JP4406267B2 - Torque transmission coupling - Google Patents

Description

  The present invention relates to a torque transmission coupling of an automobile.

  An example of such a conventional torque transmission coupling is shown in FIG. FIG. 26 shows a sectional view of the transfer of the four-wheel drive vehicle. The transfer 201 includes a torque transmission coupling 203. The torque transmission coupling 203 includes a clutch cage 205 and a sleeve 207. A friction clutch 209 is disposed between the clutch cage 205 and the sleeve 207. The outer plate of the friction clutch 209 is engaged with the clutch cage 205 side, and the inner plate is engaged with the sleeve 207 side.

  A pressure ring 211 is disposed opposite to the friction clutch 209. The pressure ring 211 is engaged with the transfer case 215 via the pin 213 in the rotational direction, and is movable in the direction along the rotational axis. A support ring 217 is disposed opposite to the pressure ring 211. A cam mechanism having a ball 219 is provided between the support ring 217 and the pressure ring 211.

  A gear 221 is engaged with the support ring 217. The gear 221 is interlocked with the shaft 223. The shaft 223 is linked to the drive shaft 231 of the servo motor 229 via a gear 221 and a pinion 227.

  An output shaft 233 to the rear wheel side is coupled to the clutch cage 205. The output shaft 233 is linked to an input shaft 235 that receives rotational input from the engine.

  A gear 237 is interlocked and connected to the sleeve 207. A countershaft 239 that outputs to the front wheel side is rotatably supported by the transfer case 215. The countershaft 239 is provided with a gear 241. A chain 243 is wound around the gear 241 and the gear 237.

  Therefore, the torque transmitted from the engine to the input shaft 235 is directly transmitted to the rear wheel side via the output shaft 233. Further, it is transmitted to the front wheel side according to the engagement of the friction clutch 209. The engagement of the friction clutch 209 is performed by driving a servo motor 229.

  When the servo motor 229 is driven, the pinion 227 rotates in conjunction with the drive shaft 231, and the gear 221 rotates through the gear 225 and the shaft 223. By this rotation, the support ring 217 rotates within a range of 180 degrees and rotates relative to the pressure ring 211. By this relative rotation, the cam mechanism provided with the ball 219 works, and the pressure ring 211 moves to the friction clutch 209 side with respect to the support ring 217. By this movement, the friction clutch 209 is fastened.

  When the friction clutch 209 is engaged, the clutch cage 205 and the sleeve 207 are engaged according to the engagement force, and torque is also applied from the output shaft 233 to the gear 237 side via the clutch cage 205, the friction clutch 209, and the sleeve 207. Transmission takes place. Torque is transmitted from the gear 237 to the countershaft 239 via the chain 243 and the gear 241 and output to the front wheel side.

  However, in the above structure, the support ring 217 that rotates relative to the pressure ring 211 on the fixed side at a low speed is driven by the servo motor 229 via the pinion 227 and the gears 225 and 221 to rotate at a reduced speed. The reduction mechanism using the gear 221, the gear 225, and the pinion 227 is large, and it is difficult to mount the reduction mechanism in a narrow space in the transfer 201.

  Further, if the reduction gear 5 is downsized without increasing the reduction ratio from the servo motor 229 to the support ring 217, the rotation of the support ring 217 becomes steep, and it is difficult to finely adjust the friction clutch 209. Become. In addition, if the reduction ratio is reduced, the servo motor 229 itself must be enlarged in order to obtain the fastening force of the friction clutch 209, which may increase the weight.

  Furthermore, since the servo motor 229 is offset on the rotational axis parallel to the rotational axis of the friction clutch 209, the overall weight balance is poor, which may cause vehicle body vibration.

Japanese Patent No. 2715340

  The problem to be solved is that it is difficult to mount in a narrow space due to the increase in size, and that fine adjustment of the fastening becomes difficult when the size is reduced, and the size increases when the reduction ratio is reduced.

The present invention includes an input / output rotating member for transmitting torque input / output, a friction engagement portion provided between the input / output rotating members and transmitting torque between the input / output rotating members by friction engagement, and a pair of A planet gear that meshes with the gear and a planet carrier that supports the planet gear, and one of the pair of gears is supported non-rotatably on the support side, and the planet carrier is driven to rotate, or the planet The carrier is supported non-rotatably on the support side and one of the pair of gears is driven to rotate, or one of the pair of gears meshes with the planetary gear on the inner peripheral side and the other is engaged on the outer peripheral side. One of the gears is rotationally driven, and a cam mechanism provided between the pair of gears or between the planet carrier and the support side is rotated relative to the pair of gears or between the planet carrier and the support side. By Wherein comprising a pressing gear set input by the rotary drive is converted into the direction of the pressure along the axis of rotation frictionally engage the frictional engaging unit, and a rotary actuator for performing the rotation driving by the driving rotary shaft, A torque transmission coupling that performs the relative rotation due to a difference in gear ratio or meshing radius between each of the pair of gears and the planetary gear , wherein the rotary actuator is an electric motor and the friction The engaging portion and the rotary shaft core are arranged to coincide with each other, and both end sides of the drive rotary shaft are supported on the support side via a pair of bearings, and one of the input / output rotary members is an axis of the drive rotary shaft One side in the axial direction with respect to the drive rotation shaft passing through the center is supported on the support body side via a bearing, and the other side is supported on the other input / output rotation member via a bearing, The other entry output rotary member is characterized by being spline-coupled to the outer circumference of the end portion of the rotary shaft, wherein the support side to be supported through a bearing for transmitting the driving force.

The torque transmission coupling according to the present invention includes an input / output rotating member for performing torque input / output transmission and a friction engagement provided between the input / output rotating members for transmitting torque between the input / output rotating members by friction engagement. A planet gear that meshes with the gear, and a planet carrier that supports the planet gear, and one of the pair of gears is non-rotatably supported on the support side, and the planet carrier is driven to rotate. Or, the planetary carrier is supported on the support side so as not to rotate, and one of the pair of gears is rotationally driven, or one of the pair of gears meshes with the planetary gear on the inner peripheral side and the other is engaged on the outer peripheral side. Thus, one of the pair of gears is rotationally driven, and a cam mechanism provided between the pair of gears or between the planet carrier and the support side is provided between the pair of gears or between the planet carrier and the support. A pressure gear set that frictionally engages the friction engagement portion by converting the input by the rotation drive into a pressing force in a direction along the rotation axis by relative rotation between the rotation drive and the rotation drive. A torque transmission coupling that performs the relative rotation due to a difference in gear ratio or meshing radius between the pair of gears and the planetary gear, wherein the rotary actuator includes: And an electric motor which is arranged with the friction engagement portion and the rotation axis aligned with each other, and both ends of the drive rotation shaft are supported on the support side via a pair of bearings, and one of the input / output rotation members Passes through the axis of the drive rotary shaft, one side in the axial direction with respect to the drive rotary shaft is supported on the support side via a bearing, and the other side bears on the other of the input / output rotary members. Is supported through the ring, the other of said input and output rotary member is spline-coupled to the outer circumference of the end portion of the rotary shaft, wherein the support side to be supported through a bearing for transmitting the driving force.
For this reason, when rotational driving is performed by the rotary actuator, the pair of gears are relatively rotated at a low speed due to the difference in gear ratio or meshing radius between the pair of gears and the planetary gear, or the pair of gears. And the planet carrier can be relatively rotated at a low speed. By this low-speed relative rotation, the input by the rotational drive can be converted into a pressing force in a direction along the rotation axis, and the friction engagement portion can be frictionally engaged.

