of neighbouring crystals. The scale of the pattern shows the size of the tin-rich crystals. The effect is limited to tin-rich white metals. Possible Causes Thermal ratcheting arises from thermal cycling of the bearing metal and, when this occurs, the effect is most obvious on the trailing edges of thrust pads where temperature changes are greatest. It occurs because of differential rotations of the tin-rich crystals caused by differences in the thermal coefficient of expansion in different directions in the tin-rich crystals. Where dovetails are used thermal cycling can lead to thermal fatigue because of the difference in thermal expansion of white metal and steel, particularly where the use of dovetails results in to poor bonding of the white metal to the backing (see Section 3). Actions No corrective action is normally required for thermal ratcheting, though the very slight damage can easily be removed by a light polish with abrasive paper. Excessive temperature rise can, of course, result in wiping failure as described in Section 1. Dovetails can inhibit the production of a sound metallurgical bond between steel and white metal and should not be used. See Section 3: Premature Fatigue.

8.Electrical Damage Both electromagnetic and electrostatic voltages generated on rotors can discharge through the thin oil films in bearings and cause electrical erosion damage. This produces a characteristic matt finish that is frequently replicated on the journal or thrust collar. Electrical Erosion Damage in Journal Bearing (1) Electrical arcing damage in loaded area of journal bearing Electrical arcing damage to the journal corresponding to Photograph ‘a’ Points to Note Main Characteristics

This damage occurred in an electric motor after a short run on test following a mechanical overhaul. Damage is confined to the loaded area on the bottom half bearing, but is continuous on the journal. Note the shiny melt pits on the surface of the bearing

Cause

Electrical fault with leakage to earth through bearing oil film.

Note

Possible Confusion with Other Types of Damage Dirt erosion (see Section 2: Erosion of White Metal Journal Bearing) and cavitation erosion (see Section 4) give a superficially similar matt finish on the bearing surface, but both of these have an arrow shape pointing against the direction of motion. Moreover, there is no corresponding damage on the journal

Comment Electrical Erosion Damage in Journal Bearing (2) Spark erosion damage on top half of ring-oil bearing Points to Note Main Characteristics

Note the characteristic roughening at the leading edge of the top half bearing. In this case so much damage had been caused to the bottom half that it had wiped and the evidence of electrical damage had been removed.

Cause

Electrical fault in motor causing sparking through oil film at area of minimum film thickness.

Note

Evidence of electric arcing was also present on journal. Possible Confusion with Other Types of Damage Dirt erosion (see Section 2: Erosion of White Metal Journal Bearing) and cavitation erosion (see Section 4) give a superficially similar matt finish on the bearing surface, but both of these have an arrow shape pointing against the direction of motion. Moreover, there is no corresponding damage on the journal

Comment

This was the third successive failure to the motor after an overhaul. The wiping in

the first two was attributed to lubricant breakdown caused by dirt in the oil; this was also considered responsible for the damage to the journal. A more careful examination after the third failure revealed the erosion damage on the top half bearing and that the damage to the journal was pitting, not scoring.

Electrical Arcing Damage on Journal of Vertical Electric Motor Electrical arcing damage on the journal of a vertical electric motor Points to Note Main Characteristics

This is an unusual type of electric damage, similar in appearance to ‘washboarding’ in rolling bearings.

Cause

The damage shown in the Photograph occurred on a 4 MW vertical motor on a standby water pump. The motor was run each month for about an hour. This pattern of operation lead to the laying down of wear dust from the carbon brushes that in time caused an explosion that was traced to arcing in the end cover. The damage to the journal was found when the machine was stripped to investigate the explosion. Similar damage was found on other motors in the pumping station that were subject to the same operating regime.

Note

The direction of motion in the Photograph is horizontal. Note that most of the lines of marks follow the direction of motion, some are at an angle of about 45°. Possible Confusion with Other Types of Damage

Comment

No explanation was found to account for the effect, but it is worth illustrating as an example of electrical damage.

Electrostatic Erosion Damage on Journal Bearing Electrostatic erosion damage to bottom half bearing from centrifugal compressor Journal corresponding to the bearing shown in Photograph ‘a’ Points to Note Main Typical ‘frosting’ appearance of electrostatic arcing damage. This tends to be on a Characteristics finer scale than electromagnetic current erosion, but at low magnification (5-10X), the characteristic shiny pits can be seen Cause Electrostatic built up on rotor of centrifugal compressor caused by droplet impingement on the wheels. Note

Note that the damage to the bearing is limited to the region of low oil film thickness, the loaded region, of the bottom half bearing, but is continuous on the journal.

