- Email: [email protected]
On the wear and failure of high speed roller bearings
Wear, 116 (1987)
361-
361
380
ON THE WEAR AND FAILURE OF HIGH SPEED ROLLER BEARINGS T. SAVASKAN* Atlantic Industrial Research Institute, Technical Halifax, Nova Scotia B3J 2X4 (Canada)
University
of Nova Scotia, PO Box 1000,
D.E.VEINOT Defence Research Establishment B3Y 327 (Canada)
Atlantic,
(Received January 2, 1986;revised
PO Box 1012, Dartmouth,
July 22, 1986;accepted
Nova Scotia
October 15,1986)
Summary A laboratory bearing test rig was constructed to study wear and failure mechanisms in high speed roller bearings under different conditions of loading and lubrication. The design of the rig permitted three bearings to be tested simultaneously. A test run of 720 h duration was performed during which wear of the bearings was monitored using ferrography and atomic absorption oil analysis. Wear debris particles collected from the bearings by ferrography were studied with scanning electron microscopy using energydispersive X-ray analysis. The results of ferrography were correlated with the metallographic observations made on the worn bearings. It was found that for lightly loaded high speed roller bearings rolling contact fatigue including surface pitting is not a significant wear or failure mechanism. However, misalignment due to shaft deflection above a critical point cannot be tolerated. Wear of high speed engine bearings can successfully be monitored by ferrography but, as the wear condition accelerates, frequent oil sample collections are beneficial.
1. Introduction Failure of rolling element bearings may result from a number of causes, the most common are rolling contact fatigue, wear, material defects or improper operating conditions (i.e. insufficient lubrication, improper loading or excessive misalignment).
*Present address: Trabzon, Turkey. 0043-1648/87/$3.50
Department
of Mechanical Engineering,
Karadeniz
University,
0 Elsevier Sequoia/Printed in The Netherlands
362
High speed rolling element bearings which operate at speeds greater than 24.5 m s-l or a bearing speed (DN) of the order of lo6 mm min’ are manufactured with special high precision and optimum internal clearances [ 11. Therefore excessive wear or dimensional changes due to thermal effects cannot be tolerated. Single-row cylindrical roller bearings are the most suitable in high speed engine applications, but the wear or failure of high speed roller bearings has received limited attention. Although cylindrical roller bearings have high load-carrying capability in the radial direction, the characteristic feature which affects the design of such bearings is their speed, not the load [2]. In high speed engine applications roller bearings support only the load caused by imbalance forces which may develop owing to the loss of turbine blades but they require accurate shaft alignment. In high speed applications excessive misalignment (or shaft deflection) can cause sudden bearing failure [31. Bearing failure due to material defects has been greatly reduced by using clean and properly heat-treated steels. Rolling contact fatigue can cause surface and subsurface-initiated bearing failures under extreme stress conditions. Subsurface microstructural changes, due to rolling contact fatigue, may occur in the bearings if the maximum shear stresses exceed the elastic limit of the bearing steel [4]. These changes may greatly reduce the fatigue life of bearings by initiating cracks which lead to spalling on the bearing surfaces. However, subsurface-initiated failures are not usual for high speed rolling bearings because the subsurface shear stresses which develop in the components of lightly loaded high speed engine bearings are normally well below the elastic limit of the bearing steel [2]. Wear of rolling bearings may result from metal-to-metal contact due to inadequate or improper lubrication. In high speed applications it is difficult to provide sufficient lubrication to all the bearing components, particularly to inner ring raceway and guide flanges, owing to the development of high centrifugal forces on the elements of lubricating oil. Hence many high speed bearings are now “under race” lubricated. Inadequate lubrication causes the bearings to operate at epicyclic speeds, giving rise to sliding of the rolling elements on the raceways. Both the rolling and the sliding action of the rolling elements may cause microstructural changes resulting in the formation of friction layers on the bearing surfaces [ 51. Formation of friction layers on the surfaces of high speed engine bearings can upset the hardness balance between the bearing components and result in excessive wear leading to failure [ 51. Since excessive wear of high speed engine bearings may result in catastrophic failure, it is necessary to monitor bearing wear during operation. Laboratory-simulated bearing wear tests may provide useful information on monitoring wear and studying wear mechanisms and failure modes of the high speed engine bearings. The purpose of this paper is to report on the development of a laboratory bearing test rig and to describe the observations made on three high speed roller bearings used for a test run.
363
2. Rearing test rig A bearing test rig was constructed to study wear of high speed roller bearings under different operating conditions. The design of the rig permitted three bearings to be tested s~ul~eou~y. The components of the test rig are shown in Fig. 1. These include a shaft assembly, a belt drive system, three bearing housing assemblies, numbered 1,2 and 3 respectively from the belt end of the shaft, lubrication supply lines and thrust bearings to prevent axial movement of the shaft. The inner rings of the test bearings are press fitted onto the shaft. The shaft is made of AISI 4140 steel which is hardened to 50 HRC and centreground to a surface finish of 0.3 pm r.m.s. The shaft is supported by the three test bearings while simple thrust bearings placed at both ends of the shaft prevent axial movement. Each thrust bearing consists of a hardened steel ball and a friction plate made of polished tool steel. The shaft is rotated by a belt attached to the pulley of a drive system. The drive system consists of an 11 kW (15 h.p.) a.c. electric motor running at a speed of 1750 rev min-‘. A speed ratio of 15 gives a rotational shaft speed between 25 000 and 26 000 rev mm-’ for the system. The lub~cation system contains an oil tank, an electrically driven oil pump, a pressure regulating valve, pressure gauges, a flow meter, an oil filter, oil flow valves and a cooling coil.
3. Test conditions In addition to the load on the bearings due to belt pull, the bearings were further loaded by shimming the number 3 bearing housing. To determine the static loading on the bearings, a calibration procedure was used. This involved using strain gauges to measure shaft deflections produced by applying known loads to the shaft through the number 2 bearing and
Fig. 1. A view of the bearing test rig.
364 TABLE
1
Data for bearing Running (h)
0 400 500 550 600 650 700
- 400 - 500 -550 -600 -650 - 700 - 720
time
testa Load on bearings (N) (+30 N)
Oil supply
(1 min-' )
rate
Temperature after IO h of running (“C)
Bearing number
Bearing number
Bearing number
1
2
3
1
2
3
1
2
3
Oil
1595 1664
892 1030
103 176
1.0 1.0 0.5 0.25 0.1 0 0
1.0 0.5 0.5 0.25 0.25 0.2 0.2
0.5 0 0 0 0 0 0
87 83 87 96 97 75 77
85 87 88 99 95 85 86
88 89 78 80 70 60 61
72 71 70 44 52 46 47
aShaft speed, 25 700 rev min-‘; rev min-‘; angular misalignment 0.0184”.
DN = 642 500 mm min-I; speed of bearing of number 2 bearing: 0 - 700 h, 0.0161”;
cage, 10 440 700 - 720 h,
through the belt. Using this procedure, two calibration curves of load us. shaft deflection were produced. These curves were used to determine the load applied to the shaft by the belt and the load developed on the number 2 bearing due to both belt pull and shimming. This procedure allowed the calculation of static loads imposed on the number 1 and number 3 bearings. The calculated static load values for each bearing are listed in Table 1. During 700 h of run the static load on each bearing was kept constant but after 700 h the load was increased slightly by increasing the shimming on the number 3 bearing housing. Each bearing was separately lubricated using an ester-based synthetic oil meeting MIL-L-23699C specification, under a constant oil pressure selected within a range of 200 - 300 kPa. The oil flow rate to each bearing was also separately selected and maintained constant for each selected running period to study the effect of the lubrication supply rate on the wear of the bearings. These flow rates are shown in Table 1. To accelerate wear of the bearings, the oil flow rate to each bearing was reduced systematically following each selected running period (see Table 1). After 400 h of operation the oil supply to the number 3 bearing was completely cut off and the number 1 bearing was run dry after 700 h. A 25 pm nominal size filter was used in the oil supply line. The thrust bearing at the belt end of the shaft was continuously lubricated by a simple oil dripping device, while the oil mist from the test bearings provided lubrication to the thrust bearing inside the machine. The bearing test machine was operated for approximately 15 h a day. For safe operation, test variables were monitored continuously by a computer interfaced to the machine. If either the speed of the shaft or the bearing cage, the axial shaft movement or excessive wandering of the drive
365
belt exceeded predetermined limits the computer would switch off the motor. The rotational speeds of the shaft and the bearing cages were continuously measured using fibre optic light guides. The signal collected from either shaft or cage by the fibre light guide was amplified and displayed on the screen of an oscilloscope in the form of a continuous square wave. The frequency of the wave corresponding to the rotational speed of either shaft or cage was read on a frequency counter and the speeds also recorded by the computer. The average speeds of both the shaft (25 700 rev min-‘) and the bearing cage (10 440 rev min-‘) varied by less than 100 rev min-’ during operation. The temperature of each bearing and of the lubricating oil was continuously monitored using type T ~e~ocouples and a digital the~ometer. Thermocouples were attached to the outer ring of each bearing. During each test period, the temperatures of the bearings and the lubricating oil stabilized after approximately 3 h, following sharp initial increases. After 5 h of run the increase in temperature was minor and the steady state temperatures of the bearings and the lubricating oil taken after 10 h of run are listed in Table 1. To monitor the wear of the bearings, oil sampfes were collected from each ~d~idu~ bearing every 10 h and from the oil tank at selected irregular intervals. The samples were subsequently analysed using direct reading (DR) ferrography, analytical ferrography and atomic absorption analysis. Ferrograms were made from the oil samples every 5.0 h in the first 650 h of run and then every 10 h until termination of the run. Oil samples were analysed for iron, copper and silver content by flame atomic absorption every 10 h during the entire run. The test run was terminated after 720 h owing to failure of the number 2 bearing.
