AGMA 13FTM13-2013 Gear Failure Analysis and Lessons Learned in Aircraft High-Lift Actuation.pdf

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1、13FTM13 AGMA Technical Paper Gear Failure Analysis and Lessons Learned in Aircraft High-Lift Actuation By A. Wang, S. Gitnes, L. El-Bayoumy and J. Davies, MOOG Inc. Aircraft Group2 13FTM13 Gear Failure Analysis and Lessons Learned in Aircraft High-Lift Actuation Anngwo Wang, Seth Gitnes, Lotfi El-Ba

2、youmy and Jonathan Davies, MOOG Inc. Aircraft Group The statements and opinions contained herein are those of the author and should not be construed as an official action or opinion of the American Gear Manufacturers Association. Abstract Several gear failure cases and lessons learned in the develop

3、ment phase of aircraft high lift actuation systems are presented, including leading edge geared rotary actuators, and trailing edge geared rotary actuators, sector gears and pinions, and offset gearboxes. The high lift system of an aircraft, which contains trailing edge flaps and/or leading edge sla

4、ts, increases lift for takeoff, controls flight during cruise, and reduces speed while increasing lift for shorter landing distance. Many of these components contain highly loaded gears to increase the power to weight ratio. Because of requirements on weight or envelope and consideration of cost, th

5、e gears are always designed to the limit with reasonable margins of safety in a high lift system. The structure which supports the gears is limited in size and simplified, and the gear material and heat treatment are selected for easy manufacturing. Therefore, when misalignment and/or deflection of

6、the gears are large enough to cause reduction in tooth contact area, the stress on gears becomes large enough to cause damage. The failure modes can be classified as spalling or pitting at the location of concentrated loads. Most of the problems can be resolved by providing correct lead modification

7、 to alleviate the concentrated loading, while some need increase of the gear diameters, design modifications, or introduction of materials with higher allowable. Copyright 2013 American Gear Manufacturers Association 1001 N. Fairfax Street, Suite 500 Alexandria, Virginia 22314 September 2013 ISBN: 9

8、78-1-61481-070-4 3 13FTM13 Gear Failure Analysis and Lessons Learned in Aircraft High-Lift Actuation Anngwo Wang, Seth Gitnes, Lotfi El-Bayoumy and Jonathan Davies, MOOG Inc. Aircraft Group Introduction The high lift system of an aircraft, including trailing and/or leading edge slats/flaps, increase

9、s lift for take-off, controls flight during cruise, and reduces landing distance for touch-down. This flight control system is usually composed of power drive units (PDUs), torque tubes, bevel gear boxes, offset gearboxes, leading edge (LE) geared rotary actuators (GRAs), trailing edge (TE) GRAs, an

10、d leading edge sector gears and pinions (Figure 1). The system also includes other components, such as torque limiters, slip clutches, no-back drive devices, and wing tip brakes to provide system protection from overloading. Many of these components contain different types of gears that are usually

11、highly loaded to increase the power to weight ratio. Because of the requirements on weight or envelope and consideration of cost, the gears in a high lift system are always designed with minimal margins. The structure which supports the gears is limited in size or simplified, and the gear material a

12、nd heat treatment are selected for easy manufacturing. Deflections and misalignments between meshing gears cause edge loading which generates noise and high bending and contact stresses. The deflection emanates from the high loading and the misalignment comes from wing bending or the deflection of g

13、ear shafts and housings. Irrespective of the load, once the misalignment and/or deflection cause the contact area to shift and diminish, the stress becomes large enough to cause problems. AGMA publishes an atlas of failure modes 1 that identify some common types of wear. Bajpai et. al. 2 combines a

14、contact analysis model and a wear prediction model to describe the evaluation of tooth surface wear of spur and helical gear pairs. In these previous papers, the focus is on the outcome of failures, and not the reasons and sources of gear wear. Drago 3 discussed spalling caused by tip interference a

15、nd showed that a contact failure can lead to a more severe tooth fracture. Drago et. al. 4also discussed the relation between micropitting to specific lubrication additive package combinations, gear quality and surface finish, which can be resolved by profile modifications. Errichello et. al. 5 inve

16、stigated that the root cause of macropitting is the geometric stress concentration caused by tip-to-root interference. All these previous three papers deal with failures caused by tooth geometries and quality of gears. Errichello 6 investigated a gear set which failed due to lubrication breakdown. T