Possible Confusion with Other Types of Damage Dirt erosion (see Section 2: Erosion of White Metal Journal Bearing) and cavitation erosion (see Section 4) give a superficially similar matt finish on the bearing surface, but both of these have an arrow shape pointing against the direction of motion. Moreover, there is no corresponding damage on the journal Comment

Electrostatic build up can result from either solid particle or liquid droplet impingement on the rotor. A build up of about 20V at least is necessary for damage to be caused.

Electrostatic Erosion Damage on Thrust Pad Electrostatic erosion damage on thrust pad

Points to Note Main Characteristics

Note the characteristic ‘frosting’, that is caused by spark erosion pitting. The shiny melt pits are obvious at low magnification (5-10X)

Cause

Electrostatic built up on rotor of centrifugal compressor caused by droplet impingement on the wheels.

Note

Note that the damage to the bearing is limited to the outer edge of one pad. This suggests that this particular pad was misaligned and proud of the other pads. Possible Confusion with other Types of Damage

Comment

Electrostatic build up can result from either solid particle or liquid droplet impingement on the rotor. A build up of about 20V at least is necessary for damage to be caused.

Electrical Damage: Summary Characteristics Both the passage of electric current and electrostatic discharge across a bearing oil film produces a characteristic matt appearance on the bearing surface (sometimes referred to as ‘frosting’}and this damage is frequently replicated on the opposing steel surface(journal or thrust collar). As discharge occurs at the point of minimum film thickness, the damage is confined to the loaded region of the bearing, but is continuous on the opposing surface as this rotates and passes through the minimum film thickness. Examination at low magnification (5-10X) reveals shiny, rounded pits from which metal has been removed by melting. The bearing surface is gradually degraded and, when it can no longer support a load-carrying oil film, breakdown occurs and a wiping failure follows (see Section 1). This is a secondary effect, but the problem is that it can remove the evidence of electrical discharge, though in the case of journal bearings some signs of this may remain in the unloaded half of the bearing and evidence of discharge should be found on the opposing steel surface.

Possible Causes Electromagnetic damage is most frequently caused by leakage currents in elevctric motors, though it can also occur in large rotary machines (compressors, turbines) where the rotor has been handled with magnetic grabs. A potential of at least 250 mV is necessary for damage to be caused. Electrostatic damage is most likely in large steam turbines and high-speed centrifugal compressors where particles are or droplets of liquid are present in the gas stream. It also occurs in compressors subject to on-line washing by the periodic injection of solvents. A potential built up of about 20V is required for damage to occur. The diagram opposite shows the principles sources of electrical discharge that can cause damage to bearings. Actions Electric motors should be provided with insulated bearing housings and fitted with earthing straps unless shaft voltages can be guaranteed to be less than 250mV. The conditions required for the generation of electrostatic potential on a rotor - particle/droplet size and velocity - are difficult to define precisely. Thus, where there is a risk of liquid droplets or solid particles being present in the gas stream in turbines and compressors, the shaft should be grounded by earthing brushes if the bearing oil film can sustain a potential of 20V or greater. Principal Sources of Electrical Discharges that Can Cause Damage to Bearings

9. Incompatible Journal or Thrust Collar Material The design philosophy with plain bearings is that, should failure occur, the shaft or thrust collar should be unharmed with any damage limited to the bearing. This imposes little restriction on the choice of the shaft or thrust collar, which is normally of steel, apart from the general rule that they should have a hardness at least three times that of the bearing material. This means that mild steel is perfectly satisfactory with white metal bearings, though hardened steels are required for harder bearing alloys, such as the lead-bronzes. Again, this poses no particular problems, though nitrided steels must be cleaned by lapping before use, otherwise they are liable to cause bearing damage. The one exception is that steels containing more than 1.5% chromium should not be used where peripheral speeds exceed 25 m/sec to avoid the risk of a destructive failure mechanism known as wire-wool failure (black-scab failure). The diagram below shows an analysis made in 1966 of reported wire-wool failures. These were shown to occur with to steels having chromium contents in the range 3-18% in bearings operating with surface speeds greater than 80 ft/sec (25 m/sec). Failure of Bearings with Nitrided Steel Shaft

Bearing that had been operated with a nitrided shaft that had not been surface lapped after nitriding Points to Note Main Characteristics

The bearing damage is in the form of sharp vee-shaped grooves pointing against the direction of rotation.

Cause

The presence of hard particles loosely adhering to the nitrided surface.

Note

Possible Confusion with Other Types of Damage Both dirt erosion (see Section 2: Wear) and cavitation erosion (see Section 4: Cavitation Erosion) produce arrow-shaped damage on the bearing surface; in these cases, however, the whole of the arrowhead has a matt appearance.

Comment

Nitrided steels should be cleaned and polished before being used as bearing surfaces.