  That is, a pair of gears and a planetary gear meshing with the gears are provided so that the gear ratios or meshing radii between the pair of gears and the planetary gears are different, and the rotational drive of the rotary actuator is greatly decelerated to increase the pressurizing force. Therefore, the speed reduction mechanism and the rotary actuator can be downsized and formed compact.

  Therefore, it can be arranged very easily even in a narrow space such as a transfer. Further, since the rotary actuator can be reduced in size, weight can be reduced. Furthermore, since the rotational drive of the rotary actuator can be greatly decelerated and converted to the applied pressure, the fastening fine adjustment of the friction engagement portion can be easily performed.

  Since the rotary actuator and the friction engagement portion are arranged with the rotation axis aligned with each other, the overall weight balance is good and vehicle body vibration and the like can be suppressed. Moreover, since an actuator is not attached to the outside of the support body, interference with peripheral members can be prevented.

  The planetary carrier is supported on the support side so as to be relatively rotatable at a fixed angle, and the planetary carrier is interposed between the planetary carrier and the support side, and rotates in the same direction when rotated by the rotary actuator. A biasing member for regulating, and a displacement detecting means for detecting a displacement amount when the planetary carrier is thus rotationally displaced to the biasing member, and fastening of the friction engagement portion based on the detected displacement amount Since the force can be obtained, fine adjustment of the frictional engagement portion and the like can be performed accurately.

  The torque transmission device of the present invention includes an output side of a transfer of a four-wheel drive vehicle, an input side to a rear differential device, a propeller shaft between the transfer and the rear differential device, an axle shaft on the front wheel side, an axle on the rear wheel side. As a starting clutch between the shaft, the output of the prime mover and the transmission, or as a differential limiting device of the differential device, it can be accurately transmitted as any of the torque transmission couplings. .

A pair of gears and a planetary gear meshing with the gears are provided for the purpose of being able to be easily arranged in a narrow space by downsizing, and for easy fastening fine adjustment and weight reduction. This was achieved by changing the gear ratio or meshing radius between the planetary gears.
Example 1
FIG. 1 is a skeleton plan view of a four-wheel drive vehicle of a laterally mounted front engine and a rear drive base (FR base), showing the arrangement of a torque transmission coupling according to the first embodiment of the present invention.

  As shown in FIG. 1, the torque transmission coupling 1 is provided on the rear wheel output side of the transfer case 5 of the transfer 3. The transfer case 3 is attached to the vehicle body side and is a support body side. A conduction shaft 7 is rotatably supported in the transfer case 5. The transmission shaft 7 is provided with a bevel gear 9 and a spur gear 11. The bevel gear 9 meshes with a pinion gear 10 provided on the rotary shaft 61, and the spur gear 11 meshes with a spur gear 17 that is interlocked and connected to the differential case 15 side of the front differential device 13.

  Torque is input to the ring gear 23 from the engine 19 through the transmission 21 to the front differential device 13. Left and right front wheels 29 and 31 are linked to the front differential device 13 via left and right axle shafts 25 and 27.

  A propeller shaft 35 is coupled to the torque transmission coupling 1 via a constant velocity joint 33. A drive pinion shaft 39 is coupled to the propeller shaft 35 via a constant velocity joint 37. The drive pinion gear 41 of the drive pinion shaft 39 is engaged with the ring gear 45 of the rear differential device 43. The ryant differential device 43 is rotatably supported by the differential carrier 47. The left and right rear wheels 53 and 55 are linked to the rear differential device 43 via left and right axle shafts 49 and 51.

  Therefore, when torque is input from the engine 19 to the ring gear 23 of the front differential device 13 via the transmission 21, torque is transmitted to the left and right front wheels 29 and 31 via the axle shafts 25 and 27. On the other hand, torque is transmitted to the torque transmission coupling 1 via the differential case 15, the spur gears 17 and 11, the transmission shaft 7, the bevel gear 9, and the pinion gear 10.

  Torque is transmitted from the torque transmission coupling 1 to the ring gear 45 of the rear differential device 43 via the constant velocity joint 33, the propeller shaft 35, the constant velocity joint 37, the drive pinion shaft 39, and the drive pinion gear 41. Torque is transmitted from the rear differential device 43 to the left and right rear wheels 53 and 55 via the left and right axle shafts 49 and 51.

  Therefore, when the torque transmission coupling 1 is in the torque transmission state, the front wheels 29 and 31 and the rear wheels 53 and 55 can travel in the four-wheel drive state. When the torque transmission coupling 1 is not in the torque transmission state, the vehicle can travel in a two-wheel drive state with the front wheels 29 and 31.

  The details of the torque transmission coupling 1 are as shown in FIGS. FIG. 2 is a longitudinal sectional view of the torque transmission coupling 1 and its surroundings. FIG. 3 is an enlarged cross-sectional view of the main part.

  As shown in FIGS. 2 and 3, the torque transmission coupling 1 includes a clutch housing 57 and a clutch hub 59. The clutch housing 57 is configured as an input rotation member in the present embodiment, and is spline-fitted to the rotation shaft 61. The clutch housing 57 is fixed in the axial position relative to the rotary shaft 61 between a snap ring 62 attached to the rotary shaft 61 and the axial end surface of the nut 65. A unit bearing 63 is attached to the rotary shaft 61 and fastened with a nut 65. The unit bearing 63 is detachably attached to the support portion 67 of the transfer case 5 by bolt fastening or the like.

  The clutch hub 59 constitutes an output rotation member in the present embodiment, and is formed integrally with the rotation shaft 69. The rotating shaft 69 is rotatably supported by a bearing 72 on a housing 71 on the support side. The housing 71 is fastened and fixed to the transfer case 5 with bolts and nuts or the like.

  A coupling flange 73 is spline-engaged with the outer end of the rotating shaft 69. The coupling flange 73 is fastened to the rotating shaft 69 by a nut 75 and is prevented from coming off. A seal 77 is provided between the coupling flange 73 and the housing 71. The coupling flange 73 is coupled to the constant velocity joint 33.

  A friction multi-plate clutch 79 is provided between the clutch housing 57 and the clutch hub 59 as a friction engagement portion. The friction multi-plate clutch 79 has an outer plate engaged with the clutch housing 57 and an inner plate engaged with the portion 59. Therefore, torque transmission between the clutch housing 57 and the clutch hub 59 can be performed by the friction engagement of the friction multi-plate clutch 79.

