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Failure analysis of a pinion
Engineering Failure Analysis 12 (2005) 287–298 www.elsevier.com/locate/engfailanal
Failure analysis of a pinion C.R. Das, S.K. Albert, A.K. Bhaduri *, S.K. Ray Materials Technology Division, Indira Gandhi Centre for Atomic Research, Kalpakkam 603102, India Received 29 March 2004; accepted 30 March 2004 Available online 2 September 2004
Abstract A pinion, which was a part of an air-motor driving a diesel generator (DG) of a power plantÕs emergency powersupply system, was found to have failed into three parts during a routine check-up of the DG. This pinion, attached to the air-motor shaft is active only for a short duration during the start up, when the motion of the air-motor shaft is transferred to the ring gear attached to the DG which in turn starts the DG. Once the DG has started, the pinion disengages from the ring gear and the air-motor stops. The pinion that failed had been in operation for a long time and experienced several start-up operations before failure. A systematic failure analysis was carried out to find out the reasons for this unexpected fracture of the pinion during operation. The results indicate that the fracture was caused by fatigue with a fatigue crack initiating from the fillet of one of the pinion teeth. Due to misalignment that was present, the pinion did not mesh properly with the ring gear during the start-up operation and this led to a high stress concentration at the root of the pinion. The misalignment also led to severe wear and excessive heat generation at the mating surface. Misalignment, improper heat treatment, non-uniform distribution of sulphide inclusions and sharp corners would have all contributed to the failure of the pinion. Ó 2004 Elsevier Ltd. All rights reserved. Keywords: Gear failures; Misalignment; Inclusions; Fatigue crack; Heat treatment
1. Introduction Continuous operation of power plants necessitates the use of diesel generators (DGs) for providing emergency power supply during unexpected power failures. One of the common methods used for the start-up of these DGs is to use an air-motor that is driven by air released from an air compressor in the event of a power failure. A pinion attached to the shaft of an air-motor drives a large ring gear connected
*
Corresponding author. Tel.: +91 4114 280232; fax: +91 4114 280081. E-mail address: [email protected] (A.K. Bhaduri).
1350-6307/$ - see front matter Ó 2004 Elsevier Ltd. All rights reserved. doi:10.1016/j.engfailanal.2004.03.010
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to the DG. Under idling conditions, the pinion remains unengaged and the gap between the pinion and the ring gear is about 2.5–3.0 mm. During the start-up of the air-motor, the pinion receives a short burst of air that moves the pinion forward to mesh with the ring gear, after which the main air-valve opens for running the air-motor and driving the pinion, which in turn drives the large ring gear of the DG. Thus, before the air-motor applies full torque, the pinion rotates slowly as it meshes with the ring gear. To enable smooth meshing with the gears, the pinion has chamfered teeth to avoid abutment of the gear. Even if abutment occurs, the gear can ‘‘fall into’’ mesh as the air-motor begins to rotate. Once the DG has started-up, the air supply to the air-motor stops and the pinion is withdrawn to its idling position. Thus, the air-motor runs only for a short duration and the loading of the pinion is only for a few seconds during each startup operation. The failure, investigated in this paper, occurred while an air motor was being started up for a routine DG testing operation, and resulted in the DG failing to start. Although this pinion failure occurred during testing of the emergency power supply system, the consequences of a recurrence of this failure during an actual power failure could be serious. Hence, this failure analysis was carried out to identify the causes and possible remedies. For convenience, relevant nomenclature of a gear is shown in Fig. 1. The failed pinion had spur teeth, with fillet radius of 1.78 mm (0.07 in.) on the outer side for meshing with the ring gear of the DG, and had spline helical teeth, without any radius in the fillet on the inner side for meshing with the air-motor shaft. It had 12 spur teeth, with a pitch circle diameter of 50.8 mm (2 in.) and outer diameter (OD) of 57.15 mm (2.25 in.), and was designed for a torque of 340 N m (250 ft lb) at 1200 rpm. The ring gear had 318 spur teeth, with a pitch circle diameter of 1346.2 mm (53 in.) and OD of 1352.6 (53.25 in.). The photographs of the failed pinion (Figs. 2 and 3) show that the crack had initiated preferentially along the root of the teeth, where stress concentration was high, and then propagated. The arrow in Fig. 2 points to one of the fracture locations, while the arrows in Fig. 3 indicate the crack initiation points at the fracture surface.
Fig. 1. Nomenclature of the gear teeth.
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Fig. 2. Assembly of fracture pinion.