Wire-Wool Failure of Journal and Thrust Bearing Wire-wool failure of journal bearing Wire-wool failure: damage to journal and thrust collar corresponding to the bearing in Photograph ‘a’ above. Points to Note Main Characteristics

Identified by extensive damage to both bearing and the journal and/or thrust collar with wear products from the shaft collected in the bearing housing where they look like wire wool. A black scab is also often present on the damaged surface, hence the alternative name ‘black scab failure’.

Cause

This failure occurs when a small particle of hard dirt or swarf gets embedded in the bearing material, but continues to rub against the steel counterface. At high speed the temperature generated by the frictional rub carburises the chromium in the steel in the presence of a hydrocarbon oil, producing hard chromium carbides that embed in the soft bearing material and act as cutting tools on the journal or thrust collar. The process continues by accretion of the embedded carbides and may result in a journal being turned completely through.

Note

Although wire-wool failures have been most frequently encountered with whitte metals bearings, they can also occur with copper bearing alloys. Possible Confusion with other Types of Damage

Comment

The best solution is to limit the chromium content of the steel to 1.5% where surface speeds are greater than 25 m/sec.

Unsuitable Journal or Thrust Collar Materials Summary There are two quite separate types of bearing damage that arise from using unsuitable materials for the journal or thrust collar. Damage with Nitrided Steel Shafts Characteristics The bearing surface of a nitrided steel shaft has to be cleaned and polished before going into service otherwise there is a risk of scoring of the bearing, with a characteristic vee-shape pointing against the direction of rotation Cause Failure to clean the surface after nitriding. This is in fact a manufacturing error (see Section 10), though for descriptive purposes it is most conveniently treated here. Wire-Wool Failure Characteristics This is the only form of bearing failure that results in extensive damage to both bearing surfaces. It is readily identified by the presence of a mass of ‘wire wool’ in the bearing housing. Cause Embedding of a hard particle of dirt or swarf in the bearing material that remains proud and rubs against the steel surface creating high temperature through friction. Wire-wool failure is limited to alloy steels containing more than 3% chromium operating with surface speeds greater than 25 m/sec. Actions The best action is to avoid using steels containing more than 1.5% chromium in bearings with surface speeds greater than 25 m/sec. The optimum solution with existing machines is either to

sleeve the journal and replace the thrust collar with a plain low carbon steel; though hard chrome plating of the bearing surfaces will also remove the risk. Careful control has to be exercised over the chromium plating to ensure that the chromium is properly bonded to the steel surface.

10. Manufacturing Faults and Assembly Errors Bearing failures can occur as a result of manufacturing or assembly errors. A few examples are given, but obviously there is no general pattern as the failure mechanism depends on the type of error. In the event of unusual effects that do not match with any of those illustrated in the previous Sections, a careful examination should be made of the whole bearing assembly for any manufacturing defects or assembly error that could account for the damage. This should include the back of the bearing or bearing shell for peculiar markings and a check of the journal for taper, barrelling or corrosion damage.. There are two obvious categories: manufacturing faults

assembly errors

Manufacturing Faults 1. Barrelled Shaft Points to Note Main Characteristics Cause Note

Possible Confusion with Other Types of Damage

Comment Manufacturing Faults 2. Undersize Bearing Shells or Oversize Housing Bore Points to Note Main Characteristics Cause Note Comment Manufacturing Errors

Possible Confusion with Other Types of Damage

3. Incorrect drilling of oil feed hole Fatigue damage to journal bearing (a) in line with incorrectly positioned feed hole partially drilled in back of shell (b) Points to Note Main Characteristics

A small patch of fatigue damage at centre line of bottom half bearing, Photograph ‘a’. Fatigue would be expected to occur at the position of maximum, not at the bottom of the bearing

Cause

Examination of the back of the bearing shell shows a partially drilled hole in line with the fatigue damage, Photograph ‘b’. This was obviously a machining error that was discovered before the oil feed hole was completely drilled through. This has resulted in inadequate support of the white metal that has failed in fatigue under the low loading at the centre of the bottom half of the bearing.

Note

This was a main bearing from a reciprocating compressor, with sufficient alternating loading to cause the fatigue. Possible Confusion with Other Types of Damage While this is clearly a fatigue failure, the location of the fatigue patch suggests some abnormality.

Comment

The back of the bearing shell should always be examined, particularly in cases of abnormal damage.

Manufacturing Errors 4. Incorrectly Inserted Thermocouple Causing Distortion of Bearing Surface Points to Note Main Characteristics

This bearing was removed because of a very high temperature when it was put into service.

Cause

The surface of the pad was distorted by incorrect assembly of the thermocouple. This led to reduction in the oil film thickness in the distorted area with abnormally high temperature resulting from the thin oil film.

Note

It is normal to fit more than one pad with a thermocouple or RTD. This should help to indicate if one of the pads is behaving abnormally. Possible Confusion with Other Types of Damage

Comment

Assembly Errors 1. Misalignment Pair of bearing shells showing wiping at opposite sides of bottom and top shell Points to Note Main Characteristics

The wiping at opposite sides of the top and bottom halves of the bearing shows misalignment causing breakdown of the oil film by edge loading.