A pressing member 81 is disposed opposite to the end between the clutch housing 57 and the clutch hub 59. The pressing member 81 is integrally provided with a pressure receiving portion 83 on the inner peripheral side thereof. On the inner periphery of the pressure receiving portion 83, a support boss portion 85 is provided in a circular shape.
A pressure gear set 87 is provided adjacent to the pressing member 81. The pressure gear set 87 includes a pair of gears 89 and 91, a planetary gear 93 that meshes with the gears 89 and 91, and a planet carrier 95 that supports the planetary gear 93.

  In the present invention, any one of the pair of gears 89 and 91, the planetary gear 93, and the planetary carrier 95 is supported by the housing 71 on the support side, and any of the other is rotationally driven, and the other is relatively rotated. The input by the rotational drive is converted into the applied pressure in the direction along the rotational axis, and the friction multi-plate clutch 79 is frictionally engaged.

  In the present embodiment, a gear 89 that is one of the pair of gears 89 and 91 is non-rotatably supported on the housing 71 side that is the support body side. The gear 89 is formed in a ring shape, its outer peripheral surface is spline-engaged with the inner peripheral surface of the housing 71, and the rear surface on one end side is abutted against the housing 71 in a direction along the rotation axis.

  The gear 91 is supported so as to be rotatable relative to the gear 89. The gear 91 is integrally provided with a pressurizing portion 97 in a circular shape. The pressing part 97 is supported on the outer peripheral surface of the support boss part 85 so as to be relatively rotatable. A needle bearing 99 is interposed between the pressurizing unit 97 and the pressurizing receiving unit 83.

  A cam mechanism 103 including a ball 101 is provided between the pair of gears 89 and 91. The ball 101 is disposed opposite to the cam surfaces formed on the gears 89 and 91, respectively. On the inner peripheral surfaces of the gears 89 and 91, tooth portions 90 and 92 are provided. The tooth portion 90 and the tooth portion 92 are slightly different in the number of teeth.

  The planetary gear 93 is provided with tooth portions 107 and 109 in the front and rear directions along the rotation axis with the circular recess 105 interposed therebetween. The tooth portion 107 meshes with the tooth portion 90 of the one gear 89, and the other tooth portion 109 meshes with the tooth portion 92 of the other gear 91. The recess 105 escapes the ball 101.

  The gear ratio between the gear 89 and the planetary gear 93 and the gear 91 and the planetary gear 93 is set to be slightly different depending on the difference in the number of teeth of the tooth portion 90 and the tooth portion 92.

  The planetary gear 93 is rotatably supported on the planet carrier 95. The planetary carrier 95 includes carrier plates 111 and 113. Carrier pins 115 are attached to the carrier plates 111 and 113. The planetary gear 93 is rotatably supported on the carrier pin 115.

  The carrier plates 111 and 113 are fixed to the outer peripheral side of the ring 117 by welding or the like. The ring 117 is splined to the end of the hollow rotary drive shaft 119. The rotation drive shaft 119 is an output shaft of the electric motor 121 that is a rotation actuator. The rotation drive shaft 119 is rotatably supported on the housing 71 side by bearings 123 and 125. As a result, the electric motor 121 as the rotary actuator and the frictional multi-plate clutch 79 as the friction engagement portion are arranged with their rotational axes aligned. The electric motor 121 is disposed inside the housing 71 and is stably supported by the housing 71.

  When the friction multi-plate clutch 79 is not engaged, the clutch housing 57 and the clutch hub 59 can rotate relative to each other. Therefore, even if the torque transmitted from the engine 19 side to the pinion gear 10 as described above is input to the clutch housing 57 via the rotary shaft 61, the torque is not transmitted to the clutch hub 59 side, and the torque transmission cup is not transmitted. The ring 1 is not transmitting torque. That is, as described above, traveling in a two-wheel drive state by driving the front wheels 29 and 31 can be performed.

  When the electric motor 121 is rotationally driven, torque is transmitted to the ring 117 via the rotational drive shaft 119, and the planetary carrier 95 rotates integrally. When the planetary carrier 95 rotates, the planetary gear 93 revolves around the rotation axis of the rotation drive shaft 119 via the carrier pin 115. Due to the revolution of the planetary gear 93, the planetary gear 93 meshes with the gears 89 and 91 and rotates.

  In this case, the gear ratio between the gear 89 and the planetary gear 93 is slightly different from the gear ratio between the gear 91 and the planetary gear 93, and the gear 89 is non-rotatably supported with respect to the housing 71. Yes. For this reason, the gear 91 is greatly decelerated and rotates relative to the gear 89 at a low speed. By this relative rotation, the cam surfaces of the gears 89 and 91 ride on the ball 101, and the cam mechanism 103 generates thrust.

  The thrust of the cam mechanism 103 is received on the housing side via the gear 89 and acts on the gear 91 as a reaction force. The gear 91 is moved by the action of the thrust, and the pressurizing unit 97 integral with the gear 91 pressurizes the pressurizing receiving unit 83 in the direction along the rotation axis via the needle bearing 99.

  By this pressurization, the pressing member 81 moves in the same direction, and the friction multi-plate clutch 79 is fastened with the clutch housing 57. The friction multi-plate clutch 79 exhibits a frictional engagement force according to the fastening force of the pressing member 81, and transmits torque between the clutch housing 57 and the clutch hub 59.

  Therefore, the torque transmitted from the rotating shaft 61 of the transfer 3 is transmitted from the clutch housing 57 to the clutch hub 59 via the friction multi-plate clutch 79. Torque is transmitted from the clutch hub 59 to the rotating shaft 69 and output from the rotating shaft 69 to the rear wheels 53 and 55 as described above. Thus, the vehicle can travel in a four-wheel drive state by driving the front wheels 29 and 31 and the rear wheels 53 and 55.

  Since the rotation transmitted from the rotary drive shaft 119 to the gear 91 is greatly decelerated through the planetary gear 93, the electric motor 121 can be downsized and compactly formed while the friction multi-plate clutch 79 is securely engaged. Can do.

  Since the electric motor 121 can be miniaturized and formed compact, weight reduction can be achieved. Moreover, it can arrange | position very easily also in narrow spaces, such as a transfer, by the whole size reduction.

By adjusting the driving force of the electric motor 121, the fastening force of the friction multi-plate clutch 79 can be adjusted, and the torque transmission to the rear wheels 53, 55 can be finely adjusted by the adjustment. In this case, the rotation transmitted from the rotational drive shaft 119 to the gear 91 is greatly decelerated via the planetary gear 93. For this reason, the gear 91 rotates at a very low speed with respect to the rotational drive of the electric motor 121, and the fine adjustment of the friction multi-plate clutch 79 can be easily performed. Thereby, torque adjustment can be arbitrarily and easily performed according to the traveling state of the vehicle such as starting traveling, cornering traveling, and rough road traveling.
(Example 2)
4 and 5 relate to the second embodiment of the present invention, FIG. 4 is a longitudinal sectional view of the torque transmission coupling 1A and its periphery, and FIG. 5 is an enlarged sectional view of the main part. The basic configuration is the same as that of the first embodiment, and the corresponding components will be described with the same reference numerals.