17、he main objective of this presentation is to find out how to solve these problems if the cause is due to edge loading from the misalignment of gear mounting and/or deflection of supporting structure. In this paper, several different gear failure cases in the development phase of high lift systems ar

18、e presented, including leading edge geared rotary actuators, and trailing edge geared rotary actuators, sector gears and pinions, and offset gearboxes. The failure modes can be classified as spalling or pitting at the location of concentrated loads. Most of the problems can be resolved by providing

19、correct lead modifications to alleviate the concentrated loading, while some can only be corrected by increasing the gear diameters, design modifications, or introduction of materials with higher allowable. Detailed analyses to predict deflections and misalignments on system and component levels is

20、the key to the amount of lead modification, from which increased local contact stresses can be calculated. Figure 1. Schematic of the high lift system of an aircraft (left LE shown, TE similar) 4 13FTM13 The cases presented in this paper are from the pre-production risk mitigation units (RMUs). Norm

21、ally the RMUs were designed with little margin to achieve minimum weight. Most system or component deflections can be predicted by analyses during the design phase. If there are any unanticipated misalignments or deflections, then failures will show up and the situation can be improved with correcti

22、ons in the production units. In this way designs are optimized for minimum weight. Leading edge rotary geared actuators The cross section of a typical leading edge geared rotary actuator is shown in Figure 2. There are three gear meshes on each one of the planet gears. The output is on the left side

23、 of the actuator. The output planet gears are overhung and balanced by the planet gears on the right. The center planet gears act as a pivot point. Because the output planet gears are overhung, it is called a cantilever GRA. Trailing edge geared rotary actuators The cross section of a typical traili

24、ng edge geared rotary actuator is shown in Figure 3, and the gear schematic is shown in Figure 4. The output consists of two load paths from two end ring gears. The sun gear drives the right end planet gears. The stiffness difference between the right and left load paths causes the compound planet g

25、ear to tilt. Thus, not only the planet gear loads due to meshing with the ring gears have to be considered, but also the misalignment from the two load paths needs to be included for selecting the optimum crowning which also reduces load mal-distribution across the gear tooth faces and increases loc

26、al contact stresses. Therefore, excessive crowning must be avoided. Figure 2. Typical leading edge GRA Figure 3. Typical trailing edge GRA Figure 4. Schematic of the output stage of a typical trailing edge GRA 5 13FTM13 Leading edge sector gears and pinions A typical leading edge sector gear and pin

27、ion set is shown in Figure 5. The pinion has to be crowned to allow for wing bending if a spherical bearing mount is not possible. This gear set is exposed to outside environment and grease or dry film lubrication may be depleted between service intervals. Crowning radii have to be optimized so that

28、 contact stresses can be minimized and the risk of running dry can be mitigated. From all the above applications in a high lift system, we can understand the importance of a good face load distribution between mating gear teeth. If misalignments become excessive, gears will suffer edge loading. Five

29、 different failure examples from pre-production risk mitigation units are the subject of the next section. Failure examples Example 1, Offset Idler Gear Set An offset gear set with one bearing very close to one end of the gearbox and another support far at the other end is shown in Figure 6. This is

30、 a very special layout, and is not the normal offset gearbox straddle mounted between two bearings. One can calculate the face load distribution according to AGMA 9. Because of the deflection, gears are edge loaded and pitted as shown in the figure. To solve the problem, the gears need to be crowned

31、 to accommodate this misalignment. The face width is 0.80 inch. The total slope including the deflection and manufacturing errors is 0.0048 in/in. After solving the equations as discussed in 7, the crowning radius is 91 inch, and the crowning center is at the end of the tooth. A contact stress of 26

32、6 ksi is calculated under the max tangential load of 1000 lbf. From Figure 7, we can see that the crowning has eliminated the pitting problem, so that the full tooth is now sharing the load. Figure 5. Typical leading edge sector gear and pinion set Figure 6. Damaged gear without crowning 7 6 13FTM13

33、 Figure 7. Good gear contact with crowning 7 Lessons learned The lesson learned in this case is that the location of bearing supports has significant impact on the load distribution of gears. Of course, it is better to have gears straddle mounted and sit just in the middle of two rigid supporting be