Cause

Misalignment

Note

A similar result could occur by running with a tapered shaft, but in this case the damage would be on the same side of each bearing half. Possible Confusion with Other Types of Damage

Comment

The low melting point and strength of white metal (compatibility) are valuable properties, allowing it to wipe or deform to compensate for a limited amount of misalignment.

Assembly Errors 2. Dirt Between Thin Shell Bearings and Housing A thin-walled bearing shell showing local overheating in its bore caused by dirt trapped between the shell and the housing Points to Note Main Characteristics Cause Note Comment

Possible Confusion with Other Types of Damage

Mechanical Engineering Solutions

ROLLING BEARING PROBLEMS & FAILURES INTRODUCTION Examination of failed components is a crucial element in failure diagnosis and is frequently the starting point in any failure investigation. The illustrations that follow have been specifically selected to highlight particular mechanisms of failure. It should not be overlooked, however, that the initial cause may lead to secondary damage that obscures the original event. Care and experience may be needed to separate primary and secondary effects. It is important to realize that rolling bearings have a finite life, the repeated high stressing at the contacts between rolling elements and tracks eventually resulting in fatigue damage to the surfaces. There are, however, a large number of other reasons why rolling bearings fail and it is crucial to distinguish between these and normal fatigue when investigating failures. Following are the main mechanisms of bearing damage and failure, but examination of used bearings that have not failed can still give valuable information on the conditions under which the bearing has been operating. This forms the first section. 1. Track markings 2. Fatigue Failure 3. Premature Fatigue 4. Fitting Damage 5. Lubrication Failures 6. Loss of Internal Clearance 7. Wear 8. Fretting Damage 9. Manufacturing Defects 10. Electrical Damage

1. Track markings Even if fatigue failure has not occurred, rolling bearings usually show track marks on the races in the regions where the rolling elements have been in contact. The shape and position of these track marks can give useful guidance on the conditions under which the bearing has been operating (figure 1a).

Figure 1a Tracking patterns on a ball bearing races arising from a variety of operating situations

Figure 1b is a photograph of the tracking pattern on a misaligned ball bearing with associated ball pocket wiping on the cage.

2. Fatigue Failure The life of a rolling bearing is defined as the time the first detectable fatigue pit forms on any of the contacting surfaces (figures 2a, b). This is a somewhat random effect, determined by metallurgical imperfections in the steel and, while modern steel-making techniques have produced steels that are significantly cleaner and have more uniform structures, crystal lattice imperfections or other

defects still provide the loci for fatigue. This can occur on either the races or the rolling elements, depending on the design of the bearing, the direction of loading and variations in steel quality.

Figure 2a

Figure 2b

Once a fatigue pit has formed, it sets up a point of weakness and fatigue damage progresses round the component (figures 2c, d). In this condition, although the bearing may have become noisy, it is still performing its function of guiding the shaft and cannot really be said to have failed in the conventional usage of the term. Collapse of the bearing occurs when the component fractures from a stress raiser set up by the fatigue pits (figure 2e). Practical experience suggests that once the first fatigue pits are formed the bearing has only used up 70-80% of its life to collapse.

Figure 2c

Figure 2e

Figure 2d

3. Premature Fatigue 3a. Misalignment Misalignment of roller bearings causes edge loading and overstressing of the raceway and can result in premature fatigue. This is easily recognised as the damage is clearly restricted to one side (figures 3.1a, b).

Figure 3.1a Edge pitting of the inner race of a taper roller bearing due to misalignment

Figure 3.1b Edge pitting of a cylindrical roller bearing due to misalignment

3b. Incorrect fitting methods Pressing the inner race on to a shaft by applying force to the outer race will cause indentations at the rolling element contacts by plastic deformation (figure 3.2a). Alternatively, misalignment when offering up the races of a separable cylindrical roller bearing can cause scoring at the point of contact of the roller ends and the race (figure 3.2b).

Figure 3.2a Fitting impact damage on angular contact

Figure 3.2b

Both these mechanisms result in damage to the bearing surface that acts as stress raisers and reduces the fatigue life. 3c. Corrosion when stationary Corrosion can occur when the bearing is stationary and is not adequately protected by the lubricant. For example, if a grease-packed bearing is installed in a machine and left standing for some time without running, the grease is not spread over all the surfaces and distributed so that it forms a good seal in the end covers. In this event moisture can condense on the bearing surfaces. Corrosion occurs at the narrowing contact between rolling element and race by a mechanism known as crevice attack (figure 3.3). Again this produces points of weakness that act as initiators for fatigue.