  In the torque transmission coupling 1A of the present embodiment, the gear 89A of the pressure gear set 87A is formed integrally with the ring 117A. A needle bearing 127 is provided between the gear 89 </ b> A and the housing 71. The gear 89A and the gear 91A are provided side by side in a direction along the rotation axis. A cam mechanism 103A including a ball 101 is interposed between the pair of gears 89A and 91A. The tooth portions 90A and 92A of the gear 89A and the gear 91A are formed with slightly different numbers of teeth and mesh with the tooth portion 129 of the planetary gear 93A.

  The planetary carrier 95A of this embodiment includes a carrier pin 115A and a housing 71, and the carrier pin 115A is screwed and fixed to the housing 71. Thereby, the planetary carrier 95A is configured to be supported on the support side so as not to rotate. The planetary gear 93A is rotatably supported between the carrier pin 115A and the housing 71. A plurality of planetary gears 93A supported by carrier pins 115A are provided at predetermined intervals in the circumferential direction of gears 89A and 91A.

  When the electric motor 121 is rotationally driven, one gear 89A is integrally rotated through the rotational drive shaft 119. When the gear 89A is rotationally driven, the planetary gear 93A meshing with the gear 89A rotates, and the gear 91A meshing with the planetary gear 93A is interlocked. That is, both the gear 89A and the gear 91A rotate.

  The gear ratio between the planetary gear 93A and the gear 89A and the gear ratio between the planetary gear 93A and the gear 91A are set slightly different as described above. For this reason, the gear 91A rotates relative to the gear 89A at a low speed while rotating together with the gear 89A. Due to this relative rotation, the cam mechanism 103A works in the same manner as described above to generate thrust.

  The gear 89A is supported on the housing 71 side via a needle bearing 127. Therefore, the thrust is received on the housing 71 side, and the gear 91A moves to the pressure receiving portion 83 side by the reaction force. By this movement, the friction multi-plate clutch 79 can be fastened through the pressing member 81 as described above.

  Therefore, also in this embodiment, there can exist an effect substantially the same as 1st Embodiment.

In addition, the planet carrier 95A can be constituted by the carrier pins 115A and the housing 71, so that a simple structure can be formed and the overall can be made more compact. Also, weight reduction can be achieved.
(Example 3)
6 and 7 relate to Embodiment 3 of the present invention, FIG. 6 is a longitudinal sectional view of the torque transmission coupling 1B and its periphery, and FIG. 7 is an enlarged sectional view of the main part. The basic configuration of the present embodiment is the same as that of the second embodiment, and the corresponding components will be described with the same reference numerals.

  In the torque transmission coupling 1B according to the present embodiment, the number of teeth of the gear 89A of the pressure gear set 87B and the tooth portions 90B and 92B of the gear 91A are set to be the same. The tooth portions 107B and 109B of the planetary gear 93B are formed of, for example, face gears, and the outer diameter thereof is set so that the tooth portion 109B is larger than the tooth portion 107B.

  The carrier pin 115A of the planetary carrier 95A is screwed and fixed obliquely with respect to the housing 71. In this state, the teeth 107B and 109B of the planetary gear 93B are engaged with the gears 89A and the teeth 90B and 92B of the gear 91A, respectively. .

  Therefore, in this embodiment, the meshing radii between the pair of gears 89A and 91A and the planetary gear 93B are set to be different.

The operation of the present embodiment is almost the same as that of the second embodiment, and the planetary gear 93B rotates due to the rotational drive of the gear 89A, and the gear 91A rotates with the gear 89A due to the difference in meshing radius while rotating with respect to the gear 89A. Relative rotation at low speed. As a result, the frictional multi-plate clutch 79 is engaged as described above. Therefore, in this embodiment as well, it is possible to achieve substantially the same operational effects as in the second embodiment.
Example 4
8 and 9 relate to a fourth embodiment of the present invention, FIG. 8 is a longitudinal sectional view of the torque transmission coupling 1C and its periphery, and FIG. 9 is an enlarged sectional view of the main part. The basic configuration is the same as that of the first embodiment, and the corresponding components will be described with the same reference numerals.

  In the torque transmission coupling 1C of the present embodiment, one gear 89C of the pressure gear set 87C is integrally provided at the end of the rotary drive shaft 119C. Therefore, the electric motor 121 is configured to rotationally drive one 89C of the pair of gears 89C and 91C.

  The other 91C of the pair of gears is provided integrally with the housing 71 on the support side, and is supported on the support side so as not to rotate.

  The planetary carrier 95C includes carrier plates 111C and 113C. A planetary gear 93C is rotatably supported by carrier pins 115C fixed to the carrier plates 111C and 113C.

  A cam mechanism 103C including a ball 101 is provided between the housing 71 and the carrier plate 113C. The ball 101 faces the cam surface formed on the inner wall surface of the housing 71 and the cam surface formed on the side surface of the carrier plate 113C.

  As described above, the gear 91C is formed integrally with the housing 71, and the cam mechanism 103C is interposed between the housing 71 and the carrier plate 113C. Thus, the cam mechanism 103C is interposed between the other side 91C of the pair of gears 89C and 91C and the planetary carrier 95C.

  The tooth portions 90C and 92C of the pair of gears 89C and 91C have different pitch circle radii, and the tooth portion 92C is set larger. Since the tooth portions 129C of the planetary gear 93C are engaged with the tooth portions 90C and 92C, the meshing radii between the pair of gears 89C and 91C and the planetary gear 93C are different.

  When the electric motor 121 is rotationally driven, the gear 89C is integrally rotated. When the gear 89C is driven to rotate, the planetary gear 93C rotates while meshing with the pair of gears 89C and 91C. At this time, the meshing radius between the pair of gears 89C and 91C and the planetary gear 93C is different, so that the planetary gear 93C revolves at a low speed around the rotation axis of the rotation drive shaft 119C. Due to this revolution, the planetary carrier 95C is largely decelerated with respect to the housing 71 on the gear 91C side via the carrier pin 115C and relatively rotated at a low speed. This relative rotation causes the cam mechanism 103C to work and generate thrust. This thrust is received on the housing 71 side, and the planetary carrier 95C moves relative to the housing 71 toward the pressing member 81 by the reaction force. By the movement of the pressing member 81, the friction multi-plate clutch 79 can be fastened.

  Therefore, the present embodiment can provide substantially the same effects as the first embodiment.

  In addition, since the gear 89C is integrally provided on the electric motor 121 side and the gear 91C is integrally provided on the housing 71 side, the number of parts can be reduced and the device can be formed more compactly.

  FIG. 10 shows an embodiment according to a modification of the fourth embodiment, and is a longitudinal sectional view of the torque transmission coupling 1D and its periphery.

