Bearing Damage Analysis [PDF]

Zitiervorschau

BEARING DAMAGE ANALYSIS: Recognizing and preventing damage in automotive bearings can dramatically increase bearing life and decrease the potential for improper handling, installation and adjustment. It also reduces instances of bearing failure, thereby increasing the safety of vehicle passengers. The most common types of bearing damage that may result in a reduction of bearing or application life are often caused by insufficient maintenance practices, mishandling, improper adjustment practices or inadequate lubrication. The following offers a quick reference to the common causes of bearing damage in automotive applications.

Roller-end scoring Metal-to-metal contact from breakdown of lubrication film.

Cone large rib and roller large end scoring “Welding” and heat damage from metal-tometal contact.

Roller large end deformation Metal flow from excessive heat generation.

Total bearing lock-up Rollers skew, slide sideways and lock-up bearing.

Staining Surface stain with no significant corrosion from moisture exposure.

Etching Rusting with pitting and corrosion from moisture/water exposure.

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EXCESSIVE PRELOAD OR OVERLOAD

Line spalling Roller-spaced spalling from bearings operating after etching damage.

Rapid and deep spalling Caused by unusually high stresses. Full race width fatigue spalling is caused by heavy loads creating a thin lubricant film and possible elevated temperatures.

Geometric stress concentration Spalling from misalignment, deflections or heavy loading.

Inclusion origin Spalling from oxides or other hard inclusions in bearing steel.

Point surface origin Spalling from debris or raised metal exceeding the lubricant film thickness.

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IMPROPER FIT

Cone bore damage Fractured cone due to out-of-round or over-sized shaft.

Cup spinning Loose cup fit in a rotating wheel hub.

Abrasive wear Fine abrasive particle contamination.

Bruising Debris from other fatigued parts, inadequate sealing or poor maintenance.

Grooving Large particle contamination imbedding into soft cage material.

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EXCESSIVE END PLAY

Scalloping Uneven localized wear resulting from excessive end play.

Cage pocket wear Heavy contact between the rollers and cage pocket surfaces caused by bearing operating too loosely. RECOGNIZING AND PREVENTING DAMAGES OF BEARINGS: 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.

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

Abrasive Wear

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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.

Cage damage

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

Cage breakage

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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

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BEARING SPEED CAPABILITIES: The speed capability of a bearing in any application is subject to a number of factors including : • • • •

temperature bearing setting lubrication bearing design

The relative importance of each of these factors depends on the nature of the application. The effect of each factor is not isolated; each contributes, in varying degrees depending on the application, to the overall speed capability of the design. An understanding of how each of these factors affects performance as speeds change is required to achieve the speed capabilities inherent in tapered roller bearings. Measuring speed The usual measure of the speed of a tapered roller bearing is the circumferential velocity of the midpoint of the inner race large end rib (fig. 5-2), and this may be calculated as : Rib speed: Vr = pDmn / 60000 (m/s), where:

Vr = pDmn / 12 (ft/min)

Dm = Inner race rib diameter mm, in,

n = Bearing speed rev/min

Fig. 5-2 Inner race rib diameter. The inner race rib diameter may be scaled from a print or approximated as the average of the inner race inside diameter and the outer race outside diameter. The rib diameter at the midpoint of the roller end contact can be scaled from a drawing of the bearing, if available, or this diameter can be approximated as the average of the bearing I.D. and O.D. DN values (the product of the inner race bore in mm and the speed in rev/min) are often used as a measure of bearing speed. There is no direct relationship between the rib speed of a tapered roller bearing and DN value because of the wide variation in bearing cross sectional thickness. However, for rough approximation, one metre per second rib speed is about equal to 16 000 DN for average section bearings. One foot per minute is equal to approximately 80 DN.

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Speed capability guidelines

Fig. 5-3 Speed capability guidelines for various types of lubrication systems Fig. 5-3 is a summary of guidelines relating to speed and temperature. There are no clear-cut speed limitations for tapered roller bearings regardless of the bearing design or lubrication systems. The Timken Company recommends that testing be performed for all new high-speed applications. Bearing design Standard tapered roller bearings can operate at speeds up to about 30 m/s (6 000 ft/min or approximately 500 000 DN) ; specially designed high speed tapered roller bearings can operate successfully at speeds of over 200 m/s (40 000 ft/min or about 3 200 000 DN). These speeds can be achieved for either of these cases provided there is proper setting, adequate lubrication, no shock, vibration or unusual loading, and there is adequate heat dissipation. Bearing material limitations Standard bearing steels cannot maintain the desired minimum hot hardness of 58 Rc much above 135 °C (275 °F). Special steels that retain their hardness at elevated temperatures are available. Timken CBS 600 TM VIMVAR steel should be considered for temperatures between 150 to 230 °C (300 to 450 °F) and Timken CBS 1 000 TM VIMVAR steel should be used for temperatures above 230 °C (450 °F). WARNING: Never spin a bearing with compressed air. The force of the compressed air may cause the rollers to be expelled with great velocity, creating a risk of serious bodily harm. Proper bearing maintenance and handling practices are critical. Failure to follow installation instructions and failure to maintain proper lubrication can result in equipment failure, creating a risk of serious bodily harm. BEARING DYNAMICS AND SOUNDS:

Preface A growing awareness of noise pollution, prompted in part by government regulations, has been noticeable during recent years. One is hard-pressed to single out an industry that has not been affected, either as a user or a supplier. In its role as a supplier, The Timken Company can look back on a long history --predating the current emphasis on noise abatement by many years--of actively practicing noise control. This

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philosophy is exemplified not only by an extensive sound test program in its production facilities but also by an ongoing commitment to research in both fundamental and practical aspects of bearing related sound. It is very useful to picture the bearing as playing one of two very distinct roles. In one of these, its passive role as a transmitter, the bearing merely provides a path for energy transfer between the rotating and the stationary member, while in the second or active role, it causes its immediate environment to be excited by virtue of its rotation. It is important to recognize this distinction, particularly in situations calling for a diagnosis. Essentially, bearings play a significant role in the transmission of vibration in rotating equipment, however, they usually are not the predominant source of vibration. Nomenclature Symbol Description

Units

d0

Cone raceway mean diameter

mm, in

D0

Cup raceway mean diameter

mm, in

DW0

Roller mean diameter

mm, in

f

Excitation frequency

Hz

i

Harmonic index of carrier frequency, 0, 1, 2, 3, ..

j

Harmonic index of modulating frequency, 0, 1, 2, 3, ..

K1, K2, K3 Geometry-related constants Kbearing

Bearing stiffness

N/m, lbf/in

Khousing

Housing stiffness

N/m, lbf/in

Ksystem

System stiffness

N/m, lbf/in

S

Rotational speed

rpm

Z

Number of rollers per row

a (alpha) ½ included cup angle

degree

ß (beta) ½ included cone angle

degree

n (nu)

degree

½ included roller angle

The bearing as a transmitter Simply put, a bearing may be thought of as a massless spring/damper connecting a shaft to its housing. Typically, the interest lies in determining how vibration is transferred from housing to shaft or vice versa. For example, the excitation of meshing gears is carried along the shaft, through the bearing and to the exposed housing surface, where some of the energy is converted to airborne noise. Tapered roller bearings enjoy an advantage not found in other types of rolling element bearings. Since two bearings typically are adjusted against one another, the setting will govern the axial force. This influences the stiffness of the bearings and thereby the stiffness of the system. By merely varying the bearing setting, it may be possible to shift any unwelcome resonances out of the frequency range of interest. Maximum stiffnesses of approximately 1.75 x 109 N/m (10 x 106 lbf/in) are common in tapered roller bearings. In manipulating the system stiffness, it is essential that the stiffness of the bearing supports (housing) be taken into account. In simple conceptual terms: 1/Ksystem = 1/Kbearing + 1/Khousing While the prediction of stiffness in bearings is cumbersome at best, dealing with their damping characteristics is even more elusive. It has been demonstrated, however, that bearing setting will affect the amount of damping which can be realized.

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The bearing as an exciter The excitation potential of any rolling element bearing is determined primarily by the topography of its rolling surfaces. For example, a severe force pulse would result if a gross imperfection, such as a spall of sufficient size, were present in an operating bearing. Similarly, small imperfections such as brinell marks, nicks and any other deviations from perfect roundness of the components, will cause smaller fluctuation in the dynamic force. Hertzian theory tells us that even minute deformations can result in forces of significant magnitude. Therefore, this is the mechanism causing the bearing to act as an exciter. Surface irregularities of various origins lead to dynamic forces. These forces do not remain localized but are transmitted quite readily into the supporting structure. The dependence upon a number of rather unwieldy variables prohibits the mathematical determination of the magnitude of these forces in any one bearing. Their frequencies can be determined very accurately, though, from the gross dimensions of the bearing and its operating speed. Three constants can be defined in terms of either the angles or the diameters of the bearing:

These constants, along with the operating speed (S), the number of rollers (Z) and a harmonic index (i), permit the calculation of certain frequencies. They, in turn, identify specific disturbances (Table 1). Table 1 Type of Disturbance and Resulting Excitation Frequencies Disturbance

Frecuency, Hz

Eccentricity of Rotating Member

f0 = S/60

Out-of-Round of Rotating Member

f1i = i * f0

Roller Irregularity, e.g., nick or spall f2i = 2 * k1 * k3 * f1i Cone Irregularity, e.g., nick or spall f3i = Z * k1 * f1i Cup Irregularity, e.g., nick or spall

f4i = Z * k1 * k2 * f1i

Roller Size Variation (Rotating Cone) f5i = k1 * k2 * f1i Roller Size Variation (Rotating Cup) f6j = k1 * f1i Measurement considerations The frequencies listed in Table 1 are applicable whenever a bearing is evaluated. A typical approach employs an accelerometer attached on or near the bearing. By performing a narrow band frequency analysis of the acceleration signal, one can usually determine if the bearing is damaged or meets a user established vibration criterion. To avoid ambiguity when identifying the acceleration spikes occurring at the above frequencies, the bandwidth must be sufficiently narrow. For example, as the operating speed decreases, so should the bandwidth. It is not uncommon to observe modulation, particularly when the signal is obtained in a direction perpendicular to the axis of the bearing. Under these circumstances, the predominant evidence will be found at the frequencies f2i ± f5j or f2i ± f6j where i and j denote harmonic indices.

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Up to this point it has been assumed that the bearing operates with a 360º load zone. If this is not the case, such as when operating with radial load and end play, the rollers moving in and out of the load zone cause a spectrum that tends to have a "smeared" appearance. As one of the final steps in its quality assurance program, The Timken Company subjects its bearings to a vibration analysis in highly specialized, accelerometer-equipped test machines. In addition, the following rationale is employed: "The vibration (dynamic force) level of a bearing, operating at a specific speed and under a specific preload, is compared to and must meet an established standard. If this is the case, then by implication the geometric imperfections are of such small magnitude that the bearing's potential to act as an exciter is considered acceptable." Note that this implies that the merit of the bearing is strictly a function of the geometric imperfections, not one of speed and/or load and/or the bearing supports. The vibration signature may, of course, differ under other combinations of speed and load. Acoustic implications The mechanical energy in the bearing-generated dynamic forces and those presented to the bearing from the rotating member for transmission to the stationary member, will first be transferred to the structure supporting the bearing. The energy then permeates the structure and will be partially converted to acoustic energy upon arriving at an air/solid interface. Depending upon the mass, stiffness, geometry and boundary crossings characterizing the structure, the mechanical energy will undergo modifications. As a result of this transfer function, the prevailing acoustic energy (or airborne sound) will be a function not only of the mechanical vibration of the bearing but also the attenuation/amplification characteristics of each particular structure. One such structure is the quality assurance equipment employed by The Timken Company. Bearings are tested for vibration in a relatively unenclosed configuration, i.e., one in which a large percentage of the bearing surface is exposed. Clearly, this condition is acoustically quite different from one in which the bearing is fully enclosed, as for example, in a machine tool. The structure greatly influences the outcome of an acoustic measurement. Since sound is mainly caused by transverse vibration of the housing walls, a stiffer housing tends to be less noisy than one that is less rigid. Thus, any comparisons made or conclusions drawn between dissimilar structures are at best haphazard. The design of the structure can profoundly affect the overall noise characteristics of the system. This is the most important reason for not attaching sound level specifications, dB(A), to bearings. Design considerations Usually, resonances can be shifted or minimized by selective design, i.e., the shrewd manipulation of mass and/or stiffness. Where possible, impedance mismatches should be part of the design. For example, the vibration path between some electric motors and their bases is interrupted by rubberlike inserts. Also, consideration should be given to damping, either in the form of visco-elastic layers or mechanical discontinuities. The latter is realized wherever bolts, rivets or interference fits. Within this context, the excitation potential of the bearing can be optimized by a variety of different techniques. An increase in the operating speed of a bearing causes an upward shift toward the frequency range of maximum hearing sensitivity. Simultaneously, the overall vibration level increases. A variation in preload/end play of the bearing can be utilized to bring about a "most favorable" condition. Run-in will typically result in some "quieting". The same effect can be observed by going from a condition of marginal lubrication to one of "adequate" lubrication, but there is a point where additional lubricant flow no longer produces a benefit. It is good practice to fully enclose the bearing to minimize the direct acoustic path. Assistance is readily available from Timken Company sales engineers. Their experience can assist the user in selecting the proper bearing.

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