4. Results 4.1. ~etui~o~p~y of new bearings Figure 2 is a diagram which shows the d~en~ons of the test bearings. The numbers 1 and 3 bearings were manufactured from AISI M-50 high speed tool steel, the chemical composition of which is given in Table 2. The typical microstructure of the M-50 bearing steel consisted of a dispersion of undissolved carbides in a matrix of tempered martensite as shown in Fig. 3. The carbides were found to be MC and M&J types in which molybdenum, vanadium and chromium were the major alloying elements. The carbides were distributed in the form of stringers in some areas but no non-metallic inclusions were found in the steel. The cages of these bearings were made of AISI 4340 steel and plated with silver to a measured thickness of approximately 27 pm.
366 I=d,,
Fig. 2. Diagram of test bearing: bore B, 25 mm; outside d.iameter D, 47 mm; width of inner ring Wi, 16 mm; width of outer ring W,, 12 mm; inner ring raceway diameter, 29.65 mm; outer ring raceway diameter, 42.35 mm; pitch diameter, 36.5 mm; roller diameter d, 6.35 mm; roller length 1, 6.35 mm; number of rollers, 14. Fig. 3. Microstructure of M-50 bearing steel. TABLE 2 Chemical composition of bearing steels Bearing number
AZSZ grade steel
1 and 3 2
M-50 52100
Composition
(wt.%)
C
Si
Mn
S
Cr
MO
V
0.86 1.0
0.34
0.35 0.32
0.02
4.15 1.60
6.62 -
1.24 -
The number 2 bearing was produced from AISI 52100 steel, the chemical composition of which is given in Table 2. The microstructure of this steel consisted of spheroidal carbides, (Fe, Cr),C, homogeneously distributed in a matrix of tempered martensite as shown in Fig. 4. The cage of this bearing was made of brass and plated with silver to a thickness of 16 pm. However, the inner ring of this bearing was deliberately replaced with one made of AISI M-50 steel, to study the effect of different steel components on the wear of the bearings. The dimensions of both inner rings (M-50 and 52100 steels) were exactly the same. X-ray diffraction analysis showed no indication of retained austenite in either AISI M-50 or 52100 bearing steel. The measured hardnesses for the components of the bearings are listed in Table 3. It can be seen that the hardnesses for the steel components of each bearing are approximately the same. Although the hardness of the M-50 steel is slightly higher than the hardness of the 52100 steel, both hardness values meet normal engineering requirements [ 6 1.
Fig. 4. Microstructure
of 52100 bearing steel.
Fig. 5. Results of DR ferrography obtained from oil samples from bearings and tank: number 1 bearing, 0, L, 0, S: number 2 bearing, A, L, A, S; number 3 bearing, 0, L, w, S; tank, X, L, +, S. (Large wear particles (L); small wear particles (S).)
TABLE 3 Measured hardnesses for components of test bearings Bearing number
Component
Hardness (HRC)
1 and 3 (M-50)
Roller Inner ring Outer ring
62.2 63.4 64.0
(252100)
Roller ring Inner Outer ring
60.6 62.1 61.9
4.2. Ferrograph y Oil samples collected from the bearings and the oil tank were analysed by DR ferrography and the results are plotted in Fig. 5. These results show that there was no significant wear before 600 h. A more severe wear situation developed after 650 h of operation. After 700 h a very severe wear situation developed suddenly and led to failure of the number 2 bearing at 720 h. Results of DR ferrography obtained from the oil samples of the bearings showed irregular variations in the amount of wear, instead of a steady increase in wear with time. This is because the results indicate the instan-
Fig. 6. Wear debris collected from number 2 bearing at 600 h. Fig. 7. Wear debris particles collected from number 2 bearing at 700 h.
taneous wear of each bearing only at the time of oil collection, not the cumulative wear. However, the results obtained from the oil tank correspond to cumulative wear of all bearings and hence provide a relatively early indication of the severe wear situation. Wear debris particles deposited on the ferrograms were studied by both optical microscopy and scanning electron microscopy (SEM) combined with energy-dispersive X-ray analysis. Ferrograms made of oil from the bearings and oil from the tank during 600 h of run showed no significant wear in any of the bearings. Figure 6 shows the entry deposits of a ferrogram made from oil collected from the number 2 bearing at 600 h. It can be seen that the entry deposit contained strings of tightly joined small flat particles (platelets) which were found to be mainly AISI 52100 bearing steel. However, after 650 h of run a significant wear situation started to develop particularly in the number 2 bearing as indicated by the large thick wear particles found on the ferrogram of oil from the number 2 bearing at 700 h (Fig. 7). The particles were from both M-50 and 52100 steel parts of the bearing. Similar particles were also observed on the ferrograms made from oil samples of the number 1 and number 3 bearings at this time (700 h). However, some of these particles probably originated from the number 2 bearing and were carried in the lubricating oil as it passed the 25 E.trn filter. The number 1 bearing showed very little wear even at 710 h. Figure 8 shows some large flat particles of 52100 steel deposited on a ferrogram made from oil collected from the number 1 bearing at 710 h which most likely originated from the number 2 bearing. At 710 h, very large (about 0.3 mm long) wear particles were found on the ferrogram of the number 2 bearing as shown in Fig. 9. Energy-dispersive X-ray analysis showed that these large particles were from the inner ring raceway (M-50 steel) of the bearing, indicating very severe wear of this component. Large silver particles were also found on the same ferrogram as shown in Fig. 10, indicating severe cage wear. Therefore analytical ferrography was in agreement with the DR ferrography in indicating the development of a very severe wear condition in the number 2 bearing after 700 h. The ferrogram of the number 3 bearing
Fig. 10. Silver particles collected from number 2 bearing at 710 h. Fig. 11. Wear particles collected from number 1 bearing at 720 h.
obtained at 710 h showed wear particles similar to those collected from the number 1 bearing at the same time. Ferrograms made from the oil samples of all the bearings and the tank collected at 720 h (during failure of the number 2 bearing) showed a drastic increase in the quantity of debris particles. Figure 11 shows the wear particles collected from the number 1 bearing at 720 h. The particles are from both M-50 and 52100 bearing steels. Small silver particles were also found below the entry deposit of the same ferrogram. A ferrogram made from oil collected from the number 2 bearing during its failure at 720 h showed dense entry deposits with large particles of both M-50 and 52100 bearing steels (Fig. 12). Silver and brass particles were found below the entry deposit of the same ferrogram. Wear particles collected from the number 3 bearing at 720 h are shown in Fig. 13. Both M-50 and 52100 bearing steel wear particles were found. In addition, silver and brass particles were found on the same ferrogram. The presence of wear debris particles from AISI 52100 steel and the brass cage on the ferrograms of both the number 1 and number 3 bearings suggests that the drastic increase in the quantity of wear particles resulted from the failure of the number 2 bearing. Again, it is likely that
Fig. 12. Wear debris collected Fig. 13. Wear particles
from number
from number
Fig. 14. Silver and brass particles
2 bearing
3 bearing
collected
collected
during failure at 720 h. at 720 h.
from oil tank at 720 h.
wear debris from the number 2 bearing was carried to the number 1 and number 3 bearings in the lubricating oil as it passed the filter. A ferrogram made from oil collected from the tank at 720 h showed a mixture of dense wear debris from the M-50 and 52100 bearing steels and particles from the brass cage. Figure 14 shows brass and silver particles below the entry deposit of this ferrogram. Atomic absorption spectrometric analysis of the oil samples collected from the individual bearings at 10 h intervals during the run indicated a change in wear status only on the samples collected at 720 h (termination of run) and on these samples the concentrations of copper, silver and iron increased substantially. Thus atomic absorption analysis was only able to indicate the change in status at 720 h, whereas ferrographic analysis (debris analysis) indicated this change much earlier. 4.3. Metallography of test bearings The surfaces of the bearing components were cleaned in solvents and electroplated with nickel to retain the edges of the samples. Each specimen was prepared using standard metallographic techniques and etched in Nital (6% Nital was used for the M-50 bearing steel and the samples of 52100 steel were etched in 2% Nital).
Fig. 15. Worn roller surface of number 1 bearing. Fig. 16. Worn edge of the roller shown in Fig. 15.