  In the torque transmission coupling 1D of the present embodiment, a sleeve 131 protruding into the housing 71 is provided in the transfer case 5 with respect to the basic structure of the fourth embodiment, and the sleeve 131 and the boss portion of the clutch housing 57 are provided. 133 is provided with a seal 135 interposed therebetween.

  Therefore, in this embodiment, the transfer case 5 and the housing 71 can be closed by the seal 135, and an appropriate type of lubricating oil or the like can be used for both the transfer 3 and the torque transmission coupling 1D. it can.

  The setting of the input / output relationship is arbitrary, and the clutch housing 57 side can be configured as an output rotating member and the clutch hub 59 side can be configured as an input rotating member. The friction engagement portion only needs to generate a friction engagement force by fastening, and is not limited to the friction multi-plate clutch 79 but can be arbitrarily selected such as a cone clutch.

  The arrangement of the torque transmission couplings 1, 1 </ b> A, 1 </ b> B, 1 </ b> C, and 1 </ b> D is not limited to the one attached to the output side of the transfer 3, but the torque transmission couplings 1 </ b> E, 1 </ b> F, 1 </ b> G, 1 </ b> H, 1 </ b> I, 1 </ b> J It is also possible to select and arrange as appropriate.

  The torque transmission coupling 1E is interposed in the propeller shaft 35. In this case, the pinion gear 10 and the like are omitted, and the rotation shafts 61 and 69 of the torque transmission couplings 1, 1 </ b> A, 1 </ b> B, 1 </ b> C, and 1 </ b> D are coupled to the propeller shaft 35.

  By adjusting the fastening of the torque transmission couplings 1, 1 </ b> A, 1 </ b> B, 1 </ b> C, and 1 </ b> D interposed in the propeller shaft 35, the torque transmission can be adjusted toward the rear wheels 53 and 55.

  When the torque transmission coupling 1E is in a torque non-transmission state, the rotation from the rear wheels 53 and 55 is not transmitted to the constant velocity joint 33, the rotating shaft 61, etc. upstream of the torque transmission coupling 1E. Accordingly, energy loss can be suppressed.

  The torque transmission couplings 1F and 1G are interposed in axle shafts 49 and 51, respectively. In this case, the pinion gear 10 and the like are omitted, and the rotation shafts 61 and 69 of the torque transmission couplings 1, 1 </ b> A, 1 </ b> B, 1 </ b> C, and 1 </ b> D are coupled to the axle shafts 49 and 51. The torque transmission couplings 1F and 1G can be provided only on one of the axle shafts 49 and 51.

  When the torque transmission couplings 1F and 1G are in a torque non-transmission state, the rotation from the rear wheels 53 and 55 is not transmitted to the rear differential device 43 side, and energy loss during two-wheel drive is further suppressed. can do.

  The torque transmission couplings 1H and 1I are interposed in the axle shafts 25 and 27 on the front wheels 29 and 31 side. In this case, the pinion gear 10 and the like are omitted, and the rotation shafts 61 and 69 of the torque transmission couplings 1, 1A, 1B, 1C, and 1D are coupled to the axle shafts 25 and 27. The torque transmission couplings 1F and 1G can be provided only on one of the axle shafts 25 and 27.

  The functions of the torque transmission couplings 1H and 1I are substantially the same as those of the torque transmission couplings 1F and 1G.

  The torque transmission coupling 1J is provided on the drive pinion shaft 39 and is disposed in the differential carrier 47 of the rear differential device 43. In this case, the rotation shaft 61 and the pinion gear 10 of each torque transmission coupling 1, 1A, 1B, 1C, 1D become the drive pinion shaft 39 and the drive pinion gear 41, and the coupling flange 73 side is coupled to the constant velocity joint 37 side.

  The torque transmission coupling 1T is provided as a starting clutch between the output of the engine 19 as a prime mover and the transmission 21.

The torque transmission coupling 1U is provided as a differential limiting device for the rear differential device 43, which is a differential device.
(Example 5)
FIG. 11 is a skeleton plan view of a four-wheel drive vehicle of a longitudinally mounted front engine and a rear drive base (FR base), showing the arrangement of a torque transmission coupling according to a fifth embodiment of the present invention. Components corresponding to those in FIG. 1 will be described with the same reference numerals.

  In the present embodiment, a torque transmission coupling 1K is provided in the transfer 3A. In this torque transmission coupling 1K, the pinion gear 10 of the rotating shaft 61 in the structure of FIGS. 2 to 10 is omitted, and the rotating shaft 61 is coupled so as to input torque from the transmission 21 of FIG. The rotation shaft 69 of the torque transmission coupling 1K is coupled to the propeller shaft 35 via the constant velocity joint 33 on the coupling flange 73 side.

  A sprocket 141 is integrally provided on the rotating shaft 61. A chain 147 is looped between the sprocket 141 and a sprocket 145 provided on the transmission shaft 143. The transmission shaft 143 is connected to the transmission shaft 151 side via the propeller shaft 149. The pinion gear 153 of the transmission shaft 151 meshes with the ring gear 23 of the front differential device 13.

  Therefore, by the engagement control of the friction multi-plate clutch 79, torque is transmitted to the propeller shaft 35 via the friction multi-plate clutch 79 on the one hand. On the other hand, torque can be input from the transmission 21 to the front differential device 13 via the sprocket 141, the chain 147, the sprocket 145, the transmission shaft 143, the propeller shaft 149, the transmission shaft 151, the pinion gear 153, and the ring gear 23.

  By controlling the engagement of the friction multi-plate clutch 79 of the torque transmission coupling 1K according to the running state, the torque distribution to the rear wheels 53 and 55 is controlled according to the running state, and directly connected to the front wheels 29 and 31. Torque is transmitted in the state, and two-wheel drive and accurate four-wheel drive can be performed.

  The transmission shaft 143 can be provided as a torque transmission coupling 1L. In this case, the sprocket 145 is provided in the clutch housing 57 of FIGS. 2 to 10, the pinion gear 10 of the rotating shaft 61 is omitted, and the rotating shaft 61 is coupled to the propeller shaft 149 on the front wheel side as the transmission shaft 143. The rotating shaft 69 is rotatably supported on the transfer case 5 side.

Therefore, by controlling the engagement of the friction multi-plate clutch 79 of the torque transmission coupling 1L according to the traveling state, the torque distribution to the front wheels 29 and 31 is controlled according to the traveling state, and the rear wheels 53 and 55 are connected. Torque is transmitted in a directly connected state, and two-wheel drive and accurate four-wheel drive can be performed.
(Example 6)
12 to 14 show Embodiment 6 of the present invention. FIG. 12 shows the arrangement of the torque transmission coupling, a skeleton plan view of a four-wheel drive vehicle having a horizontally mounted front engine and a rear drive base (FR base), and FIG. 13 is a longitudinal section of the torque transmission coupling 1M and its surroundings. 14 and 14 are enlarged cross-sectional views of the main part. The basic configuration of this embodiment is the same as that of the second embodiment shown in FIGS. 4 and 5, and components corresponding to those shown in FIGS.