Fig. 3. Fracture piece of pinion.
2. Investigations 2.1. Visual examination Visual examination of the fracture surfaces of the failed pinion, after ultrasonic cleaning with acetone, showed the presence of burn marks on the edge of a spur tooth (Fig. 4) and on the upper land of a helical tooth (Fig. 5). It was further observed that the front of the spur tooth had a smooth worn-out surface, while
Fig. 4. Burn mark in one edge of the gear teeth.
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Fig. 5. Gouging mark on the land of the spline helical tooth of the pinion.
Fig. 6. Contact and non-contacted area on a spur tooth surface.
the rear side had a rough surface indicating improper contact with the ring gear during operation (Fig. 6). Gouging marks were also observed on the land of the spline helical tooth (Fig. 7). Small-size specimens were cut from location D in the spur teeth (Fig. 4) and prepared for metallographic examination.
Fig. 7. Scouring mark on the helical tooth.
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2.2. Material identification As no information with respect to the chemical composition and the heat treatment condition of the pinion material was available with the plant, the first task in the failure analysis was the material identification. Chemical analysis carried out using a direct-reading optical emission spectrometer at several locations of the broken pinion indicated that the pinion material was AISI 8620 steel (Table 1), which is usually used for gears in the surface-hardened condition. The elemental contents at different locations of the specimen from the failed pinion were similar for all elements, except sulphur, which although within the specified limit, varied between 0.020 and 0.033 wt%. 2.3. Microstructural examination The microstructure of the gear-tooth near the surface (Fig. 8) showed a dark etched region close to the pinion surface and a lightly etched region away from it, clearly indicating that the pinion material was surface hardened. The high carbon content at the case caused faster etching in this region, and therefore it was confirmed that the pinion material was case carburised. The dark etching also precludes case nitriding, as the nitrided layer would have shown up as a white layer [1]. Hence the gear material was confirmed as case hardened AISI 8620 steel. Optical microscopic examination of metallographically polished surfaces, in the unetched condition, of specimens taken from the burnt area (location D in Fig. 4) revealed the presence of an array of stringer-type inclusions (Fig. 9). From the morphology of the inclusions these could be identified as sulphide inclusions. The volume fraction of these sulphide inclusions, estimated using the point-count method [2], varied between 2% and 11% at different locations. Microcracks were also observed in specimens obtained from the failed pinion from regions close to the location of fracture (at location D). One of these microcracks initiated from the root of the spur tooth where the hardened surface had been damaged (Fig. 10). Similar microcracks were also observed at the sharp corners of the helical spline tooth in the inner diameter (ID) of the pinion, with one of the microcracks initiating from the fillet of the helical spline tooth (Fig. 11). 2.4. Hardness The hardness of the case and core of the failed pinion, measured using a Rockwell hardness tester with 150 kg load, was 41–55 RC at the helical spline gear surface, 53–55 RC at the spur gear surface, and 41 RC at the core. This indicated that while case hardening was proper on the spur gear side, uniform case hardening was not achieved on the helical spline gear side. Microhardness measurements, using a Shimadzu Microhardness tester HMV 2000 with 200 gm load, across the cross-section of the failed pinion (at locations indicated in Fig. 11) showed that hardness of the surface was clearly different for the inner (helical) tooth and outer (spur) tooth of the pinion (Fig. 12). Also,
Table 1 Chemical composition of the pinion gear material (determined at different locations) Elements wt%
C 0.206–0.208
Mn 0.80–0.82
P 0.018
S 0.020–0.033
Si 0.29–0.30
Al 0.02–0.25
Cr 0.68–0.69
Elements wt%
Cu 0.104–0.106
Ni 0.44–0.45
Mo 0.17–0.18
Co <0.03
Nb <0.008
Ti <0.003
V <0.14
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Fig. 8. Microstructure of the case and core in the spur tooth.
Fig. 9. Micrograph showing typical sulphide inclusions.
Fig. 10. Typical crack from the root of the spur tooth.
the hardness variation along the pitch line (i.e. across the cross-section of the spur tooth) indicated that the surface hardening across the active and inactive flanks of the teeth was fairly symmetrical, with very little variation among the different teeth (Fig. 13).
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Fig. 11. Typical root crack from the helical spline tooth.
Microhardness (VHN200)
1000
root of a sure gear root of a helical gear
900 800 700 600 500
-800 -600 -400 -200
0
200 400 600
Distance along the the root ( µm) Fig. 12. Hardness distribution along the roots.