Fig. 17. Outer ring raceway of number 1 bearing after run. Fig. 18. Wear track on inner ring raceway of number 1 bearing. After 720 h of run, SEM examination revealed only small pits and fine score marks on the roller surfaces of the number 1 hearing as shown in Fig. 15.However, wear (scoring and pitting) was more severe on the roller edges as shown in Fig. 16. This observation suggests that roller skewing or tilting occurred during operation and resulted in metal-to-metal contact. Evidence of scoring and pitting was also found on the outer ring raceway of this bearing but the surface appeared less damaged than the roller surfaces, as shown in Fig. 17. The inner ring raceway appeared similar to the outer raceway but the rollers were more damaged than either raceway. Figure 18 shows the details of the wear track on the inner ring raceway. Although pits and score marks were observed on the rollers and the raceways of this bearing, the bearing was still in a usable condition and no noticeable damage to the cage was observed. However, the number 2 bearing failed completely at 720 h. Figure 19 shows the bearing after failure. It can be seen that the cage was cracked and the cage pockets were broken. Some rollers of this bearing were severely deformed resulting in macroscopic dimensional and shape changes. SEM examination showed that some areas of the roller surfaces, including the
372
Fig. 19. Photograph Fig. 20. Surface
of number
damage
after failure.
of a roller of number
Fig. 21. End face of a deformed Fig. 22. Cage material
2 bearing
2 bearing.
roller of number
(brass and silver) attached
2 bearing. to end face of a roller shown
in Fig. 21.
edges, were fractured. A typical example of the surface damage on the rollers is shown in Fig. 20. Damage was even more severe at the roller edges. The worst case example of this is shown in Fig. 21. This suggests that the observed gross deformation of the roller caused by heavy edge loading and rubbing occurred as a result of roller skewing. Energy-dispersive X-ray analysis showed that the end face of this roller was almost completely covered with material from the bearing cage. Figure 22 shows the particles of cage material (brass and silver) attached to the roller end face. This indicates that the roller ends were rubbing against the bearing cage under extreme pressure. This is an example of failure of high speed roller bearings caused by misalignment and development of excessive thrust load on the bearing due to shaft deflection. Examination of the rollers at higher magnifications showed that most surface areas were covered by small pits similar to those observed on the number 1 bearing. However, pitting was not a significant factor in the failure of this bearing. Hardness readings obtained from the roller ends showed a softening (3.5 HRC) in the average hardness of the rollers. This softening may have resulted from either formation of surface cracks or
Fig. 23. Wear damage
(pitting
Fig. 24. Inner ring raceway
Fig. 25. Details of damage Fig. 26. Surface
and scoring)
of number
on outer
2 bearing
on inner ring raceway
microstructure
of number
ring raceway
of number
2 bearing.
after failure.
of number
2 bearing.
1 bearing.
overtempering of surface material due to frictional heat, or a combination of both. Under visual examination, the centre of the outer ring raceway of this bearing appeared smooth but the areas along both sides of the raceway showed significant wear damage. Figure 23 shows the typical wear damage (pitting and scoring) observed on the outer ring raceway. The damaged inner ring raceway of this bearing is shown in Fig. 24. It can be seen that the failure of the inner ring raceway resulted from the removal of excessive surface material, producing a groove approximately 0.07 mm deep. An SEM micrograph, Fig. 25, shows the details of this groove. The observations suggest that the removal of raceway material resulted from extreme metal-to-metal contact (the ploughing action of the rollers on the raceway). Again, the cause of failure was misalignment produced by shaft deflection due to the increased shaming which also resulted in the movement of the inner ring on the shaft during failure. Visual inspection of the number 3 bearing after the run showed no significant damage. Examination of the bearing using SEM showed pits and score marks on the bearing components similar to those observed with the
Fig. 27. Cracks
in a roller of number
Fig.
material
28. White
(embedded
Fig. 29. Wear debris embedded Fig. 30. Friction bearing.
2 bearing. debris)
in surface
on the surface
material
layer with a fine microstructure
of a roller of number
of a roller of number on the surface
2 bearing.
2 bearing.
of a roller
from number
2
number 1 bearing. As observed with the number 1 bearing, the rollers suffered more surface damage (pitting and scoring) than either raceway. Damage was more severe at the roller edges than on the roller surfaces. Again, this was a result of roller skewing caused by shaft deflection. No microstructural changes were found in the components of the number 1 bearing. A typical surface microstructure of the components of this bearing is shown in Fig. 26. In addition, microhardness measurements obtained from the components of this bearing showed no distinct change below the surface. Hence metallographic examination and hardness measurements indicated no significant wear in this bearing. Many cracks were observed in the rollers of the number 2 bearing which failed at 720 h. Figure 27 shows the cracks on a section of a roller. A white etching material was also observed on some areas of the roller surfaces, particularly on the roller edges. Figure 28 shows the white etching material on the surface of an unplated roller. Examination of this material at higher magnifications suggested that it was not a part of the roller but was wear debris embedded in the surface material. Different wear debris particles
Fig. 31. White etching material on inner ring raceway of number 2 bearing. Fig. 32. SEM micrograph
of the same area shown in Fig. 31.
were also found on the roller surfaces. Figure 29 shows the heavily deformed wear debris particles of both the bearing steel (AISI 52100) and the brass cage embedded in the surface of a roller. The white etching material and the other embedded debris particles were found only in some surface regions. However, examination using SEM at higher magnifications revealed a distinct surface layer with a fine mi~rost~ctu~ in most surface areas of the rollers over a depth of approximately 6 pm below the surface. A typical example of this is shown in Fig. 30. No microstructural changes were found in the outer ring raceway of this bearing. However, a white etching material was observed on the inner ring raceway as shown in Fig. 31, An SEM micrograph, Fig. 32, of the same area suggests that the white etching material was not a part of the raceway but was wear debris embedded in the surface material as observed on the rollers of this bearing. Apart from the embedded wear debris, no distinct microstructural changes were found in the inner ring raceway. No microstructural or hardness changes were found in the components of the number 3 bearing. Hence no signific~t wear occurred in the number 3 bearing. 5. Discussion As shown in Table 1, the operating temperatures of the bearings increased slightly as the oil supply rates were reduced over the range 1 -0.25 1 min-I. This was expected because lubricating oil acts as a coolant for the bearings and hence a decreased coolant supply rate would result in increased bearing temperatures. However, below 0.25 1 min-i the operating temperatures of the bearings decreased with decreasing oil supply rate. In particular, when the oil supply to either the number 1 or number 3 bearing was completely cut off the operating temperature of each bearing reached the lowest level. It is well established that the load-independent friction torque increases with the amount of lubricant as a result of increased churning [7]. This can explain the higher temperatures at higher flow rates, despite the
376
increased cooling. The measured temperatures of the bearings, therefore, suggest that lowering the oil supply rate below a certain range (0.2 1 mini) or oil mist lubrication may be beneficial in reducing operating temperatures of high speed engine bearings. This observation is also in agreement with the results of Hargreaves and Higginson [8] who reported that friction torque and operating temperatures for cylindrical roller bearings could be greatly reduced by running at low lubricant supply rates or with oil mist lubrication. The number 1 and number 3 bearings showed no significant wear but the number 2 bearing failed completely. These observations may be explained in terms of the running conditions (loading and lubrication) and production of bearings including bearing materials. The number 1 bearing was sufficiently lubricated until the last 20 h. Although this bearing was the most heavily loaded and run dry for the last 20 h, it showed no evidence of significant wear, except minor surface pitting and scoring. Surface examination showed some evidence of roller skewing but the misalignment of this bearing was not high enough to cause severe wear or failure. Although the number 2 bearing was less heavily loaded than the number 1 bearing, it failed completely. The number 2 hearing occupies a critical point on the shaft because it constrains the shaft deflections produced by belt pull and shimming of the number 3 bearing housing. Shaft deflections due to the loading procedure produce relative race misalignments in the bearings. The calculated values of angular misalignment for the number 2 bearing are listed in Table 1. Excessive bearing misalignment due to increased shaft deflections may result in large skew motion of the rollers leading to seizure and failure. Large shaft deflections may also produce complex forces on the bearings during rotation of the shaft. One of these forces is the development of excessive thrust (axial) load on the bearings which in this case leads to an observed movement of the inner ring of the number 2 bearing on the shaft during failure. Development of thrust load on the number 2 bearing also resulted in fracture of the roller edges, the cage and edges of the guide flanges. Although cylindrical roller bearings can be used to carry combined radial and thrust loads [ 91, development of excessive thrust load may result in sudden bearing failure [ 31. This bearing was operated successfully for 700 h but failed suddenly at 720 h following a small increase (25 pm) in shimming of the number 3 bearing housing. The magnitude of the shaft deflection that shimming produced depends on the amount of shimming and the shaft length (distances between the bearings). Shaft deflection above a critical point, the value of which depends on the shaft material, shaft length, load and speed, may result in failure of either the shaft or the bearing, or both [ 31. In the present case this resulted in the failure of the number 2 bearing. This bearing was successfully operated with an angular misalignment of 0.0161” for 700 h but it failed in 20 h after the misalignment was increased to 0.0184“ by the increase in shimming. Hence the present work suggests that misalignment above a critical point (0.0161”) cannot be tolerated in these bearings,