  The torque transmission coupling 1M of the present embodiment is attached to the rear differential device 43 side. A housing 71M on the support side that accommodates the torque transmission coupling 1M is fastened and coupled to a differential carrier 47M on the support side by bolts 154 or the like. The drive pinion gear 41 of the drive pinion shaft 39 that is the rotation shaft of the torque transmission coupling 1M is engaged with the ring gear 45 of the rear differential device 43. The connecting shaft 73 of the rotating shaft 69 of the torque transmission coupling 1M is connected to the constant velocity joint 37 side. Since the electric motor 121 is housed and supported inside the housing 71M and disposed on the foremost side in the vehicle traveling direction of the torque transmission coupling 1M, the cooling efficiency is improved.

  The friction multi-plate clutch 79 of this embodiment is disposed on the outer peripheral side of bearings 155 and 157 that rotatably support the drive pinion shaft 39 that is the rotating shaft.

  Specifically, the vertical wall 159 of the clutch hub 59M as an input rotating member is arranged at the end of the clutch hub 59M while moving toward the rotating shaft 69 side. An inner cylindrical portion 161 is integrally provided on the inner peripheral side of the clutch housing 57M as an output rotating member, and the inner peripheral portion 165 of the vertical wall 163 provided at the end portion of the inner cylindrical portion 161 is splined to the end portion of the drive pinion shaft 39. Combined. A support portion 67M provided on the differential carrier 47M is also projected on the inner peripheral side of the inner cylinder portion 161, and the bearings 155 and 157 are supported on the support portion 67M. The bearings 155 and 157 support the drive pinion shaft 39, which is the rotating shaft, so as to be rotatable with respect to the support portion 67M. Further, a bearing 166 is disposed between the rotary shaft 69 and the inner peripheral portion 165 of the clutch housing 57M, and is in a supporting relationship with each other.

Therefore, in this embodiment, in addition to the operational effects of the second embodiment, the bearing span of the drive pinion shaft 39 can be increased and the drive pinion shaft 39 can be reliably supported by the support portion 67M. In addition, since the support portion 67M is housed on the inner peripheral side of the inner cylinder portion 161, the support portion 67M can be formed compactly as a whole by effectively using the internal space.
(Example 7)
15 and 16 show Embodiment 7 of the present invention. FIG. 15 is a longitudinal sectional view of the torque transmission coupling 1N and its periphery, and FIG. 15 is an enlarged sectional view of the main part. The basic configuration of this embodiment is the same as that of the sixth embodiment shown in FIGS. 13 and 14, and components corresponding to those in FIGS. 13 and 14 are denoted by the same reference numerals.

  In this embodiment, the oil passage 167 is provided in the support portion 67N on the differential carrier 47N side that supports the bearings 155 and 157. The oil passage 167 extends from one side of the support portion 67N to the other side and guides the lubricating oil to the bearing 155. The oil passage 167 is provided in the built-up portion 169 provided on the outer periphery of the upper portion of the support portion 67N, and is formed so as to incline downward from the differential carrier 47N to the end surface 171 of the built-up portion 169. The end surface 171 is located at one end of the outer periphery of the bearing 155, and the outer periphery of the bearing 155 is opened at this portion. The upper surface of the built-up portion 169 is inclined corresponding to the inclination of the oil passage 167, and the inner cylinder portion 161N of the clutch housing 57N is also formed in a tapered shape corresponding to this inclination. In the differential carrier 47N, a guide wall 173 is provided at the end of the oil passage 167 and is continuous with one side wall of the oil passage 167.

  During the meshing rotation of the pinion gear 10 and the ring gear 45, the scattered gear oil in the differential carrier 47N is guided by the guide wall 173 to the oil path 167, or the scattered gear oil directly reaches the oil path 167. The gear oil in the oil passage 167 flows to the outer peripheral surface of the bearing 155 due to the inclination of the oil passage 167, and the bearing 155 is sufficiently lubricated by the gear oil.

Therefore, in this embodiment, in addition to the function and effect of the sixth embodiment, the bearing 155 can be sufficiently lubricated with gear oil even if the support portion 67N is lengthened to increase the bearing span.
(Example 8)
17 and 18 show an eighth embodiment of the present invention. FIG. 17 is a longitudinal sectional view of the torque transmission coupling 1P and its periphery, and FIG. 18 is an enlarged sectional view of the main part. The basic configuration of the present embodiment is the same as that of the seventh embodiment shown in FIGS. 15 and 16, and components corresponding to those in FIGS.

  In this embodiment, the support portion 67P and the built-up portion 169P are slightly extended from the bearing 155 so as to protrude in the direction along the rotation axis, and between the support portion 67P tip inner periphery and the seal sliding ring 175. A seal 177 was provided. The seal sliding ring 175 is fastened and fixed between the nut 65 and the inner race of the bearing 155. With this configuration, the oil passage 167P extends from the outer peripheral surface of the bearing 155 to the space between the inner race and the outer race of the bearing 155.

  Note that the electric motor 121P of the present embodiment is formed long, and the gear 89A of the pressure gear set 87A is integrally provided at the end of the rotary drive shaft 119P.

  In this embodiment, the scattered gear oil that reaches the oil passage 163P flows to the outer periphery of the bearing 155 with the inclination of the oil passage 163P. From the outer periphery of the bearing 155, gear oil flows between the inner race and the outer race, and the bearing 155 is reliably lubricated. Excess oil when lubricating the bearing 155 flows on the inner peripheral side of the support portion 67P, and can return to the differential carrier 47P while lubricating the other bearing 157. Since the friction multi-plate clutch 79 side is partitioned by a seal 177 with respect to the bearing 155 side, for example, automatic transmission oil or the like different from the gear oil can be used. The automatic transmission oil can accurately lubricate the friction multi-plate clutch 79 and the like separately from the bearing 155 and the like.

  Therefore, in this embodiment, in addition to the function and effect of the seventh embodiment, the bearing 155 can be more reliably lubricated, and the bearing 155 side and the friction multi-plate clutch 79 side can be reliably and accurately lubricated with appropriate oils, respectively. can do.

Moreover, since the electric motor 121P is formed long, the outer diameter can be reduced at this portion.
Example 9
19 to 21 show Embodiment 9 of the present invention. 19 is a longitudinal sectional view of the torque transmission coupling 1Q and its surroundings, FIG. 20 is a sectional view showing the displacement detecting means and its surroundings, and FIG. 21 is an enlarged sectional view of the main part. The basic configuration of the present embodiment is the same as that of the seventh embodiment shown in FIGS. 15 and 16, and components corresponding to those in FIGS.

  In this embodiment, a displacement sensor 179 constituting a displacement detecting means is provided in order to obtain the fastening force of the friction multi-plate clutch 79.

  Specifically, the planet carrier 181 that supports the planetary gear 93A is composed of a pair of carrier plates 183 and 185 and carrier pins 187, and is supported by a housing 71Q on the support side so as to be relatively rotatable by a fixed angle. The carrier plate 185 is positioned in the direction along the rotation axis by a stopper 188 provided on the inner surface of the housing 71Q.