Hardness (VHN200gm)
850
tooth-1 tooth-2
800 750 700 650 600 550 0
1
2 3 4 Distance (mm)
5
Fig. 13. Hardness distribution across the spur tooth thickness.
293
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2.5. Fracture surface Fractography, using a scanning electron microscope (SEM), of the fracture surface (Fig. 4) showed a river-like pattern indicating a fatigue-type [3] of fracture (Fig. 14). While a fibrous fracture surface along with dimples was observed at the centre of the hub (Fig. 15), an intergranular mode of fracture was observed at the rear end of the hub (Fig. 16). Additionally, the presence of extensive sub-surface cracks [4] at the loading surface of the spur tooth (Fig. 17) was an indication that during operation the spur tooth was subjected to a contact stress that would have been high enough to initiate a fatigue crack at its root. The presence of a high contact stress was supported by the damaged surface (Fig. 18) in the burnt region. SEM fractography showed that the fracture morphology changed significantly from the front to the rear of the pinion hub. At the edges with the fracture was intergranular, while in the centre the fracture was predominantly ductile with long fibre-like inclusions along the fracture surface. Although the pinion material was essentially ductile, the presence of elongated sulphur stringers could have assisted the fracture process. The intergranular fracture mode at the edge of the pinion can be attributed to the presence of a large number of carbides present along the prior austenite grain boundaries of the carburised zone.
3. Stress calculations The stress experienced by the spur tooth during operation was estimated using the design torque of 340 N m (250 ft lb). The bending stress on the loaded tooth can be calculated using the equation r¼
W tp ; K v FJ
ð1Þ
where Wt is the transmitted load, p is the circular pitch, Kv is the dynamic factor, F is the pinion height and J is the geometry factor [5]. The torque T is given by T ¼ d2W t ;
ð2Þ
where d is the pinion diameter. Using T = 340 N m and d = 50.8 mm (2 in) in Eq. (2), the value of Wt is calculated as 13.4 kN. Using p = 0.24 mm1, (12 teeth in 2 in diameter), Kv = 0.816, F = 28 mm and J = 0.355 in Eq. (1), the bending stress r was calculated to be 396 MPa. Although the pinion height F
Fig. 14. Typical brittle fracture observed in the failed region.
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Fig. 15. Typical fracture surface in the high sulphur steel.
Fig. 16. Intergranular fracture surface in the edge side of the hub.
Fig. 17. Rolling-contact fatigue.
was 28 mm, examination of the tooth surface showed that the actual contact height of the tooth was only 22 mm. With this value of F = 22 mm (instead of 28 mm), the actual bending stress r was estimated to be 505 MPa, i.e., 109 MPa more than the design bending stress of 396 MPa.
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Fig. 18. Presence of sliding marks on the burnt surface.
Also, the ultimate tensile strength (UTS) and the endurance limit for high cycle fatigue conditions for the pinion material was estimated from the measured hardness of the material. The hardness at the surface and core of the pinion was measured as RC55 (or 564 BHN) and RC41 (or 379 BHN), respectively. The tensile strength can be approximately calculated from the Brinell hardness (BHN) using the equation [6] UTS 3:4 BHN:
ð3Þ
From Eq. (3), the approximate UTS values for the hardened surface and the core was calculated to be 1917 and 1289 MPa, respectively. Using these UTS values, the Endurance Limit Se of a gear tooth can be calculated using the simplified equation [5] S e ¼ K a K b K c K d K e K f S 0e ;
ð4Þ
where the surface factor Ka = 0.62, size factor Kb = 0.93 (for module 4), reliability factor Kc = 0.868 (for 95% reliability), temperature factor Kd = 1 (assumed), modifying factor for stress concentration Ke = 1, miscellaneous-effect factor Kf = 1.455 (for UTS of 1917 MPa of the hardened surface) and 1.3 (for UTS of 1289 MPa of the core), and the endurance limit of a rotating-beam specimen S 0e ¼ 0:5 UTS ¼ 958:5 MPa (for the hardened surface) and 644.5 MPa (for the core). Using these values in Eq. (4), the endurance limit Se was estimated to be 697 and 419 MPa for hardened surface and the core of the pinion, respectively. For the properly heat-treated AISI 8620 pinion material, the specified UTS is 1960 MPa [7] for which the achievable endurance limit is 665 MPa. This is considerably higher than the endurance limit of 419 MPa estimated above from the hardness measured on the pinion material. This discrepancy, between reported UTS of the material and the actually estimated from the hardness is too high to be ascribed to errors in the estimates and strongly suggests that the heat treatment given to the failed pinion was inadequate. Additionally, the reduced contact area on the pinion teeth (of 22 mm instead of 28 mm) would have resulted in the teeth experiencing a bending stress of 505 MPa (under design torque conditions), which is higher compared to both the pinionÕs design bending stress of 396 MPa, and the estimated endurance limit for the core of 419 MPa. These material and/or contact deficiencies could also have contributed to the failure of this pinion. In fact many failures have been reported due to improper heat treatment [8].