371
The number 3 bearing was the least heavily loaded but was operated without an oil supply after 400 h. After completion of the run this bearing showed no evidence of significant wear, except minor pitting and scoring. SEM examination showed some evidence for roller skewing but neither the level of load nor the magnitude of misalignment was sufficient to cause serious damage. Although the bearing was run without an oil supply for 320 h, no distinct evidence of sliding or skidding was observed. This observation suggests that the oil mist produced from the lubricated bearings prevented sliding and metal-to-metal contact in this bearing. SEM examination showed that the rollers of the bearings suffered more pitting than the raceways. Since surface pitting can result from fatigue even under ideal operating conditions, it is possible that the rollers were subjected to either a higher number of stress cycles or a higher magnitude of stress per shaft revolution than either raceway. The number of stress cycles that bearing components undergo per shaft revolution depends on the bearing parameters, including the number of rollers and their locations. In high speed applications development of centrifugal forces may increase both the number and the level of contact stresses on the rollers and the outer ring raceway. In the present case the centrifugal forces acting on each roller (about 34 N) were insignificantly low and the contact stress on the outer raceway of each bearing was lower than the stress produced on the inner ring raceway because of contact geometry. However, the rollers and the outer raceway of each bearing would be expected to receive a higher number of contact stresses than the inner raceway, per shaft revolution, because of a continuously acting centrifugal force on each roller. In addition, the surfaces of the relatively small rollers would be expected to attain the highest temperatures in operation. Therefore the combination of both the higher number of stress cycles and the higher temperatures might be the reason for the observed heavier pitting of the roller surfaces. No distinct microstructural changes or friction layers were found on the components of either the number 1 or the number 3 bearing, although these bearings were run without oil supply for some period of time. Similar to the surface examination, this observation suggests that the oil mist lubrication prevented sliding and metal-to-metal contact in these bearings. Microstructural changes were found only on the rollers of the number 2 bearing which were made of AISI 52100 steel. These changes resulted in the formation of a distinct surface layer with a very fine microstructure. In addition, white etching wear debris was found embedded in the surface material of the rollers. Similar microstructural changes were also observed on the rollers and the raceways of a used high speed turbine engine bearing [ 51. The transformation products, including white etching material and the surface layer with a fine microstructure, were identified as friction martensite derived from the original microstructure of the bearing steel owing to the heat of friction. White etching wear debris was also found on the heavily damaged inner ring raceway of this bearing but this raceway showed no distinct microstructural changes. Comparison of the observations made on the
378
different bearings and the different bearing components suggested that AISI M-50 bearing steel is more resistant to friction-induced microstructural changes than AISI 52100 steel. Although the radial loads on the number 1 and the number 2 bearings were much higher than the dynamic loads which they normally carry in high speed engines (less than 550 N), no subsurface microstructural changes were found in any components of these bearings. Stress-induced subsurface microst~ctur~ changes may occur if the m~imum shear stresses exceed a threshold value approximately equal to the elastic limit of the bearing steel [ 41. For AISI 52100 bearing steel this value is reported to be 724 MPa [ 41. The calculated maximum shear stresses in the test bearings (577 MPa for the number 1, 454 MPa for the number 2 and 188 MPa for the number 3 bearing) were well below the proportional elastic limits of the bearing steels; between 1000 and 1600 MPa for M-50 [lo] and approx~ately ‘700 MPa for 52100 steel [ 4f. Thus mechanical wear or failure due to subsurface rolling contact fatigue would not be expected in the test bearings. To obtain shear stresses above the reported threshold value (724 MPa) in the components of the test bearings, the applied radial load on each bearing must be above 3000 N. No threshold value has been reported for AISI M-50 steel but a higher critical stress value would be expected for the stress-induced microstructural changes in this steel. Therefore, for radial loads below approximately 3000 N, subsurface rolling contact fatigue should not be a significant factor in the wear or failure of the test bearings. Ferrography, which enables monitoring of progressive changes, indicated no si~ific~t wear in the bearings until 650 h. Results of DR ferrography, and the ferrograms made of oil collected from the bearings during the first 650 h of run, indicated only a mild wear condition. However, after 650 h a more severe wear condition developed in the number 2 bearing. This was indicated by a change in the size and shape of wear debris particles collected from #is bearing at 700 h. Ferrograms made from oil samples collected from the number 1 and the number 3 bearings also indicated the change from mild to severe wear after 650 h but the compositional analysis of the wear debris showed that these particles originated from the number 2 bearing. The number 2 bearing failed suddenly at 720 h owing to increased shaft deflection (misalignment) which resulted in the movement of the inner ring on the shaft. During this period of time, ferrography provided sufficient evidence of the extreme operating condition by showing a drastic increase in the size and quantity of the wear particles from the number 2 bearing at 710 h. In addition, ferrograms showed the presence of relatively large silver particles in the lubricating oil at 710 h which indicated severe cage wear. Ferrographic analysis of the oil samples collected from the bearings and the tank at 720 h (during failure of the number 2 bearing) showed a large increase in the quantity of the wear debris particles. Compositional analysis of debris on the ferrograms showed that the wear particles were from all the components of the number 2 bearing, including rollers and raceways (52100
379
and M-50 steels) and cage material (brass and silver). Silver particles were present in oil collected after 700 h but brass particles were found only during failure of the number 2 bearing at 720 h. This suggests that the presence of silver particles in the oil indicates a severe wear condition in the bearings but the presence of brass particles from the cage is indicative of bearing failure. Present work suggests that under normal operating conditions wear of high speed engine bearings may successfully be monitored by ferrography but for abnormal wear conditions frequent oil sample collections are necessary. Spectrometric oil analysis (atomic absorption) indicated a change in wear status only at 720 h, after the number 2 bearing failed completely. Thus, in this case, ferrographic analysis or debris analysis provided an earlier indication of change in wear status than spectrometric oil analysis. Comparison of ferrographic results (debris analysis) with the worn surface examination suggests that normal wear particles (small platelets) resulted from pitting of rollers and the raceways of the bearings. Large and thick debris particles resulted from more severe wear (spalling and fracture of rollers and the inner raceway) of the number 2 bearing.
6. Conclusions
(1) Misalignment of high speed roller bearings above a critical point results in sudden bearing failure. (2) Rolling contact fatigue is not a significant wear or failure mechanism in high speed engine bearings. (3) Operating temperatures of high speed engine bearings may be decreased by reducing the oil supply rate or by running with oil mist lubrication. (4) AISI M-50 bearing steel is more resistant to friction-induced microstructural changes than AISI 52100 bearing steel. (5) Wear of high speed engine bearings may successfully be monitored by ferrography but as the wear condition accelerates more frequent oil collections are beneficial. (6) Ferrographic oil analysis provides an earlier indication of a change in wear condition than atomic absorption spectrometric oil analysis. (7) Small flat wear debris particles (platelets) from the bearing steel indicate a mild wear condition and large bulky particles indicate a severe wear mechanism in the test bearings. (8) The presence of silver particles in oil from the bearings indicates a severe wear condition but the presence of large brass particles (cage material) is indicative of bearing failure.
380
Acknowledgments The authors wish to thank Mr. C. Collings and Dr. A. W. Hanson of Atlantic Region Laboratory, National Research Council of Canada, for the SEM facilities. Mr. F. Roma, Mr. C. Jackson and Ms. M. Murphy provided technical assistance and Mrs. G. Veinot typed the manuscript. Finally, the authors would like to thank Professor G. P. Wilson for the management of the Contract and Dr. E. E. Laufer for discussions.
References 1 M. C. Shaw and F. Macks, Analysis and Lubrication of Bearings, McGraw-Hill, New York, 1949, Chapter 10. 2 H. J. Koeber, Design and calculation of high speed engine bearings, 60th AGARD Specialists Meet. on Aircraft Gear and Bearings, Tribological Systems, San Antonio, TX, April 1985, Paper 5. 3 T. Savaskan, E. E. Laufer and D. E. Veinot, Wear of high speed roller bearings, 60th AGARD Specialists Meet. on Aircraft Gear and Bearings, Tribological Systems, San Antonio, TX, April 1985, Paper 12. 4 J. A. Martin, S. F. Borgese and A. D. Eberdardt, Microstructural alterations of rolling steel undergoing cyclic stressing, J. Basic Eng., 59 (1966) 555 - 567. 5 T. Savaskan and E. E. Laufer, Wear in a high speed roller bearing, Met. Technol. NY, 11 (1984) 530 - 534. 6 Metals Handbook, Vol. 10, American Society for Metals, Metals Park, OH, 8th edn., 1975, pp. 416 - 436. 7 A. Sahlgren, in D. Dowson, C. M. Taylor, M. Godet and D. Berthe (eds.), Reference Speeds - A New Concept for Rolling Bearings, Proc. 7th Leeds-Lyon Symp., Leeds, September 9 - 12, 1980, Westbury House, Guildford, 1981. 8 R. H. Hargreaves and G. R. Higginson, Some effects of lubricant starvation in cylindrical roller bearings, J. Lubr. Technol., 98 (1976) 66 - 70. 9 S. R. Brown and S. Y. Poon, The lubrication of the roller-rib contacts of a radial cylindrical roller bearing carrying thrust load, ASLE Trans., 26 (3) (1983) 317 - 324. 10 J. A. Rescalvo and B. L. Averbach, Fracture and fatigue in M-50 and 18-4-l high speed steels, Metal. Trans., 1 OA (1979) 1265 - 1271.