  As shown in FIG. 20, a coil spring 189 is interposed as a biasing member between the planet carrier 181 and the housing 71Q. That is, the carrier plates 183 and 185 of the planetary carrier 181 are provided with notches 191. A spring accommodating portion 193 is provided on the housing 71Q side so as to face the notch portion 191. The coil spring 189 is interposed between the notch portion 191 and the spring accommodating portion 193. Therefore, the coil spring 189 is configured to restrict the rotation of the planet carrier 181 that rotates in the same direction when driven by the electric motor 121 by the urging force.

  At least one of the carrier plates 183 and 185 is provided with a protrusion 195 on the outer periphery. The convex portion 195 faces the concave portion 197 formed on the housing 71Q side. The convex part 195 is relatively movable in the rotational direction of the carrier plates 183 and 185 within the concave part 197, and the planetary carrier 181 is configured to be relatively rotatable relative to the housing 71Q by a certain angle.

  The displacement sensor 179 is installed at a predetermined location outside the housing 71Q. The displacement sensor 179 is linked and connected to the convex portion 195 by a link 199. Therefore, when the convex portion 195 moves, the displacement amount is input to the displacement sensor 179 via the link 199, and the rotational displacement amount of the carrier plates 183 and 185 can be detected.

  When the electric motor 121 is rotationally driven, one gear 89A is integrally rotated through the rotational drive shaft 119. When the gear 89A is rotationally driven, the planetary gear 93A meshing with the gear 89A rotates, and the gear 91A meshing with the planetary gear 93A is interlocked. That is, both the gear 89A and the gear 91A rotate.

  The gear ratio between the planetary gear 93A and the gear 89A and the gear ratio between the planetary gear 93A and the gear 91A are set slightly different as described above. For this reason, the gear 91A rotates relative to the gear 89A at a low speed while rotating together with the gear 89A. As a result of this relative rotation, the cam mechanism 103A works in the same manner as described above to generate thrust, and the frictional multi-plate clutch 79 is engaged.

  When the friction multi-plate clutch 79 is engaged, the gear 91A receives a rotation restricting force proportional to the engaging force. By this rotation restricting force, a rotational force proportional to the fastening force is transmitted to the carrier plates 183 and 185 via the planetary gear 93A and the carrier pin 187. This rotational force causes the carrier plates 183 and 185 to rotate relative to the housing 71Q against the biasing force of the coil spring 189, and the convex portion 195 moves relative to the concave portion 197. This relative movement is input to the displacement sensor 199 via the link 201, and a displacement proportional to the fastening force can be detected.

  Therefore, the detected displacement is input to the controller, the fastening force of the friction multi-plate clutch 79 is obtained by a predetermined calculation, and fine adjustment of the engagement of the friction multi-plate clutch 79 can be performed accurately.

Therefore, the present embodiment can achieve substantially the same operational effects as those of the seventh embodiment, and the fine adjustment of the friction multi-plate clutch 79 can be accurately performed.
(Example 10)
22 and 23 show a tenth embodiment of the present invention. FIG. 22 is a longitudinal sectional view of the torque transmission coupling 1R and its periphery, and FIG. 23 is an enlarged sectional view of the main part. The basic configuration of the present embodiment is the same as that of the seventh embodiment shown in FIGS. 15 and 16, and components corresponding to those in FIGS.

  In the present embodiment, the electric motor 121R is disposed outside the housing 71R.

  That is, a step portion 203 is provided in the housing 71R, and the electric motor 121R is accommodated in an outer peripheral space formed by the step portion 203. The rotation drive shaft 119R is rotatably supported by the vertical wall portion 205 of the housing 71R, and a drive gear 207 is attached to the rotation drive shaft 119R in the housing 71R. The drive gear 207 meshes with the gear 89A of the compressible gear set 87A.

  A ring portion 117R is integrally provided on the gear 89A, and the ring portion 117R is rotatably supported by the housing 71R by a bearing 125R.

  Then, by driving the electric motor 121R, the gear 89A of the pressure gear set 87A can be rotated via the rotary drive shaft 119R and the drive gear 207, and the frictional multi-plate clutch 79 can be adjusted by fastening as in the seventh embodiment. it can.

Therefore, in this embodiment, the same effects as those of the seventh embodiment can be obtained, and the housing 71R can be formed more compactly.
(Example 11)
24 and 25 show Embodiment 11 of the present invention. FIG. 24 is a longitudinal sectional view of the torque transmission coupling 1S and its periphery, and FIG. 25 is an enlarged sectional view of the main part. The basic configuration of this embodiment is the same as that of the sixth embodiment shown in FIGS. 13 and 14, and components corresponding to those in FIGS. 13 and 14 are denoted by the same reference numerals.

  In the present embodiment, the positions of the friction multi-plate clutch 79S and the electric motor 121S are switched with respect to the sixth embodiment, and the electric motor 121S is accommodated in the housing 71A connected to the differential carrier 47S, and the outer peripheral side of the bearings 155 and 157 Arranged.

  That is, the support portion 67S provided on the differential carrier 47S is projected toward the housing 71A inner side toward the housing 71S, and the bearings 155 and 157 are supported by the support portion 67S. With these bearings 155 and 157, the drive pinion shaft 39 is rotatably supported with respect to the support portion 67S.

  A rotation drive shaft 119S is rotatably supported on the outer periphery of the support portion 67S via bearings 209 and 211. The inner race of the bearing 211 is abutted and supported on the differential carry 47S side in a direction along the rotational axis.

  A gear 89A of a pressure gear set 87A is splined to the rotary drive shaft 119S. The differential carrier 47S, the housing 71A, and the housing 71S are integrally coupled by a fastening bolt (not shown).

  Of the clutch housing 57S and the clutch hub 59S of the torque transmission coupling 1S, the clutch housing 57S is integrally provided on the rotating shaft 69 as an input rotating member, and the clutch hub 59S is splined to the drive pinion shaft 39 as an output rotating member. Has been.

  When the electric motor 121S is rotationally driven, one gear 89A is rotationally driven integrally through the rotational drive shaft 119S. When the gear 89A is rotationally driven, the planetary gear 93A meshing with the gear 89A rotates, and the gear 91A meshing with the planetary gear 93A is interlocked. That is, both the gear 89A and the gear 91A rotate.

  The gear ratio between the planetary gear 93A and the gear 89A and the gear ratio between the planetary gear 93A and the gear 91A are set slightly different as described above. For this reason, the gear 91A rotates relative to the gear 89A at a low speed while rotating together with the gear 89A. Due to this relative rotation, the cam mechanism 103A works in the same manner as described above to generate thrust.

  The gear 89A is supported in the direction along the rotational axis on the differential carrier 47S side via the rotational drive shaft 119S and the bearing 211. Therefore, the thrust is received on the differential carrier 47S side, and the gear 91A moves to the pressure receiving portion 83 side by the reaction force. By this movement, the friction multi-plate clutch 79S can be fastened via the pressing member 81 as described above. A needle bearing 40 is disposed between the rotary shaft 69 and the drive pinion shaft 39, and directly supports each other.