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4. Analysis of results From the above investigations, the following possible causes of failure of the pinion could be arrived at: (1) Though the pinion operates only for a few seconds each time the DG is switched on, the pinion would have experienced a large number of cyclic loadings cumulatively during its past service. (2) The evidence of reduced contact area on the spur tooth surface (Fig. 6) indicated that improper meshing occurred between the pinion gear and the ring gear of the DG. This led to the pinion being subjected to higher bending stress [3] than the endurance limit of the core of the pinion. The inference regarding high contact stress on the pinion tooth was further supported by the observation of a significant amount of sub-surface cracks at the loading surface (Fig. 17). (3) Inadequate heat treatment of the pinion could have also resulted in a lower endurance limit of the pinion core than that achievable in the chosen pinion material. (4) Under the above conditions, a fatigue crack would have initiated at the root on the active flank of the spur gear tooth (shown in Figs. 4 and 14) as it would have been experiencing a bending stress exceeding its endurance limit. Fractographic evidence of a river-like pattern (Fig. 14) suggests the initiation of fatigue cracks in this region. (5) The presence of elongated sulphide inclusions in the pinion would have further reduced the fracture resistance of the pinion material. Also, the presence of sharp fillets in the helical spline (Fig. 11) would have enhanced crack initiation. (6) Misalignment would have also led to the generation of high stresses that caused gouging/ scouring at the land of the helical spline tooth (Figs. 5 and 7) and also excessive heat generation as evidenced from the burn marks (Figs. 4 and 5). All the above factors, mostly resulting from improper heat treatment, presence of sulphide inclusions in the root location and importantly high stress at the pinion teeth resulting from improper meshing of the pinion with the ring gear during operation contributed to the ultimate fatigue failure of the pinion.
5. Suggested remedies To avoid premature failure of the pinion, the following remedial measures have been suggested. 1. Thorough inspection and perfect alignment of the pinion with respect to the ring gear during routine maintenance would ensure proper meshing during its forward movement at the time of start-up of the air motor. 2. Provision for monitoring the torque experienced by the motor during its operation could give indications regarding improper functioning of the pinion. 3. Sharp corners that raise the stress concentration factor [9] should be avoided for reducing the bending stress on the pinion teeth. 4. Proper choice of pinion material and/or proper heat treatment of the core and uniform hardening of the case for improving the mechanical properties of the pinion core and reducing non-uniform contact fatigue of the spur gear tooth should be considered. 5. Use of steel with a low sulphur content and with a uniform distribution of inclusions is preferred for this application as sulphide inclusions, as observed in the pinion material, can adversely affect the fracture toughness of the steel.
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Acknowledgements The authors thank Mr. G. Ashok (MAPS, Kalpakkam) for his help in understanding the case history of the failed component. The authors would also like to acknowledge the support of their colleague Mrs. R. Radhika during the SEM fractography studies.
References [1] [2] [3] [4] [5] [6] [7] [8] [9]
Lakhtin Yu. Engineering physical metallurgy and heat treatment. Mir Publishers; 1990. p. 274–5. Vander Voort George F. Metallography principles and practice. New York: Mc-Graw Hill; 1999. p. 426–8. Fernandes PJL. Tooth bending fatigue failures in gears. Eng Fail Anal 1996;3(3):219–25. Guichelaar PJ, Lery Bernard S, Parikh NM. Gear manufactur and performance. Am Soc Metals 1974:27. Shigley JE. Mechanical engineeging design. Matrix ed. New York: McGraw-Hill Book Company; 1986. p. 467–508. Dieter GE. Mechanical metallurgy. SI Matrix ed. New York: McGraw Hill bok Company; 1988. p. 331. Buch A. Fatigue data handbook. Trans Tech Publisher; 1998. p. 38 [chapter 1]. ASM metal handbook, failure of gears, failure analysis and prevention, vol. 10, 8th ed. 1975. p. 507–24. Peterson RE. Stress concentration factors. New York: Wiley; 1974. p. 270–1.