361-
361
380
ON THE WEAR AND FAILURE OF HIGH SPEED ROLLER BEARINGS T. SAVASKAN* Atlantic Industrial Research Institute, Technical Halifax, Nova Scotia B3J 2X4 (Canada)
University
of Nova Scotia, PO Box 1000,
D.E.VEINOT Defence Research Establishment B3Y 327 (Canada)
Atlantic,
(Received January 2, 1986;revised
PO Box 1012, Dartmouth,
July 22, 1986;accepted
Nova Scotia
October 15,1986)
Summary A laboratory bearing test rig was constructed to study wear and failure mechanisms in high speed roller bearings under different conditions of loading and lubrication. The design of the rig permitted three bearings to be tested simultaneously. A test run of 720 h duration was performed during which wear of the bearings was monitored using ferrography and atomic absorption oil analysis. Wear debris particles collected from the bearings by ferrography were studied with scanning electron microscopy using energydispersive X-ray analysis. The results of ferrography were correlated with the metallographic observations made on the worn bearings. It was found that for lightly loaded high speed roller bearings rolling contact fatigue including surface pitting is not a significant wear or failure mechanism. However, misalignment due to shaft deflection above a critical point cannot be tolerated. Wear of high speed engine bearings can successfully be monitored by ferrography but, as the wear condition accelerates, frequent oil sample collections are beneficial.
1. Introduction Failure of rolling element bearings may result from a number of causes, the most common are rolling contact fatigue, wear, material defects or improper operating conditions (i.e. insufficient lubrication, improper loading or excessive misalignment).
*Present address: Trabzon, Turkey. 0043-1648/87/$3.50
Department
of Mechanical Engineering,
Karadeniz
University,
0 Elsevier Sequoia/Printed in The Netherlands
362
High speed rolling element bearings which operate at speeds greater than 24.5 m s-l or a bearing speed (DN) of the order of lo6 mm min’ are manufactured with special high precision and optimum internal clearances [ 11. Therefore excessive wear or dimensional changes due to thermal effects cannot be tolerated. Single-row cylindrical roller bearings are the most suitable in high speed engine applications, but the wear or failure of high speed roller bearings has received limited attention. Although cylindrical roller bearings have high load-carrying capability in the radial direction, the characteristic feature which affects the design of such bearings is their speed, not the load [2]. In high speed engine applications roller bearings support only the load caused by imbalance forces which may develop owing to the loss of turbine blades but they require accurate shaft alignment. In high speed applications excessive misalignment (or shaft deflection) can cause sudden bearing failure [31. Bearing failure due to material defects has been greatly reduced by using clean and properly heat-treated steels. Rolling contact fatigue can cause surface and subsurface-initiated bearing failures under extreme stress conditions. Subsurface microstructural changes, due to rolling contact fatigue, may occur in the bearings if the maximum shear stresses exceed the elastic limit of the bearing steel [4]. These changes may greatly reduce the fatigue life of bearings by initiating cracks which lead to spalling on the bearing surfaces. However, subsurface-initiated failures are not usual for high speed rolling bearings because the subsurface shear stresses which develop in the components of lightly loaded high speed engine bearings are normally well below the elastic limit of the bearing steel [2]. Wear of rolling bearings may result from metal-to-metal contact due to inadequate or improper lubrication. In high speed applications it is difficult to provide sufficient lubrication to all the bearing components, particularly to inner ring raceway and guide flanges, owing to the development of high centrifugal forces on the elements of lubricating oil. Hence many high speed bearings are now “under race” lubricated. Inadequate lubrication causes the bearings to operate at epicyclic speeds, giving rise to sliding of the rolling elements on the raceways. Both the rolling and the sliding action of the rolling elements may cause microstructural changes resulting in the formation of friction layers on the bearing surfaces [ 51. Formation of friction layers on the surfaces of high speed engine bearings can upset the hardness balance between the bearing components and result in excessive wear leading to failure [ 51. Since excessive wear of high speed engine bearings may result in catastrophic failure, it is necessary to monitor bearing wear during operation. Laboratory-simulated bearing wear tests may provide useful information on monitoring wear and studying wear mechanisms and failure modes of the high speed engine bearings. The purpose of this paper is to report on the development of a laboratory bearing test rig and to describe the observations made on three high speed roller bearings used for a test run.
363
2. Rearing test rig A bearing test rig was constructed to study wear of high speed roller bearings under different operating conditions. The design of the rig permitted three bearings to be tested s~ul~eou~y. The components of the test rig are shown in Fig. 1. These include a shaft assembly, a belt drive system, three bearing housing assemblies, numbered 1,2 and 3 respectively from the belt end of the shaft, lubrication supply lines and thrust bearings to prevent axial movement of the shaft. The inner rings of the test bearings are press fitted onto the shaft. The shaft is made of AISI 4140 steel which is hardened to 50 HRC and centreground to a surface finish of 0.3 pm r.m.s. The shaft is supported by the three test bearings while simple thrust bearings placed at both ends of the shaft prevent axial movement. Each thrust bearing consists of a hardened steel ball and a friction plate made of polished tool steel. The shaft is rotated by a belt attached to the pulley of a drive system. The drive system consists of an 11 kW (15 h.p.) a.c. electric motor running at a speed of 1750 rev min-‘. A speed ratio of 15 gives a rotational shaft speed between 25 000 and 26 000 rev mm-’ for the system. The lub~cation system contains an oil tank, an electrically driven oil pump, a pressure regulating valve, pressure gauges, a flow meter, an oil filter, oil flow valves and a cooling coil.
3. Test conditions In addition to the load on the bearings due to belt pull, the bearings were further loaded by shimming the number 3 bearing housing. To determine the static loading on the bearings, a calibration procedure was used. This involved using strain gauges to measure shaft deflections produced by applying known loads to the shaft through the number 2 bearing and
Fig. 1. A view of the bearing test rig.
364 TABLE
1
Data for bearing Running (h)
0 400 500 550 600 650 700
- 400 - 500 -550 -600 -650 - 700 - 720
time
testa Load on bearings (N) (+30 N)
Oil supply
(1 min-' )
rate
Temperature after IO h of running (“C)
Bearing number
Bearing number
Bearing number
1
2
3
1
2
3
1
2
3
Oil
1595 1664
892 1030
103 176
1.0 1.0 0.5 0.25 0.1 0 0
1.0 0.5 0.5 0.25 0.25 0.2 0.2
0.5 0 0 0 0 0 0
87 83 87 96 97 75 77
85 87 88 99 95 85 86
88 89 78 80 70 60 61
72 71 70 44 52 46 47
aShaft speed, 25 700 rev min-‘; rev min-‘; angular misalignment 0.0184”.
DN = 642 500 mm min-I; speed of bearing of number 2 bearing: 0 - 700 h, 0.0161”;
cage, 10 440 700 - 720 h,
through the belt. Using this procedure, two calibration curves of load us. shaft deflection were produced. These curves were used to determine the load applied to the shaft by the belt and the load developed on the number 2 bearing due to both belt pull and shimming. This procedure allowed the calculation of static loads imposed on the number 1 and number 3 bearings. The calculated static load values for each bearing are listed in Table 1. During 700 h of run the static load on each bearing was kept constant but after 700 h the load was increased slightly by increasing the shimming on the number 3 bearing housing. Each bearing was separately lubricated using an ester-based synthetic oil meeting MIL-L-23699C specification, under a constant oil pressure selected within a range of 200 - 300 kPa. The oil flow rate to each bearing was also separately selected and maintained constant for each selected running period to study the effect of the lubrication supply rate on the wear of the bearings. These flow rates are shown in Table 1. To accelerate wear of the bearings, the oil flow rate to each bearing was reduced systematically following each selected running period (see Table 1). After 400 h of operation the oil supply to the number 3 bearing was completely cut off and the number 1 bearing was run dry after 700 h. A 25 pm nominal size filter was used in the oil supply line. The thrust bearing at the belt end of the shaft was continuously lubricated by a simple oil dripping device, while the oil mist from the test bearings provided lubrication to the thrust bearing inside the machine. The bearing test machine was operated for approximately 15 h a day. For safe operation, test variables were monitored continuously by a computer interfaced to the machine. If either the speed of the shaft or the bearing cage, the axial shaft movement or excessive wandering of the drive
365
belt exceeded predetermined limits the computer would switch off the motor. The rotational speeds of the shaft and the bearing cages were continuously measured using fibre optic light guides. The signal collected from either shaft or cage by the fibre light guide was amplified and displayed on the screen of an oscilloscope in the form of a continuous square wave. The frequency of the wave corresponding to the rotational speed of either shaft or cage was read on a frequency counter and the speeds also recorded by the computer. The average speeds of both the shaft (25 700 rev min-‘) and the bearing cage (10 440 rev min-‘) varied by less than 100 rev min-’ during operation. The temperature of each bearing and of the lubricating oil was continuously monitored using type T ~e~ocouples and a digital the~ometer. Thermocouples were attached to the outer ring of each bearing. During each test period, the temperatures of the bearings and the lubricating oil stabilized after approximately 3 h, following sharp initial increases. After 5 h of run the increase in temperature was minor and the steady state temperatures of the bearings and the lubricating oil taken after 10 h of run are listed in Table 1. To monitor the wear of the bearings, oil sampfes were collected from each ~d~idu~ bearing every 10 h and from the oil tank at selected irregular intervals. The samples were subsequently analysed using direct reading (DR) ferrography, analytical ferrography and atomic absorption analysis. Ferrograms were made from the oil samples every 5.0 h in the first 650 h of run and then every 10 h until termination of the run. Oil samples were analysed for iron, copper and silver content by flame atomic absorption every 10 h during the entire run. The test run was terminated after 720 h owing to failure of the number 2 bearing.