34、arings. But sometimes, because of the restriction from the structure of gearboxes, it is unavoidable to have lead modifications to relieve the situation. Another lesson learned is that if there is a large chamfer (Figure 8) on any member of the gears in mesh, the net face width at the lowest point o

35、f single tooth contact (LPSTC) should be used to calculate the contact stress, not the face width at the root. Using this net face for contact stresses is very easy to overlook at the early stage of design process. When the design is so critical, a small percentage can make a significant difference

36、on life calculations. The increased contact stress should be recalculated considering the localized load distribution from the crowned teeth. Example 2, Sector and Pinion Gear Set A sector and pinion gear set in Figure 5 must accommodate wing deflections. Because of being exposed to outside environm

37、ent, the contact stress must be low enough so that running the gears without grease is possible. Figure 8. Large chamfer on gears affects the net face width 7 13FTM13 The baseline design is regularly lubricated, and maximum allowable misalignment is 0.0015 in/in. The calculated contact stress is 312

38、 ksi with crowning under maximum operating loads. However, the test was done with no re-grease between service intervals. It is clear from the gear and pinion shown in Figure 9 and Figure 10 that although the contact pattern is localized, yet because of the higher contact stress due to crowning, ini

39、tial lubrication is gone after a while and soon after micro-pitting and rusting will result. Therefore, the goal is to make it lube-free. For a given misalignment, a new crowning radius and face width are proposed so that the stress is low enough to eliminate the need for re-lubrication in service.

40、The increased face width comes with a weight penalty, though. Another solution is to change to a material that has a higher contact allowable. Lessons learned A common engineering practice for calculating fatigue life is to use the mean load of a load spectrum. The mean load is the time averaged loa

41、d of a spectrum which contains multiple load sets (load vs. time) and is given by equation 1. The exponent m establishes the relationship between load (i.e., stress) vs. cycles to fail. Various exponents may be used to determine the mean load, but the most common are the followings: m = 3 for bearin

42、g wear analysis 8 and m = 9 for gear contact fatigue 9. 11/mmiiMEANTOTALkiLnLOADN(1) Equation 1 is known as the Lundberg-Palmgren equation used in the mean load calculation 8 for rolling bearings. This method can be extended to gear contact life analyses, where m = 9 is derived from the slope of pit

43、ting S-N logarithmic curves 9, and the square relation between loads and contact stresses. However, this mean load method may not be correct for the sector and pinion gear set, because the high load at one tooth cannot be shared by other teeth due to the limited stroke range. The lesson learned in t

44、his failure study is that unlike other types of gears the teeth get multiple hits from mating members, the sector gear gets certain loading with specific load conditions. A cumulative damage analysis, Equation 2 9 recommended by AGMA, is more suitable. Figure 9. Contact pattern on a sector gear Figu

45、re 10. Contact pattern of a crowned sector pinion 7 8 13FTM13 11.0ifikinN(2) where ni is the number of cycles at i -th stress level Nfiis the total number of cycles to failure at i-th level; k is the total number of different stress levels. Example 3, Cluster pinions in TE GRAs The cluster pinions i

46、n the input stage of a trailing edge GRA (Figure 3) failed at the pinion gear which mated with the ring (Z) gear. A free body diagram showing the force (red dashed line with arrows) acting on the gears is shown in Figure 11. Due to the deflection and internal clearances between bearings and the moun

47、ting housing, the cluster pinions tilt downwards and inwards on the pinions at the left. Therefore, the forces between the gear mesh of cluster pinions and ring (Z) gears are concentrated at the end of the cluster pinions. The evidence of tilting gears is also shown on the input sun pinion (Figure 1

48、2). Shown in Figure 13 is the spalling failure caused by edge loading on the end of the cluster pinions. Portions of the teeth then broke away from the origin of the spalling in a way very similar to what was described in 3. It starts from the inner bottom end of the teeth, then pitting/spalling pro

49、pagates up along the profile in the direction of sliding. The crack then turns in the direction of the line of action into the material, resulting in a chunk of metal separating from the teeth. Because the damage is only at one end of the teeth, the cause of failure is edge loading, instead of tip interference. The solution to this problem is to provide lead modifications on the cluster pinions similar to Example one, so that the load concentration can be alleviated. Another option is to use carburized materials, so that even with undesi