Figure 3.3 3d. False brinnelling If a bearing is subject to vibration when stationary, fretting damage occurs at the points of contact, giving rise to what is known as ‘false brinnelling’ (figures 3.4a , b). (Note: It is different from fitting damage in that material is removed rather than displaced.) This can occur to large bearings even before they are installed. With separable bearings, e.g. cylindrical roller, taper roller bearings, it is good practice to store the bearing with the two parts separated. Non-separable bearings should be stored with the axis of rotation vertical so that the gravity loading is uniformly distributed over the rolling elements.

Figure 3.4a

Figure 3.4b

Standby machines and machines in store should be rotated slightly once per week to change the points of contact. Where there are installed standby machines, it is good practice to run the machines alternately to reduce the risk of fretting. This is particularly important in the case of two machines mounted on the same pedestal. A characteristic of all the above mechanisms is that the damage occurs when the bearing is stationary and is thus at rolling element spacing. Premature fatigue is readily distinguished from normal fatigue in that it occurs in discrete patches at rolling element spacing, rather than as a continuous band starting from the initial point of failure as is the case with normal fatigue (figures 3a, b).

Figure 3a

Figure 3b

Premature fatigue also occurs with angular contact bearings incorrectly fitted so that the axial load is taken on the low shoulder side of the race (figure 3c).

Figure 3c

4. Fitting Damage Excessive impacts when fitting a bearing on to a shaft during assembly can lead to indentations on the tracks, figure 3.2a (similar to false brinnelling, figure 3.4b, but in this case there will be a raised lip surrounding the indentation) or to fracture of the race shoulders (figures 4a, b).

Figure 4a A spherical roller bearing damaged by the application of excessive impacts during assembly

Figure 4b A cylindrical roller bearing fractured by careless fitting

Excessive interference fit can even result in fracture of the race (figure 4c).

Figure 4c 5. Lubrication Failures Rolling bearings have two separation lubrication requirements: a lubricant film to protect the surfaces of the cage where sliding contact occurs, and an elastohydrodynamic lubricating film that forms at the rolling contacts. Breakdown of either of these mechanisms give rise to characteristic types of damage. 5a. Sliding contacts

In normal operation the loading at the sliding contacts between rolling elements and cage, or between race and cage with a race-centred cage, is low and lubrication requirements are modest. Wear occurs, however, if there is insufficient oil or the grease becomes hard through oxidation or loss of oil. Wear is normally confined to the softer cage (figures 5.1a, b). While in normal circumstances the loading at the cage contacts points is low, this changes, however, if the bearing is misaligned. The overloading caused by misalignment is a common cause of cage wear and failure. The evidence of misalignment can frequently be found from track markings on the races (see Track Markings).

Figure 5.1a Wear of softer cage through loss of lubrication

Figure 5.1b Cage wear

5b. Rolling contacts There is a need to develop a satisfactory lubricant film at the rolling contacts; in the majority of cases this will be an elastohydrodynamic film, but in very heavily-loaded bearings it can be a film developed by reaction of extreme pressure additives in the oil. Failure to develop a satisfactory lubricant film gives rise to a smearing type of damage that is usually referred to as ‘surface distress (figure 5.2).

Figure 5.2 ‘Surface distress’ in deep groove ball bearing caused by breakdown of elastohydrodynamic lubrication film

A characteristic of both types of lubrication failure is that the temperature rises through increased friction loss in the bearing. This may allow a breakdown in lubrication to be detected before failure occurs, though if a lot of heat is generated sufficient thermal gradient may develop between inner and outer race (for example, on a rotating shaft, the housing usually provides a better heat sink for the outer than the shaft does for the inner race so that the latter reaches a higher temperature) to result in loss of internal bearing clearance, giving rise to a different failure mechanism (‘Loss of Internal Clearance’). 6. Loss of Internal Clearance The optimum condition for a rolling bearing is when the internal radial clearance is zero. It is possible to run bearings with pre-load so that there is no residual internal clearance and rolling occurs by elastic deformation at the contacts; this is, however, a somewhat sensitive condition with higher friction losses and the risk of differential thermal expansion between the races loading the contacts beyond the elastic limit and preventing rolling. Pre-loaded bearings are normally only used where very accurate guidance of the shaft is required (e.g. in machine tools) and more normal practice is to run with a small residual internal clearance to ensure that rolling is maintained at the contacts. Once the clearance has been lost and the elastic limit exceeded, sliding occurs at the contacts. There is a large increase in friction and the process escalates, giving a condition of thermal runaway. Either the machine stalls if there is not sufficient power to overcome the friction, or the bearing heats up so much that it deforms plastically (figures 6a, b). With a bearing mounted on a rotating shaft, the inner race reaches the highest temperature and a characteristic of this type of failure is that the severity of damage is greatest at the inner race.