Failure analysis of a pinion C.R. Das, S.K. Albert, A.K. Bhaduri *, S.K. Ray Materials Technology Division, Indira Gandhi Centre for Atomic Research, Kalpakkam 603102, India Received 29 March 2004; accepted 30 March 2004 Available online 2 September 2004
Abstract A pinion, which was a part of an air-motor driving a diesel generator (DG) of a power plantÕs emergency powersupply system, was found to have failed into three parts during a routine check-up of the DG. This pinion, attached to the air-motor shaft is active only for a short duration during the start up, when the motion of the air-motor shaft is transferred to the ring gear attached to the DG which in turn starts the DG. Once the DG has started, the pinion disengages from the ring gear and the air-motor stops. The pinion that failed had been in operation for a long time and experienced several start-up operations before failure. A systematic failure analysis was carried out to find out the reasons for this unexpected fracture of the pinion during operation. The results indicate that the fracture was caused by fatigue with a fatigue crack initiating from the fillet of one of the pinion teeth. Due to misalignment that was present, the pinion did not mesh properly with the ring gear during the start-up operation and this led to a high stress concentration at the root of the pinion. The misalignment also led to severe wear and excessive heat generation at the mating surface. Misalignment, improper heat treatment, non-uniform distribution of sulphide inclusions and sharp corners would have all contributed to the failure of the pinion. Ó 2004 Elsevier Ltd. All rights reserved. Keywords: Gear failures; Misalignment; Inclusions; Fatigue crack; Heat treatment
1. Introduction Continuous operation of power plants necessitates the use of diesel generators (DGs) for providing emergency power supply during unexpected power failures. One of the common methods used for the start-up of these DGs is to use an air-motor that is driven by air released from an air compressor in the event of a power failure. A pinion attached to the shaft of an air-motor drives a large ring gear connected
*
Corresponding author. Tel.: +91 4114 280232; fax: +91 4114 280081. E-mail address: [email protected] (A.K. Bhaduri).
1350-6307/$ - see front matter Ó 2004 Elsevier Ltd. All rights reserved. doi:10.1016/j.engfailanal.2004.03.010
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to the DG. Under idling conditions, the pinion remains unengaged and the gap between the pinion and the ring gear is about 2.5–3.0 mm. During the start-up of the air-motor, the pinion receives a short burst of air that moves the pinion forward to mesh with the ring gear, after which the main air-valve opens for running the air-motor and driving the pinion, which in turn drives the large ring gear of the DG. Thus, before the air-motor applies full torque, the pinion rotates slowly as it meshes with the ring gear. To enable smooth meshing with the gears, the pinion has chamfered teeth to avoid abutment of the gear. Even if abutment occurs, the gear can ‘‘fall into’’ mesh as the air-motor begins to rotate. Once the DG has started-up, the air supply to the air-motor stops and the pinion is withdrawn to its idling position. Thus, the air-motor runs only for a short duration and the loading of the pinion is only for a few seconds during each startup operation. The failure, investigated in this paper, occurred while an air motor was being started up for a routine DG testing operation, and resulted in the DG failing to start. Although this pinion failure occurred during testing of the emergency power supply system, the consequences of a recurrence of this failure during an actual power failure could be serious. Hence, this failure analysis was carried out to identify the causes and possible remedies. For convenience, relevant nomenclature of a gear is shown in Fig. 1. The failed pinion had spur teeth, with fillet radius of 1.78 mm (0.07 in.) on the outer side for meshing with the ring gear of the DG, and had spline helical teeth, without any radius in the fillet on the inner side for meshing with the air-motor shaft. It had 12 spur teeth, with a pitch circle diameter of 50.8 mm (2 in.) and outer diameter (OD) of 57.15 mm (2.25 in.), and was designed for a torque of 340 N m (250 ft lb) at 1200 rpm. The ring gear had 318 spur teeth, with a pitch circle diameter of 1346.2 mm (53 in.) and OD of 1352.6 (53.25 in.). The photographs of the failed pinion (Figs. 2 and 3) show that the crack had initiated preferentially along the root of the teeth, where stress concentration was high, and then propagated. The arrow in Fig. 2 points to one of the fracture locations, while the arrows in Fig. 3 indicate the crack initiation points at the fracture surface.
Fig. 1. Nomenclature of the gear teeth.
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Fig. 2. Assembly of fracture pinion.
Fig. 3. Fracture piece of pinion.