4. Results 4.1. ~etui~o~p~y of new bearings Figure 2 is a diagram which shows the d~en~ons of the test bearings. The numbers 1 and 3 bearings were manufactured from AISI M-50 high speed tool steel, the chemical composition of which is given in Table 2. The typical microstructure of the M-50 bearing steel consisted of a dispersion of undissolved carbides in a matrix of tempered martensite as shown in Fig. 3. The carbides were found to be MC and M&J types in which molybdenum, vanadium and chromium were the major alloying elements. The carbides were distributed in the form of stringers in some areas but no non-metallic inclusions were found in the steel. The cages of these bearings were made of AISI 4340 steel and plated with silver to a measured thickness of approximately 27 pm.
366 I=d,,
Fig. 2. Diagram of test bearing: bore B, 25 mm; outside d.iameter D, 47 mm; width of inner ring Wi, 16 mm; width of outer ring W,, 12 mm; inner ring raceway diameter, 29.65 mm; outer ring raceway diameter, 42.35 mm; pitch diameter, 36.5 mm; roller diameter d, 6.35 mm; roller length 1, 6.35 mm; number of rollers, 14. Fig. 3. Microstructure of M-50 bearing steel. TABLE 2 Chemical composition of bearing steels Bearing number
AZSZ grade steel
1 and 3 2
M-50 52100
Composition
(wt.%)
C
Si
Mn
S
Cr
MO
V
0.86 1.0
0.34
0.35 0.32
0.02
4.15 1.60
6.62 -
1.24 -
The number 2 bearing was produced from AISI 52100 steel, the chemical composition of which is given in Table 2. The microstructure of this steel consisted of spheroidal carbides, (Fe, Cr),C, homogeneously distributed in a matrix of tempered martensite as shown in Fig. 4. The cage of this bearing was made of brass and plated with silver to a thickness of 16 pm. However, the inner ring of this bearing was deliberately replaced with one made of AISI M-50 steel, to study the effect of different steel components on the wear of the bearings. The dimensions of both inner rings (M-50 and 52100 steels) were exactly the same. X-ray diffraction analysis showed no indication of retained austenite in either AISI M-50 or 52100 bearing steel. The measured hardnesses for the components of the bearings are listed in Table 3. It can be seen that the hardnesses for the steel components of each bearing are approximately the same. Although the hardness of the M-50 steel is slightly higher than the hardness of the 52100 steel, both hardness values meet normal engineering requirements [ 6 1.
Fig. 4. Microstructure
of 52100 bearing steel.
Fig. 5. Results of DR ferrography obtained from oil samples from bearings and tank: number 1 bearing, 0, L, 0, S: number 2 bearing, A, L, A, S; number 3 bearing, 0, L, w, S; tank, X, L, +, S. (Large wear particles (L); small wear particles (S).)
TABLE 3 Measured hardnesses for components of test bearings Bearing number
Component
Hardness (HRC)
1 and 3 (M-50)
Roller Inner ring Outer ring
62.2 63.4 64.0
(252100)
Roller ring Inner Outer ring
60.6 62.1 61.9
4.2. Ferrograph y Oil samples collected from the bearings and the oil tank were analysed by DR ferrography and the results are plotted in Fig. 5. These results show that there was no significant wear before 600 h. A more severe wear situation developed after 650 h of operation. After 700 h a very severe wear situation developed suddenly and led to failure of the number 2 bearing at 720 h. Results of DR ferrography obtained from the oil samples of the bearings showed irregular variations in the amount of wear, instead of a steady increase in wear with time. This is because the results indicate the instan-
Fig. 6. Wear debris collected from number 2 bearing at 600 h. Fig. 7. Wear debris particles collected from number 2 bearing at 700 h.
taneous wear of each bearing only at the time of oil collection, not the cumulative wear. However, the results obtained from the oil tank correspond to cumulative wear of all bearings and hence provide a relatively early indication of the severe wear situation. Wear debris particles deposited on the ferrograms were studied by both optical microscopy and scanning electron microscopy (SEM) combined with energy-dispersive X-ray analysis. Ferrograms made of oil from the bearings and oil from the tank during 600 h of run showed no significant wear in any of the bearings. Figure 6 shows the entry deposits of a ferrogram made from oil collected from the number 2 bearing at 600 h. It can be seen that the entry deposit contained strings of tightly joined small flat particles (platelets) which were found to be mainly AISI 52100 bearing steel. However, after 650 h of run a significant wear situation started to develop particularly in the number 2 bearing as indicated by the large thick wear particles found on the ferrogram of oil from the number 2 bearing at 700 h (Fig. 7). The particles were from both M-50 and 52100 steel parts of the bearing. Similar particles were also observed on the ferrograms made from oil samples of the number 1 and number 3 bearings at this time (700 h). However, some of these particles probably originated from the number 2 bearing and were carried in the lubricating oil as it passed the 25 E.trn filter. The number 1 bearing showed very little wear even at 710 h. Figure 8 shows some large flat particles of 52100 steel deposited on a ferrogram made from oil collected from the number 1 bearing at 710 h which most likely originated from the number 2 bearing. At 710 h, very large (about 0.3 mm long) wear particles were found on the ferrogram of the number 2 bearing as shown in Fig. 9. Energy-dispersive X-ray analysis showed that these large particles were from the inner ring raceway (M-50 steel) of the bearing, indicating very severe wear of this component. Large silver particles were also found on the same ferrogram as shown in Fig. 10, indicating severe cage wear. Therefore analytical ferrography was in agreement with the DR ferrography in indicating the development of a very severe wear condition in the number 2 bearing after 700 h. The ferrogram of the number 3 bearing
Fig. 10. Silver particles collected from number 2 bearing at 710 h. Fig. 11. Wear particles collected from number 1 bearing at 720 h.
obtained at 710 h showed wear particles similar to those collected from the number 1 bearing at the same time. Ferrograms made from the oil samples of all the bearings and the tank collected at 720 h (during failure of the number 2 bearing) showed a drastic increase in the quantity of debris particles. Figure 11 shows the wear particles collected from the number 1 bearing at 720 h. The particles are from both M-50 and 52100 bearing steels. Small silver particles were also found below the entry deposit of the same ferrogram. A ferrogram made from oil collected from the number 2 bearing during its failure at 720 h showed dense entry deposits with large particles of both M-50 and 52100 bearing steels (Fig. 12). Silver and brass particles were found below the entry deposit of the same ferrogram. Wear particles collected from the number 3 bearing at 720 h are shown in Fig. 13. Both M-50 and 52100 bearing steel wear particles were found. In addition, silver and brass particles were found on the same ferrogram. The presence of wear debris particles from AISI 52100 steel and the brass cage on the ferrograms of both the number 1 and number 3 bearings suggests that the drastic increase in the quantity of wear particles resulted from the failure of the number 2 bearing. Again, it is likely that
Fig. 12. Wear debris collected Fig. 13. Wear particles
from number
from number
Fig. 14. Silver and brass particles
2 bearing
3 bearing
collected
collected
during failure at 720 h. at 720 h.
from oil tank at 720 h.
wear debris from the number 2 bearing was carried to the number 1 and number 3 bearings in the lubricating oil as it passed the filter. A ferrogram made from oil collected from the tank at 720 h showed a mixture of dense wear debris from the M-50 and 52100 bearing steels and particles from the brass cage. Figure 14 shows brass and silver particles below the entry deposit of this ferrogram. Atomic absorption spectrometric analysis of the oil samples collected from the individual bearings at 10 h intervals during the run indicated a change in wear status only on the samples collected at 720 h (termination of run) and on these samples the concentrations of copper, silver and iron increased substantially. Thus atomic absorption analysis was only able to indicate the change in status at 720 h, whereas ferrographic analysis (debris analysis) indicated this change much earlier. 4.3. Metallography of test bearings The surfaces of the bearing components were cleaned in solvents and electroplated with nickel to retain the edges of the samples. Each specimen was prepared using standard metallographic techniques and etched in Nital (6% Nital was used for the M-50 bearing steel and the samples of 52100 steel were etched in 2% Nital).
Fig. 15. Worn roller surface of number 1 bearing. Fig. 16. Worn edge of the roller shown in Fig. 15.