Figure 6a Loss of internal clearance Figure 6b Loss of internal clearance failure failure

This is a serious type of failure that can lead to catastrophic damage if sufficient power is available to continue driving the machine (figure 6c). Moreover, the failure occurs so suddenly that it is not possible to provide a monitor that will give adequate warning to allow protective action to be taken. This is a serious type of failure that can lead to catastrophic damage if sufficient power is available to continue driving the machine (figure 6c). Moreover, the failure occurs so suddenly that it is not possible to provide a monitor that will give adequate warning to allow protective action to be taken.

Figure 6c Damage to motor and pump following loss of clearance failure in pump drive-end bearing The following mechanisms can lead to loss of clearance failure. 1. Transient temperature differential during start-up or acceleration The shaft generally heats up faster than the bearing housing so that the temperature differential during the start-up can be much greater than the steady state value. This is particularly severe with hollow shafts or very thick housings. 2. Use of bearing with incorrect internal clearance class for the application Where in normal circumstances there is a difference in temperature between shaft and housing (e.g. in centrifugal pumps handling hot liquids, hot gas fans, electric motors) a higher than normal clearance class of bearing should be used. This is particularly the case with chemical pumps using austenitic steel shafts that have a higher coefficient of thermal expansion than the ferritic steel of the bearing. In such cases a high clearance bearing should be used (e.g. C3) and, with hot fluids, C4. Particular care has to be taken during maintenance that the correct clearance class of replacement bearing is fitted. 3. Excessive interference fit(s) of races About 80% of the interference fit is transmitted through the race. This is allowed for in the fits recommended by the bearing manufacturers.

4. Insufficient fit of inner race on shaft Rotation of the bearing on its seating leads to increased friction and heating of the inner race and hence loss of clearance. 5. Complete loss of lubrication Loss of lubrication leads to increased friction and again to increased temperature differential. 6. Inadequate lubricant viscosity to allow generation of elastohydrodynamic lubricant film This can occur with oil-lubricated double-row bearings and bearings mounted in pairs where there is insufficient circulation of the oil to the outer row or race to prevent it rising to a temperature where its viscosity becomes too low. In such cases the loss of clearance occurs in the outer bearing or row. A positive supply of lubricant to the outer bearing or race is required when dmn>3500 mm.rev/s (dm = mean bearing diameter). 7. Gross overload resulting in excessive frictional heating A typical example of this is where two locating-type bearings are mounted on the same shaft with one mounted ‘free’ in its housing to allow for differential axial thermal expansion between shaft and housing. If the ‘free’ bearing fails to slide during a thermal change, e.g. on starting or stopping, a gross overload is created when the bearings have to resist the expansion (contraction) of the shaft. In practice this effect only becomes significant for bearings with outer diameter >100mm. 8. Use of cooling jacket round bearing Where, because of the environment, cooling is required to control the bearing temperature, this should not be in the form of a complete water jacket round the bearing, which only serves to increase the temperature gradient between inner and outer races. Where cooling is necessary it should be applied to the oil, not to the bearing. Where bearing housings are fitted with cooling jackets, these should be filled with oil or glycol to allow convection cooling. The jacket should be fitted with a stand pipe to allow for thermal expansion of the liquid. 7. Wear Abrasive particles in the lubricant will cause wear. This is normally most severe with the softer cage material and can lead to cage fracture. Wear damage can be distinguished from lubrication breakdown by the appearance of the tracks. Either these take on a specular polish if the abrasive particles are very fine, or become matt and scored with larger particles (figure 7a).

Figure 7a Matt track caused by wear with abrasive particles Wear, of course, can be caused by the particles released by fatigue pitting, but this is secondary to the fatigue damage. Care should be taken to ensure that, when electric induction heaters are used to ease the fitting of races with an interference fit, the bearing is properly demagnetised. A magnetised bearing tends to collect steel particles and this will increase the amount of wear. Wear also occurs as a secondary result of corrosion, the iron oxides so produced acting as jeweller’s rouge (figure 7b). Under extreme conditions with high speed bearings this can lead to the formation of hollow balls (figure 7c) where there is some internal discontinuity in the ball that allows it to expand to fit the gap between the races to compensate for the wear.

Figure 7b Heavily corroded ball 8. Fretting Damage

Figure 7c Hollow ball

Fretting damage can occur at the contact between the bearing and its seating / housing, particularly where a sliding fit is used, because of small oscillating movements that take place there as the rolling elements pass through the loaded zone (figure 8a). This does not normally give rise to failures, though if it is very severe it can form stress raisers that will initiate fatigue fracture of the race (figure 8b); this differs from the cracking shown in figure 4c by the absence of fretting marks where the cracking has been caused by excessive interference fit. Fretting can also cause sticking of a ‘free mounted’ bearing as discussed in 7, above. Where fretting is a problem it can be prevented by applying a thin electro-deposited coating of indium to the surface or the bore or outer diameter before fitting.