2. Investigations 2.1. Visual examination Visual examination of the fracture surfaces of the failed pinion, after ultrasonic cleaning with acetone, showed the presence of burn marks on the edge of a spur tooth (Fig. 4) and on the upper land of a helical tooth (Fig. 5). It was further observed that the front of the spur tooth had a smooth worn-out surface, while
Fig. 4. Burn mark in one edge of the gear teeth.
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Fig. 5. Gouging mark on the land of the spline helical tooth of the pinion.
Fig. 6. Contact and non-contacted area on a spur tooth surface.
the rear side had a rough surface indicating improper contact with the ring gear during operation (Fig. 6). Gouging marks were also observed on the land of the spline helical tooth (Fig. 7). Small-size specimens were cut from location D in the spur teeth (Fig. 4) and prepared for metallographic examination.
Fig. 7. Scouring mark on the helical tooth.
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2.2. Material identification As no information with respect to the chemical composition and the heat treatment condition of the pinion material was available with the plant, the first task in the failure analysis was the material identification. Chemical analysis carried out using a direct-reading optical emission spectrometer at several locations of the broken pinion indicated that the pinion material was AISI 8620 steel (Table 1), which is usually used for gears in the surface-hardened condition. The elemental contents at different locations of the specimen from the failed pinion were similar for all elements, except sulphur, which although within the specified limit, varied between 0.020 and 0.033 wt%. 2.3. Microstructural examination The microstructure of the gear-tooth near the surface (Fig. 8) showed a dark etched region close to the pinion surface and a lightly etched region away from it, clearly indicating that the pinion material was surface hardened. The high carbon content at the case caused faster etching in this region, and therefore it was confirmed that the pinion material was case carburised. The dark etching also precludes case nitriding, as the nitrided layer would have shown up as a white layer [1]. Hence the gear material was confirmed as case hardened AISI 8620 steel. Optical microscopic examination of metallographically polished surfaces, in the unetched condition, of specimens taken from the burnt area (location D in Fig. 4) revealed the presence of an array of stringer-type inclusions (Fig. 9). From the morphology of the inclusions these could be identified as sulphide inclusions. The volume fraction of these sulphide inclusions, estimated using the point-count method [2], varied between 2% and 11% at different locations. Microcracks were also observed in specimens obtained from the failed pinion from regions close to the location of fracture (at location D). One of these microcracks initiated from the root of the spur tooth where the hardened surface had been damaged (Fig. 10). Similar microcracks were also observed at the sharp corners of the helical spline tooth in the inner diameter (ID) of the pinion, with one of the microcracks initiating from the fillet of the helical spline tooth (Fig. 11). 2.4. Hardness The hardness of the case and core of the failed pinion, measured using a Rockwell hardness tester with 150 kg load, was 41–55 RC at the helical spline gear surface, 53–55 RC at the spur gear surface, and 41 RC at the core. This indicated that while case hardening was proper on the spur gear side, uniform case hardening was not achieved on the helical spline gear side. Microhardness measurements, using a Shimadzu Microhardness tester HMV 2000 with 200 gm load, across the cross-section of the failed pinion (at locations indicated in Fig. 11) showed that hardness of the surface was clearly different for the inner (helical) tooth and outer (spur) tooth of the pinion (Fig. 12). Also,
Table 1 Chemical composition of the pinion gear material (determined at different locations) Elements wt%
C 0.206–0.208
Mn 0.80–0.82
P 0.018
S 0.020–0.033
Si 0.29–0.30
Al 0.02–0.25
Cr 0.68–0.69
Elements wt%
Cu 0.104–0.106
Ni 0.44–0.45
Mo 0.17–0.18
Co <0.03
Nb <0.008
Ti <0.003
V <0.14
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Fig. 8. Microstructure of the case and core in the spur tooth.
Fig. 9. Micrograph showing typical sulphide inclusions.
Fig. 10. Typical crack from the root of the spur tooth.
the hardness variation along the pitch line (i.e. across the cross-section of the spur tooth) indicated that the surface hardening across the active and inactive flanks of the teeth was fairly symmetrical, with very little variation among the different teeth (Fig. 13).
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Fig. 11. Typical root crack from the helical spline tooth.
Microhardness (VHN200)
1000
root of a sure gear root of a helical gear
900 800 700 600 500
-800 -600 -400 -200
0
200 400 600
Distance along the the root ( µm) Fig. 12. Hardness distribution along the roots.