Fig. 17. Outer ring raceway of number 1 bearing after run. Fig. 18. Wear track on inner ring raceway of number 1 bearing. After 720 h of run, SEM examination revealed only small pits and fine score marks on the roller surfaces of the number 1 hearing as shown in Fig. 15.However, wear (scoring and pitting) was more severe on the roller edges as shown in Fig. 16. This observation suggests that roller skewing or tilting occurred during operation and resulted in metal-to-metal contact. Evidence of scoring and pitting was also found on the outer ring raceway of this bearing but the surface appeared less damaged than the roller surfaces, as shown in Fig. 17. The inner ring raceway appeared similar to the outer raceway but the rollers were more damaged than either raceway. Figure 18 shows the details of the wear track on the inner ring raceway. Although pits and score marks were observed on the rollers and the raceways of this bearing, the bearing was still in a usable condition and no noticeable damage to the cage was observed. However, the number 2 bearing failed completely at 720 h. Figure 19 shows the bearing after failure. It can be seen that the cage was cracked and the cage pockets were broken. Some rollers of this bearing were severely deformed resulting in macroscopic dimensional and shape changes. SEM examination showed that some areas of the roller surfaces, including the
372
Fig. 19. Photograph Fig. 20. Surface
of number
damage
after failure.
of a roller of number
Fig. 21. End face of a deformed Fig. 22. Cage material
2 bearing
2 bearing.
roller of number
(brass and silver) attached
2 bearing. to end face of a roller shown
in Fig. 21.
edges, were fractured. A typical example of the surface damage on the rollers is shown in Fig. 20. Damage was even more severe at the roller edges. The worst case example of this is shown in Fig. 21. This suggests that the observed gross deformation of the roller caused by heavy edge loading and rubbing occurred as a result of roller skewing. Energy-dispersive X-ray analysis showed that the end face of this roller was almost completely covered with material from the bearing cage. Figure 22 shows the particles of cage material (brass and silver) attached to the roller end face. This indicates that the roller ends were rubbing against the bearing cage under extreme pressure. This is an example of failure of high speed roller bearings caused by misalignment and development of excessive thrust load on the bearing due to shaft deflection. Examination of the rollers at higher magnifications showed that most surface areas were covered by small pits similar to those observed on the number 1 bearing. However, pitting was not a significant factor in the failure of this bearing. Hardness readings obtained from the roller ends showed a softening (3.5 HRC) in the average hardness of the rollers. This softening may have resulted from either formation of surface cracks or
Fig. 23. Wear damage
(pitting
Fig. 24. Inner ring raceway
Fig. 25. Details of damage Fig. 26. Surface
and scoring)
of number
on outer
2 bearing
on inner ring raceway
microstructure
of number
ring raceway
of number
2 bearing.
after failure.
of number
2 bearing.
1 bearing.
overtempering of surface material due to frictional heat, or a combination of both. Under visual examination, the centre of the outer ring raceway of this bearing appeared smooth but the areas along both sides of the raceway showed significant wear damage. Figure 23 shows the typical wear damage (pitting and scoring) observed on the outer ring raceway. The damaged inner ring raceway of this bearing is shown in Fig. 24. It can be seen that the failure of the inner ring raceway resulted from the removal of excessive surface material, producing a groove approximately 0.07 mm deep. An SEM micrograph, Fig. 25, shows the details of this groove. The observations suggest that the removal of raceway material resulted from extreme metal-to-metal contact (the ploughing action of the rollers on the raceway). Again, the cause of failure was misalignment produced by shaft deflection due to the increased shaming which also resulted in the movement of the inner ring on the shaft during failure. Visual inspection of the number 3 bearing after the run showed no significant damage. Examination of the bearing using SEM showed pits and score marks on the bearing components similar to those observed with the
Fig. 27. Cracks
in a roller of number
Fig.
material
28. White
(embedded
Fig. 29. Wear debris embedded Fig. 30. Friction bearing.
2 bearing. debris)
in surface
on the surface
material
layer with a fine microstructure
of a roller of number
of a roller of number on the surface
2 bearing.
2 bearing.
of a roller
from number
2
number 1 bearing. As observed with the number 1 bearing, the rollers suffered more surface damage (pitting and scoring) than either raceway. Damage was more severe at the roller edges than on the roller surfaces. Again, this was a result of roller skewing caused by shaft deflection. No microstructural changes were found in the components of the number 1 bearing. A typical surface microstructure of the components of this bearing is shown in Fig. 26. In addition, microhardness measurements obtained from the components of this bearing showed no distinct change below the surface. Hence metallographic examination and hardness measurements indicated no significant wear in this bearing. Many cracks were observed in the rollers of the number 2 bearing which failed at 720 h. Figure 27 shows the cracks on a section of a roller. A white etching material was also observed on some areas of the roller surfaces, particularly on the roller edges. Figure 28 shows the white etching material on the surface of an unplated roller. Examination of this material at higher magnifications suggested that it was not a part of the roller but was wear debris embedded in the surface material. Different wear debris particles
Fig. 31. White etching material on inner ring raceway of number 2 bearing. Fig. 32. SEM micrograph
of the same area shown in Fig. 31.
were also found on the roller surfaces. Figure 29 shows the heavily deformed wear debris particles of both the bearing steel (AISI 52100) and the brass cage embedded in the surface of a roller. The white etching material and the other embedded debris particles were found only in some surface regions. However, examination using SEM at higher magnifications revealed a distinct surface layer with a fine mi~rost~ctu~ in most surface areas of the rollers over a depth of approximately 6 pm below the surface. A typical example of this is shown in Fig. 30. No microstructural changes were found in the outer ring raceway of this bearing. However, a white etching material was observed on the inner ring raceway as shown in Fig. 31, An SEM micrograph, Fig. 32, of the same area suggests that the white etching material was not a part of the raceway but was wear debris embedded in the surface material as observed on the rollers of this bearing. Apart from the embedded wear debris, no distinct microstructural changes were found in the inner ring raceway. No microstructural or hardness changes were found in the components of the number 3 bearing. Hence no signific~t wear occurred in the number 3 bearing. 5. Discussion As shown in Table 1, the operating temperatures of the bearings increased slightly as the oil supply rates were reduced over the range 1 -0.25 1 min-I. This was expected because lubricating oil acts as a coolant for the bearings and hence a decreased coolant supply rate would result in increased bearing temperatures. However, below 0.25 1 min-i the operating temperatures of the bearings decreased with decreasing oil supply rate. In particular, when the oil supply to either the number 1 or number 3 bearing was completely cut off the operating temperature of each bearing reached the lowest level. It is well established that the load-independent friction torque increases with the amount of lubricant as a result of increased churning [7]. This can explain the higher temperatures at higher flow rates, despite the
376
increased cooling. The measured temperatures of the bearings, therefore, suggest that lowering the oil supply rate below a certain range (0.2 1 mini) or oil mist lubrication may be beneficial in reducing operating temperatures of high speed engine bearings. This observation is also in agreement with the results of Hargreaves and Higginson [8] who reported that friction torque and operating temperatures for cylindrical roller bearings could be greatly reduced by running at low lubricant supply rates or with oil mist lubrication. The number 1 and number 3 bearings showed no significant wear but the number 2 bearing failed completely. These observations may be explained in terms of the running conditions (loading and lubrication) and production of bearings including bearing materials. The number 1 bearing was sufficiently lubricated until the last 20 h. Although this bearing was the most heavily loaded and run dry for the last 20 h, it showed no evidence of significant wear, except minor surface pitting and scoring. Surface examination showed some evidence of roller skewing but the misalignment of this bearing was not high enough to cause severe wear or failure. Although the number 2 bearing was less heavily loaded than the number 1 bearing, it failed completely. The number 2 hearing occupies a critical point on the shaft because it constrains the shaft deflections produced by belt pull and shimming of the number 3 bearing housing. Shaft deflections due to the loading procedure produce relative race misalignments in the bearings. The calculated values of angular misalignment for the number 2 bearing are listed in Table 1. Excessive bearing misalignment due to increased shaft deflections may result in large skew motion of the rollers leading to seizure and failure. Large shaft deflections may also produce complex forces on the bearings during rotation of the shaft. One of these forces is the development of excessive thrust (axial) load on the bearings which in this case leads to an observed movement of the inner ring of the number 2 bearing on the shaft during failure. Development of thrust load on the number 2 bearing also resulted in fracture of the roller edges, the cage and edges of the guide flanges. Although cylindrical roller bearings can be used to carry combined radial and thrust loads [ 91, development of excessive thrust load may result in sudden bearing failure [ 31. This bearing was operated successfully for 700 h but failed suddenly at 720 h following a small increase (25 pm) in shimming of the number 3 bearing housing. The magnitude of the shaft deflection that shimming produced depends on the amount of shimming and the shaft length (distances between the bearings). Shaft deflection above a critical point, the value of which depends on the shaft material, shaft length, load and speed, may result in failure of either the shaft or the bearing, or both [ 31. In the present case this resulted in the failure of the number 2 bearing. This bearing was successfully operated with an angular misalignment of 0.0161” for 700 h but it failed in 20 h after the misalignment was increased to 0.0184“ by the increase in shimming. Hence the present work suggests that misalignment above a critical point (0.0161”) cannot be tolerated in these bearings,