Therefore, in this embodiment as well, it is possible to achieve substantially the same operational effects as in the sixth embodiment. In the present embodiment, the bearing span of the drive pinion shaft 39 can be increased, and the drive pinion shaft 39 can be reliably supported by the support portion 67S. Moreover, since the support part 67S becomes a form accommodated in the electric motor 121S inner peripheral side, it can be formed compactly as a whole by effective use of the internal space.
The arrangement of the torque transmission couplings 1M, 1N, 1P, 1Q, 1R, 1S is not limited to the arrangement on the rear differential device 43 side, and the torque transmission couplings 1, 1A, 1B, 1C, 1D, 1E in FIG. , 1F, 1G, 1H, 1I, 1T, 1U can be appropriately selected and arranged. In this case, the coupling of the shafts is performed by appropriately changing the same as in the torque transmission couplings 1, 1A, 1B, 1C, 1D.

(Example 1) which is a skeleton top view of the four-wheel drive vehicle which shows arrangement | positioning of a torque transmission coupling. (Example 1) which is a longitudinal cross-sectional view of a torque transmission coupling and its periphery. (Example 1) which is a principal part expanded sectional view of a torque transmission coupling. (Example 2) which is a longitudinal cross-sectional view of a torque transmission coupling and its periphery. (Example 2) which is a principal part expanded sectional view of a torque transmission coupling. (Example 3) which is a longitudinal cross-sectional view of a torque transmission coupling and its periphery. (Example 3) which is a principal part expanded sectional view of a torque transmission coupling. (Example 4) which is a longitudinal cross-sectional view of a torque transmission coupling and its periphery. (Example 4) which is a principal part expanded sectional view of a torque transmission coupling. (Example 4) which is a longitudinal cross-sectional view of a torque transmission coupling and its periphery. (Example 5) which is a skeleton top view of the four-wheel drive vehicle which shows arrangement | positioning of a torque transmission coupling. (Example 6) which is a skeleton top view of the four-wheel drive vehicle which shows arrangement | positioning of a torque transmission coupling. (Example 6) which is a longitudinal cross-sectional view of a torque transmission coupling and its periphery. (Example 6) which is a principal part expanded sectional view of a torque transmission coupling. (Example 7) which is a longitudinal cross-sectional view of a torque transmission coupling and its periphery. (Example 7) which is a principal part expanded sectional view of a torque transmission coupling. (Example 8) which is a longitudinal cross-sectional view of a torque transmission coupling and its periphery. (Example 8) which is a principal part expanded sectional view of a torque transmission coupling. (Example 9) which is a longitudinal cross-sectional view of a torque transmission coupling and its periphery. (Example 9) which is sectional drawing which shows a displacement detection means. (Example 9) which is a principal part expanded sectional view of a torque transmission coupling. (Example 10) which is a longitudinal cross-sectional view of a torque transmission coupling and its periphery. (Example 10) which is a principal part expanded sectional view of a torque transmission coupling. (Example 11) which is a longitudinal cross-sectional view of a torque transmission coupling and its periphery. (Example 11) which is a principal part expanded sectional view of a torque transmission coupling. It is sectional drawing of a transfer (conventional example).

Explanation of symbols

1, 1A, 1B, 1C, 1D, 1E, 1F, 1G, 1H, 1I, 1J, 1K, 1L, 1M, 1N, 1P, 1Q, 1R, 1S, 1T, 1U Torque transmission coupling 5 Transfer case (support) body)
47M, 47N, 47P, 47S Differential carrier (support)
57,57S Clutch housing (input rotating member)
59, 59S Clutch hub (output rotating member)
57M, 57N Clutch housing (output rotating member)
59M Clutch hub (input rotating member)
67M, 67N, 67P, 67S Support 71, 71M, 71N, 71Q, 71R, 71S Housing (support)
79,79S Friction multi-plate clutch (friction engagement part)
87, 87A, 87B, 87C Pressurized gear set 89, 89A, 89C, 91A, 91C Gear 93, 93A, 93B, 93C Planetary gear 95, 95A, 95C, 181 Planetary carrier 103, 103A, 103C Cam mechanism 121, 121R, 121S Electric motor (rotary actuator)
155, 157 Bearing 167 Oil passage 179 Displacement sensor (displacement detection means)
181 Planetary carrier 189 Coil spring (biasing member)

Claims (3)

An input / output rotating member for transmitting and outputting torque; and
A friction engagement portion that is provided between the input / output rotation members and transmits torque between the input / output rotation members by friction engagement;
A planet gear that meshes with the gear and a planet carrier that supports the planet gear, and one of the pair of gears is non-rotatably supported on the support side and the planet carrier is driven to rotate, or The planetary carrier is supported on the support side so as not to rotate, and one of the pair of gears is rotationally driven, or one of the pair of gears meshes with the planetary gear on the inner peripheral side and the other is engaged on the outer peripheral side. One of the pair of gears is rotationally driven, and a cam mechanism provided between the pair of gears or between the planet carrier and the support side is provided between the pair of gears or between the planet carrier and the support side. A pressure gear set that frictionally engages the friction engagement portion by converting the input by the rotation drive into a pressing force in a direction along the rotation axis by relative rotation ;
A rotation actuator that performs the rotation drive by a drive rotation shaft ,
A torque transmission coupling for causing the relative rotation due to a difference in gear ratio or meshing radius between each of the pair of gears and the planetary gear ,
The rotary actuator is an electric motor and is arranged with the friction engagement portion and the rotation axis aligned with each other, and both ends of the drive rotation shaft are supported on the support side via a pair of bearings,
One of the input / output rotation members passes through the axis of the drive rotation shaft, and one side in the axial direction with respect to the drive rotation shaft is supported on the support side via a bearing, and the other side is the other of the input / output rotation members. Supported through bearings,
2. The torque transmission coupling according to claim 1, wherein the other of the input / output rotating members is splined to an outer periphery of an end of a rotating shaft that is supported on the support side via a bearing and transmits a driving force . The torque transmission coupling according to claim 1,
The planet carrier is supported on the support side so as to be relatively rotatable at a fixed angle,
An urging member is provided between the planet carrier and the support side to restrict the rotation of the planet carrier that rotates in the same direction when rotated by the electric motor by an urging force;
Displacement detecting means for detecting a displacement amount when the planetary carrier is rotationally displaced against the biasing member;
A torque transmission coupling that obtains a fastening force of the friction engagement portion based on the detected displacement amount . The torque transmission coupling according to claim 1 or 2 ,
The output side of the transfer of a four-wheel drive vehicle, the input side to the rear differential unit, the propeller shaft between the transfer and the rear differential unit, the axle shaft on the front wheel side, the axle shaft on the rear wheel side, the output and transmission of the prime mover A torque transmission coupling, wherein the torque transmission coupling is arranged as any one of a differential limiting device of a differential device as a starting clutch between the two .

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