Figure 8a Fretting on a bearing outer race due to inadequate interference fit and local distortion of housing.

Figure 8b Fretting of a bearing inner race due to inadequate interference fit, leading to fracture of the race.

9. Manufacturing Defects Rolling bearings are manufactured in a highly competitive commercial market and are subject to rigorous inspection procedures. Nevertheless, the odd substandard bearing may get through and, although this is very uncommon, should not be discounted as a possible cause of failure. The following cases have been encountered. 1. Wrong material

One example concerns a single-row angular contact bearing in which, on removal from service, one of the balls was found to be scored, whereas all the others were in the as-new condition. Examination showed that the offending ball was of 13% Cr steel, not the standard 1% Cr, 1% C steel. This was by a chance examination; if this had not taken place, a premature failure would have occurred. 2. Wrong heat treatment Races and rolling elements of normal rolling bearings are tempered at about 140°C to give a final hardness of 600-650 HV (62-63 HRc). Some large taper roller bearings may be manufactured of case-hardened steel. A soft bearing will fail prematurely. Incorrect heat treatment can be suspected if any of the components are soft and do not exhibit heat discolouration from the failure. 3. Incorrect machining Very rarely bearings may be found to be outside the standardised external dimensions. This is usually obvious during fitting and should not cause a problem. A more serious error, however, was in a double-row angular contact bearing with a filling slot that had been machined right into the ball track (figure 9a).

Figure 9a Inner race of double-row angular contact bearing with incorrectly machined filling slot Figure 9b shows comparative Talyrond tracings from the race of a normal quality bearing and the one that was removed because of excessive noise. This would probably have suffered a lubricant breakdown failure if it had not been removed because of the noise.

(a) Poor quality finish (magnification x2500)

(b) normal quality finish (magnification x5000)

Figure 9b Talyrond traces of track of inner race 10. Electrical Damage Passage of an electrical current through a rolling bearing causes surface damage and hence decreased fatigue life. With currents above 0.5 amp a characteristic ripple or ‘washboard’ pattern is formed on the contacting surfaces (figure 10a), but even with currents as low as 0.01 amp there is still some deterioration, which although not visible results in a decrease of life as shown in table 1. Note that the ‘washboarding’ in figure 10a is very severe, and milder forms of washboarding occur, some of which can only be detected by rotating the race while viewing in good lighting. The wavelength of the washboard pattern is highly variable, and shows no obvious correlation with features such as roller spacing, electrical frequency etc.

Figure 10a Electrical ‘washboard’ damage on outer race of ball bearing

Current

Comments

0.001 amp

no effect on life

0.01 amp

ca. 20% reduction in life expectancy; failurehas normal fatigue characteristics

0.01-0.1 amp

life expectancy reduced by 20-80% over thisrange; failure has still normal fatiguecharacteristics

0.5 amp

life expectancy reduced by more than 80%;‘washboard’ damage occurs

Table 1 Effect of electric currents on the life of rolling bearings From a practical point of view, the impressed voltage should not exceed 0.3 V on ball bearings, 0.5 V on rolling bearings. Electric discharge damage does not, however, always result in ‘washboarding’. Figure 10b shows pitting damage caused by high voltage discharge. This is similar to corrosion pitting, figure 7b, but is distinguished by the absence of oxide and the rounded pits that show evidence of melting.

Figure 10b Electrical pitting damage on ball from ball Figures 10c and d show arcing damage caused by electrostatic discharge.

Figure 10c Note corresponding pattern on race and on ball, suggesting single large discharge.

Figure 10d

SEAL FAILURES The failure mechanisms of mechanical seals are covered extensively by B J Woodley in "Mechanical Seal Practice for Improved Performance" (J D Summers-Smith (ed.) 2nd edition 1992). It is not proposed to cover the same ground here, though it is perhaps worth mentioning that failures of mechanical seals are frequently catastrophic with the resulting damage making it difficult to distinguish the original cause of failure. Most failures are the result of operating faults, e.g. failure of external cooling, flush or quench resulting in breakdown of the film between the seal faces. With the seal damaged, wear continues, even with the reestablishment of the coolant supply, so that it is hard to identify that the real problem was earlier, temporary loss of coolant. In such circumstances the nature of the face materials plays an important role. With two hard faces (e.g. silicon carbide, tungsten carbide), failure is almost immediate; whereas the seal with one soft, self-lubricating face (e.g. carbon-graphite, ptfe) can continue to operate some time before failure while the soft face wears out. If failure does not manifest itself for some weeks after the event that prompted it, identification of the cause of failure is not straightforward. OIL BARRIER SEALS Floating-bush seals are commonly used on the shafts of process gas centrifugal compressors. Figure 1 shows the general arrangement, with mineral oil, normally the system lubricating oil, fed between a pair of floating bushes to about 5 m head above the pressure of the sealed gas. Despite the pressure differential, some leakage occurs across the seal (the mechanism for this is not clear, but it seems likely to be precession of the seal ring about the shaft through slight departures from roundness in the seal ring bore or the shaft causing cavitation in the oil film.). High temperature occurs in the seal oil film because the high rate of shear and negligible cooling by the leaking oil. This can lead to chemical reactions in the seal that led to failure.