Hardness (VHN200gm)
850
tooth-1 tooth-2
800 750 700 650 600 550 0
1
2 3 4 Distance (mm)
5
Fig. 13. Hardness distribution across the spur tooth thickness.
293
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2.5. Fracture surface Fractography, using a scanning electron microscope (SEM), of the fracture surface (Fig. 4) showed a river-like pattern indicating a fatigue-type [3] of fracture (Fig. 14). While a fibrous fracture surface along with dimples was observed at the centre of the hub (Fig. 15), an intergranular mode of fracture was observed at the rear end of the hub (Fig. 16). Additionally, the presence of extensive sub-surface cracks [4] at the loading surface of the spur tooth (Fig. 17) was an indication that during operation the spur tooth was subjected to a contact stress that would have been high enough to initiate a fatigue crack at its root. The presence of a high contact stress was supported by the damaged surface (Fig. 18) in the burnt region. SEM fractography showed that the fracture morphology changed significantly from the front to the rear of the pinion hub. At the edges with the fracture was intergranular, while in the centre the fracture was predominantly ductile with long fibre-like inclusions along the fracture surface. Although the pinion material was essentially ductile, the presence of elongated sulphur stringers could have assisted the fracture process. The intergranular fracture mode at the edge of the pinion can be attributed to the presence of a large number of carbides present along the prior austenite grain boundaries of the carburised zone.
3. Stress calculations The stress experienced by the spur tooth during operation was estimated using the design torque of 340 N m (250 ft lb). The bending stress on the loaded tooth can be calculated using the equation r¼
W tp ; K v FJ
ð1Þ
where Wt is the transmitted load, p is the circular pitch, Kv is the dynamic factor, F is the pinion height and J is the geometry factor [5]. The torque T is given by T ¼ d2W t ;
ð2Þ
where d is the pinion diameter. Using T = 340 N m and d = 50.8 mm (2 in) in Eq. (2), the value of Wt is calculated as 13.4 kN. Using p = 0.24 mm1, (12 teeth in 2 in diameter), Kv = 0.816, F = 28 mm and J = 0.355 in Eq. (1), the bending stress r was calculated to be 396 MPa. Although the pinion height F
Fig. 14. Typical brittle fracture observed in the failed region.
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Fig. 15. Typical fracture surface in the high sulphur steel.
Fig. 16. Intergranular fracture surface in the edge side of the hub.
Fig. 17. Rolling-contact fatigue.
was 28 mm, examination of the tooth surface showed that the actual contact height of the tooth was only 22 mm. With this value of F = 22 mm (instead of 28 mm), the actual bending stress r was estimated to be 505 MPa, i.e., 109 MPa more than the design bending stress of 396 MPa.
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Fig. 18. Presence of sliding marks on the burnt surface.
Also, the ultimate tensile strength (UTS) and the endurance limit for high cycle fatigue conditions for the pinion material was estimated from the measured hardness of the material. The hardness at the surface and core of the pinion was measured as RC55 (or 564 BHN) and RC41 (or 379 BHN), respectively. The tensile strength can be approximately calculated from the Brinell hardness (BHN) using the equation [6] UTS 3:4 BHN:
ð3Þ
From Eq. (3), the approximate UTS values for the hardened surface and the core was calculated to be 1917 and 1289 MPa, respectively. Using these UTS values, the Endurance Limit Se of a gear tooth can be calculated using the simplified equation [5] S e ¼ K a K b K c K d K e K f S 0e ;
ð4Þ
where the surface factor Ka = 0.62, size factor Kb = 0.93 (for module 4), reliability factor Kc = 0.868 (for 95% reliability), temperature factor Kd = 1 (assumed), modifying factor for stress concentration Ke = 1, miscellaneous-effect factor Kf = 1.455 (for UTS of 1917 MPa of the hardened surface) and 1.3 (for UTS of 1289 MPa of the core), and the endurance limit of a rotating-beam specimen S 0e ¼ 0:5 UTS ¼ 958:5 MPa (for the hardened surface) and 644.5 MPa (for the core). Using these values in Eq. (4), the endurance limit Se was estimated to be 697 and 419 MPa for hardened surface and the core of the pinion, respectively. For the properly heat-treated AISI 8620 pinion material, the specified UTS is 1960 MPa [7] for which the achievable endurance limit is 665 MPa. This is considerably higher than the endurance limit of 419 MPa estimated above from the hardness measured on the pinion material. This discrepancy, between reported UTS of the material and the actually estimated from the hardness is too high to be ascribed to errors in the estimates and strongly suggests that the heat treatment given to the failed pinion was inadequate. Additionally, the reduced contact area on the pinion teeth (of 22 mm instead of 28 mm) would have resulted in the teeth experiencing a bending stress of 505 MPa (under design torque conditions), which is higher compared to both the pinionÕs design bending stress of 396 MPa, and the estimated endurance limit for the core of 419 MPa. These material and/or contact deficiencies could also have contributed to the failure of this pinion. In fact many failures have been reported due to improper heat treatment [8].