371
The number 3 bearing was the least heavily loaded but was operated without an oil supply after 400 h. After completion of the run this bearing showed no evidence of significant wear, except minor pitting and scoring. SEM examination showed some evidence for roller skewing but neither the level of load nor the magnitude of misalignment was sufficient to cause serious damage. Although the bearing was run without an oil supply for 320 h, no distinct evidence of sliding or skidding was observed. This observation suggests that the oil mist produced from the lubricated bearings prevented sliding and metal-to-metal contact in this bearing. SEM examination showed that the rollers of the bearings suffered more pitting than the raceways. Since surface pitting can result from fatigue even under ideal operating conditions, it is possible that the rollers were subjected to either a higher number of stress cycles or a higher magnitude of stress per shaft revolution than either raceway. The number of stress cycles that bearing components undergo per shaft revolution depends on the bearing parameters, including the number of rollers and their locations. In high speed applications development of centrifugal forces may increase both the number and the level of contact stresses on the rollers and the outer ring raceway. In the present case the centrifugal forces acting on each roller (about 34 N) were insignificantly low and the contact stress on the outer raceway of each bearing was lower than the stress produced on the inner ring raceway because of contact geometry. However, the rollers and the outer raceway of each bearing would be expected to receive a higher number of contact stresses than the inner raceway, per shaft revolution, because of a continuously acting centrifugal force on each roller. In addition, the surfaces of the relatively small rollers would be expected to attain the highest temperatures in operation. Therefore the combination of both the higher number of stress cycles and the higher temperatures might be the reason for the observed heavier pitting of the roller surfaces. No distinct microstructural changes or friction layers were found on the components of either the number 1 or the number 3 bearing, although these bearings were run without oil supply for some period of time. Similar to the surface examination, this observation suggests that the oil mist lubrication prevented sliding and metal-to-metal contact in these bearings. Microstructural changes were found only on the rollers of the number 2 bearing which were made of AISI 52100 steel. These changes resulted in the formation of a distinct surface layer with a very fine microstructure. In addition, white etching wear debris was found embedded in the surface material of the rollers. Similar microstructural changes were also observed on the rollers and the raceways of a used high speed turbine engine bearing [ 51. The transformation products, including white etching material and the surface layer with a fine microstructure, were identified as friction martensite derived from the original microstructure of the bearing steel owing to the heat of friction. White etching wear debris was also found on the heavily damaged inner ring raceway of this bearing but this raceway showed no distinct microstructural changes. Comparison of the observations made on the
378
different bearings and the different bearing components suggested that AISI M-50 bearing steel is more resistant to friction-induced microstructural changes than AISI 52100 steel. Although the radial loads on the number 1 and the number 2 bearings were much higher than the dynamic loads which they normally carry in high speed engines (less than 550 N), no subsurface microstructural changes were found in any components of these bearings. Stress-induced subsurface microst~ctur~ changes may occur if the m~imum shear stresses exceed a threshold value approximately equal to the elastic limit of the bearing steel [ 41. For AISI 52100 bearing steel this value is reported to be 724 MPa [ 41. The calculated maximum shear stresses in the test bearings (577 MPa for the number 1, 454 MPa for the number 2 and 188 MPa for the number 3 bearing) were well below the proportional elastic limits of the bearing steels; between 1000 and 1600 MPa for M-50 [lo] and approx~ately ‘700 MPa for 52100 steel [ 4f. Thus mechanical wear or failure due to subsurface rolling contact fatigue would not be expected in the test bearings. To obtain shear stresses above the reported threshold value (724 MPa) in the components of the test bearings, the applied radial load on each bearing must be above 3000 N. No threshold value has been reported for AISI M-50 steel but a higher critical stress value would be expected for the stress-induced microstructural changes in this steel. Therefore, for radial loads below approximately 3000 N, subsurface rolling contact fatigue should not be a significant factor in the wear or failure of the test bearings. Ferrography, which enables monitoring of progressive changes, indicated no si~ific~t wear in the bearings until 650 h. Results of DR ferrography, and the ferrograms made of oil collected from the bearings during the first 650 h of run, indicated only a mild wear condition. However, after 650 h a more severe wear condition developed in the number 2 bearing. This was indicated by a change in the size and shape of wear debris particles collected from #is bearing at 700 h. Ferrograms made from oil samples collected from the number 1 and the number 3 bearings also indicated the change from mild to severe wear after 650 h but the compositional analysis of the wear debris showed that these particles originated from the number 2 bearing. The number 2 bearing failed suddenly at 720 h owing to increased shaft deflection (misalignment) which resulted in the movement of the inner ring on the shaft. During this period of time, ferrography provided sufficient evidence of the extreme operating condition by showing a drastic increase in the size and quantity of the wear particles from the number 2 bearing at 710 h. In addition, ferrograms showed the presence of relatively large silver particles in the lubricating oil at 710 h which indicated severe cage wear. Ferrographic analysis of the oil samples collected from the bearings and the tank at 720 h (during failure of the number 2 bearing) showed a large increase in the quantity of the wear debris particles. Compositional analysis of debris on the ferrograms showed that the wear particles were from all the components of the number 2 bearing, including rollers and raceways (52100
379
and M-50 steels) and cage material (brass and silver). Silver particles were present in oil collected after 700 h but brass particles were found only during failure of the number 2 bearing at 720 h. This suggests that the presence of silver particles in the oil indicates a severe wear condition in the bearings but the presence of brass particles from the cage is indicative of bearing failure. Present work suggests that under normal operating conditions wear of high speed engine bearings may successfully be monitored by ferrography but for abnormal wear conditions frequent oil sample collections are necessary. Spectrometric oil analysis (atomic absorption) indicated a change in wear status only at 720 h, after the number 2 bearing failed completely. Thus, in this case, ferrographic analysis or debris analysis provided an earlier indication of change in wear status than spectrometric oil analysis. Comparison of ferrographic results (debris analysis) with the worn surface examination suggests that normal wear particles (small platelets) resulted from pitting of rollers and the raceways of the bearings. Large and thick debris particles resulted from more severe wear (spalling and fracture of rollers and the inner raceway) of the number 2 bearing.
6. Conclusions
(1) Misalignment of high speed roller bearings above a critical point results in sudden bearing failure. (2) Rolling contact fatigue is not a significant wear or failure mechanism in high speed engine bearings. (3) Operating temperatures of high speed engine bearings may be decreased by reducing the oil supply rate or by running with oil mist lubrication. (4) AISI M-50 bearing steel is more resistant to friction-induced microstructural changes than AISI 52100 bearing steel. (5) Wear of high speed engine bearings may successfully be monitored by ferrography but as the wear condition accelerates more frequent oil collections are beneficial. (6) Ferrographic oil analysis provides an earlier indication of a change in wear condition than atomic absorption spectrometric oil analysis. (7) Small flat wear debris particles (platelets) from the bearing steel indicate a mild wear condition and large bulky particles indicate a severe wear mechanism in the test bearings. (8) The presence of silver particles in oil from the bearings indicates a severe wear condition but the presence of large brass particles (cage material) is indicative of bearing failure.
380
Acknowledgments The authors wish to thank Mr. C. Collings and Dr. A. W. Hanson of Atlantic Region Laboratory, National Research Council of Canada, for the SEM facilities. Mr. F. Roma, Mr. C. Jackson and Ms. M. Murphy provided technical assistance and Mrs. G. Veinot typed the manuscript. Finally, the authors would like to thank Professor G. P. Wilson for the management of the Contract and Dr. E. E. Laufer for discussions.
References 1 M. C. Shaw and F. Macks, Analysis and Lubrication of Bearings, McGraw-Hill, New York, 1949, Chapter 10. 2 H. J. Koeber, Design and calculation of high speed engine bearings, 60th AGARD Specialists Meet. on Aircraft Gear and Bearings, Tribological Systems, San Antonio, TX, April 1985, Paper 5. 3 T. Savaskan, E. E. Laufer and D. E. Veinot, Wear of high speed roller bearings, 60th AGARD Specialists Meet. on Aircraft Gear and Bearings, Tribological Systems, San Antonio, TX, April 1985, Paper 12. 4 J. A. Martin, S. F. Borgese and A. D. Eberdardt, Microstructural alterations of rolling steel undergoing cyclic stressing, J. Basic Eng., 59 (1966) 555 - 567. 5 T. Savaskan and E. E. Laufer, Wear in a high speed roller bearing, Met. Technol. NY, 11 (1984) 530 - 534. 6 Metals Handbook, Vol. 10, American Society for Metals, Metals Park, OH, 8th edn., 1975, pp. 416 - 436. 7 A. Sahlgren, in D. Dowson, C. M. Taylor, M. Godet and D. Berthe (eds.), Reference Speeds - A New Concept for Rolling Bearings, Proc. 7th Leeds-Lyon Symp., Leeds, September 9 - 12, 1980, Westbury House, Guildford, 1981. 8 R. H. Hargreaves and G. R. Higginson, Some effects of lubricant starvation in cylindrical roller bearings, J. Lubr. Technol., 98 (1976) 66 - 70. 9 S. R. Brown and S. Y. Poon, The lubrication of the roller-rib contacts of a radial cylindrical roller bearing carrying thrust load, ASLE Trans., 26 (3) (1983) 317 - 324. 10 J. A. Rescalvo and B. L. Averbach, Fracture and fatigue in M-50 and 18-4-l high speed steels, Metal. Trans., 1 OA (1979) 1265 - 1271.