Figure 1: Schematic diagram of floating-bush seal system Figure 2 shows a deposit of copper sulphide on the white metal lining of a seal ring. Figure 3 shows a failure when the deposit had taken up the clearance.

Figure 2: Black deposit of copper sulphide on white metal lining of seal ring

Figure 3: Failure of seal ring through loss of clearance caused by build up of copper sulphide deposit

This reaction occurs between the active sulphur present in ‘sour’ hydrocarbon gases and copper in the tin-rich white metal lining of the seal. Some relief can be obtained by using a copper-free lead white metal. Figure 4 gives a Guidance Chart for white metal selection based on operating experience using shaft peripheral velocity as an index of temperature. In very severe conditions, high speeds and high sulphur contents, sulphur reacts with lead forming a deposit of lead sulphide on the seal ring and an alternative seal design is required.

Zone A Tin-rich white metal acceptable Zone B Copperfree leadrich white metal acceptable Zone C Alternative seal design required

Figure 4: Guidance chart for selection of white metal lining for floating-bush seal rings for use with sulphur containing hydrocarbon process gas Reaction can also occur with active sulphur additives in the oil. Figure 5 shows a failed seal ring caused by operating with a high duty hydraulic oil containing zinc dialkyldithiophosphate (zddp).

Figure 5: Failure of floating-bush seal ring caused by build up of copper sulphide on white metal lining caused by reaction with zddp additive in the oil Similar failures can occur through loss of clearance caused by deposition of reaction products between the process gas and additives in the oil. Figure 6 shows a failure of a seal ring caused by reaction between ammonia in the gas and the succinic acid corrosion inhibitor in the oil. The ammonium succinate reaction product is insoluble in the oil and is deposited on the hot lining of the seal ring.

Figure 6: Failure of seal ring through deposit of ammonium succinate formed by reaction between ammonia in the gas and the corrosion inhibitor in the oil Rings with low clearance are used to limit the inward leakage of the barrier oil. This results in a low oil film thickness between the rotor and the casing and this is the most likely place for discharge to occur in the event of electrostatic build up on the rotor. Figure 7 shows electrostatic erosion of the white metal lining of a seal ring giving increased clearance and excessive inward leakage.

Figure 7: Electrostatic erosion of white metal lining of seal ring causing increased inward leakage of the seal oil Radial face seals are also used as oil barrier seals. Again, problems can occur because of high temperature in the seal. Figure 8 shows a seal ring that was removed because of excessive inward leakage of oil. This was caused by three equispaced deposits that separated the faces. Analysis showed the deposits to be oxidised oil. The ring was found to have a three-wave undulation, the thin film at the high spots causing excessive temperature leading to the deposition of oil oxidation products. Distortion had resulted from the relief of locked-in stresses during manufacture in a three-jaw chuck. The problem was solved by stress relieving before final machining.

Figure 8: Three equidistant deposits of oil oxidation roducts on a radial face seal Similar failures have occurred through the breakdown of oil additives at the high temperature in the seal. Figure 9 shows the presence of additive breakdown products on a radial face seal removed because of excessive leakage. The oil used in the lubrication system contained e.p. additives that broke down in the seal.

Figure 9: Deposit of e.p. additive breakdown products on radial face seal GLANDS ON RECIPROCATING RODS Lubrication of the glands on reciprocating rods is essential. If adequate lubrication is not provided, severe scoring of the rod results even with packings of self-lubricating materials. Figure 10 shows scoring on the rod of a 240 bar methanol ram pump at the point where it passed through the filled ptfe packing. The local safety authority required that there should be no leakage of methanol. Scoring occurred when the gland was tightened up to stop the leakage. The problem was solved by fitting a lantern ring at the centre of the packing and supplying a low viscosity lubricating oil.

Figure 10: Scoring on chrome-plated rod of 240 bar methanol pump as the result of overtightening the gland A similar failure occurred on a 3-row ram pump discharging light aldehydes at 250 bar using glass-fibre reinforced ptfe chevron rings to seal the rods. The pump had been specified to operate without leakage. Predictably the seals failed. Automatic packings require adequate viscosity in the pumped liquid and wetting of the rod to ensure lubrication of the rings. The viscosity of the light aldehydes was inadequate and failure was inevitable, even with self-lubricating ptfe rings.