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4. Analysis of results From the above investigations, the following possible causes of failure of the pinion could be arrived at: (1) Though the pinion operates only for a few seconds each time the DG is switched on, the pinion would have experienced a large number of cyclic loadings cumulatively during its past service. (2) The evidence of reduced contact area on the spur tooth surface (Fig. 6) indicated that improper meshing occurred between the pinion gear and the ring gear of the DG. This led to the pinion being subjected to higher bending stress [3] than the endurance limit of the core of the pinion. The inference regarding high contact stress on the pinion tooth was further supported by the observation of a significant amount of sub-surface cracks at the loading surface (Fig. 17). (3) Inadequate heat treatment of the pinion could have also resulted in a lower endurance limit of the pinion core than that achievable in the chosen pinion material. (4) Under the above conditions, a fatigue crack would have initiated at the root on the active flank of the spur gear tooth (shown in Figs. 4 and 14) as it would have been experiencing a bending stress exceeding its endurance limit. Fractographic evidence of a river-like pattern (Fig. 14) suggests the initiation of fatigue cracks in this region. (5) The presence of elongated sulphide inclusions in the pinion would have further reduced the fracture resistance of the pinion material. Also, the presence of sharp fillets in the helical spline (Fig. 11) would have enhanced crack initiation. (6) Misalignment would have also led to the generation of high stresses that caused gouging/ scouring at the land of the helical spline tooth (Figs. 5 and 7) and also excessive heat generation as evidenced from the burn marks (Figs. 4 and 5). All the above factors, mostly resulting from improper heat treatment, presence of sulphide inclusions in the root location and importantly high stress at the pinion teeth resulting from improper meshing of the pinion with the ring gear during operation contributed to the ultimate fatigue failure of the pinion.
5. Suggested remedies To avoid premature failure of the pinion, the following remedial measures have been suggested. 1. Thorough inspection and perfect alignment of the pinion with respect to the ring gear during routine maintenance would ensure proper meshing during its forward movement at the time of start-up of the air motor. 2. Provision for monitoring the torque experienced by the motor during its operation could give indications regarding improper functioning of the pinion. 3. Sharp corners that raise the stress concentration factor [9] should be avoided for reducing the bending stress on the pinion teeth. 4. Proper choice of pinion material and/or proper heat treatment of the core and uniform hardening of the case for improving the mechanical properties of the pinion core and reducing non-uniform contact fatigue of the spur gear tooth should be considered. 5. Use of steel with a low sulphur content and with a uniform distribution of inclusions is preferred for this application as sulphide inclusions, as observed in the pinion material, can adversely affect the fracture toughness of the steel.
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Acknowledgements The authors thank Mr. G. Ashok (MAPS, Kalpakkam) for his help in understanding the case history of the failed component. The authors would also like to acknowledge the support of their colleague Mrs. R. Radhika during the SEM fractography studies.
References [1] [2] [3] [4] [5] [6] [7] [8] [9]
Lakhtin Yu. Engineering physical metallurgy and heat treatment. Mir Publishers; 1990. p. 274–5. Vander Voort George F. Metallography principles and practice. New York: Mc-Graw Hill; 1999. p. 426–8. Fernandes PJL. Tooth bending fatigue failures in gears. Eng Fail Anal 1996;3(3):219–25. Guichelaar PJ, Lery Bernard S, Parikh NM. Gear manufactur and performance. Am Soc Metals 1974:27. Shigley JE. Mechanical engineeging design. Matrix ed. New York: McGraw-Hill Book Company; 1986. p. 467–508. Dieter GE. Mechanical metallurgy. SI Matrix ed. New York: McGraw Hill bok Company; 1988. p. 331. Buch A. Fatigue data handbook. Trans Tech Publisher; 1998. p. 38 [chapter 1]. ASM metal handbook, failure of gears, failure analysis and prevention, vol. 10, 8th ed. 1975. p. 507–24. Peterson RE. Stress concentration factors. New York: Wiley; 1974. p. 270–1.