JP7559711B2 - Geared motor and clutch actuator using the same - Google Patents

Description

The present invention relates to a geared motor and a clutch actuator using the same.

Conventionally, a clutch actuator is known that is provided between a first transmission part and a second transmission part that are capable of rotating relative to one another, and that is capable of changing the state of a clutch that changes between an engaged state that allows the transmission of torque between the first transmission part and the second transmission part, and a disengaged state that blocks the transmission of torque between the first transmission part and the second transmission part.

For example, the clutch actuator described in Patent Document 1 includes a geared motor as a drive unit for pressing the clutch. The geared motor includes an electric motor and a reducer that can reduce the torque from the electric motor and output it. The reducer has a sun gear, planetary gears, and two ring gears.

JP 2006-90533 A

In the geared motor described in Patent Document 1, the axial position of the carrier subassembly, which is made up of a planetary gear, pin, and carrier, is regulated between parts that are synchronized with the carrier and sun gear, and a relatively large relative rotation occurs in the position regulation part. As a result, the sliding speed between the carrier subassembly and the position regulation part is relatively high. When torque acts on the geared motor's reducer during the process of pressing the clutch, the planetary gear tilts, and an axial load acts on the entire carrier. As a result, the axial position regulation part may be subject to severe wear due to sliding stress. Here, sliding stress corresponds to the product of the sliding distance and the axial load.

To reduce the amount of wear, measures such as (1) reducing the sliding stress, i.e., reducing the sliding distance or axial load, and (2) improving the specific wear rate can be considered. However, with the configuration of the geared motor in Patent Document 1, (1) it is not possible to reduce the sliding stress, and (2) in order to improve the specific wear rate, it is necessary to at least heat treat the carrier, which may complicate the manufacturing process. In addition, while the above-mentioned problems may be solved by supporting the entire carrier subassembly with bearings or the like, this may result in an increase in size, etc.

In addition, in the geared motor of Patent Document 1, frictional force occurs in the position restriction section, so the "actual efficiency of the entire reducer" is lower than the "theoretical efficiency considering only meshing efficiency," and there is a risk of the performance of the clutch actuator being degraded. As a result, for example, when engaging the clutch, it is necessary to output extra torque from the electric motor to compensate for the decrease in efficiency, which may increase the current and power consumption and the amount of heat generated by the electric motor. This may result in a decrease in the reliability of the clutch actuator.

The object of the present invention is to provide a geared motor and clutch actuator that can improve wear resistance and reducer efficiency with a simple configuration.

The geared motor according to the present invention comprises a housing (12), an electric motor (20), a reducer (30), and a rotating part (40). The electric motor is provided in the housing and is capable of outputting torque when electricity is applied. The reducer is capable of reducing and outputting the torque from the electric motor. The rotating part rotates due to the torque from the reducer.

The reducer has a sun gear (31), multiple planetary gears (32), a pin (335), a planetary gear bearing (36), an annular carrier (33), an annular first ring gear (34), and an annular second ring gear (35). Torque is input to the sun gear from an electric motor. The planetary gears can revolve in the circumferential direction of the sun gear while rotating while meshing with the sun gear. The pin is provided at the center of rotation of the planetary gears.

The planetary gear bearing is provided between the planetary gear and the pin. The carrier rotatably supports the planetary gear by supporting one end of the pin, and is rotatable relative to the sun gear. The first ring gear is capable of meshing with the planetary gear. The second ring gear is capable of meshing with the planetary gear and is formed so that the number of teeth in the tooth portion differs from that of the first ring gear, and outputs torque to the rotating portion.

The planetary gear, pin, planetary gear bearing and carrier constitute a carrier subassembly (330). At least one of the multiple planetary gears has a first planetary gear annular surface (901) which is an annular surface at one axial end and a second planetary gear annular surface (902) which is an annular surface at the other axial end. The sun gear has a sun gear annular surface (911) which is an annular surface a part of which abuts against and faces a part of the first planetary gear annular surface so as to be slidable against. The second ring gear or rotating part has an output side annular surface (921) which is an annular surface a part of which abuts against and faces a part of the second planetary gear annular surface so as to be slidable against.

When the first planetary gear annular surface abuts the sun gear annular surface, or when the second planetary gear annular surface abuts the output side annular surface, the carrier subassembly is restricted from moving relative to the housing in the axial direction.

In the present invention, the axial direction, i.e., the position of the carrier subassembly can be regulated by the sun gear annular surface of the sun gear and the output side annular surface of the second ring gear or the rotating part. Here, the sliding speed between the first planetary gear annular surface and the sun gear annular surface, and the sliding speed between the second planetary gear annular surface and the output side annular surface, are extremely low compared to the sliding speed between the carrier subassembly and the position regulation part in Patent Document 1. Therefore, the sliding distance between the first planetary gear annular surface and the sun gear annular surface, and the sliding distance between the second planetary gear annular surface and the output side annular surface can be extremely small, and the sliding stress can be significantly reduced. Therefore, the amount of wear of the planetary gear, sun gear, second ring gear, or the rotating part can be reduced, and the wear resistance can be improved.

In addition, torque loss between the carrier sub-assembly and the sun gear and second ring gear serving as position regulation components or between the rotating portion can be reduced, thereby improving the efficiency of the entire reducer.
The planetary gear has a cylindrical planetary gear body (322), a first protrusion (323) protruding cylindrically from one axial end face of the planetary gear body, and a second protrusion (324) protruding cylindrically from the other axial end face of the planetary gear body. The first planetary gear annular surface is formed on the end face of the first protrusion. The second planetary gear annular surface is formed on the end face of the second protrusion.

1 is a cross-sectional view showing a clutch device to which a geared motor and a clutch actuator according to a first embodiment are applied; 1 is a cross-sectional view showing a portion of a geared motor, a clutch actuator, and a clutch device according to a first embodiment; FIG. 2 is a cross-sectional view showing a portion of the clutch actuator according to the first embodiment. 2 is a schematic diagram showing a planetary gear and its vicinity of the geared motor according to the first embodiment; FIG. FIG. 11 is a cross-sectional view showing a part of a clutch actuator according to a comparative example. FIG. 13 is a schematic diagram showing a planetary gear and its vicinity of a geared motor according to a comparative example. FIG. 2 is a diagram showing a planetary gear and its vicinity of the geared motor according to the first embodiment. 8 is a schematic cross-sectional view taken along line VIII-VIII in FIG. 7 . 4 is a diagram showing the relationship between the rotation angle of the sun gear and the contact points between the planetary gear and the first and second ring gears; FIG. FIG. 11 is a diagram showing the relationship between the torque acting on the second ring gear and the axial load acting on the carrier. FIG. 1 is a diagram showing the relationship between the slip ratio and the coefficient of friction between two parts that rotate relative to one another. FIG. 11 is a cross-sectional view showing a portion of a clutch actuator according to a second embodiment. FIG. 11 is a cross-sectional view showing a portion of a clutch actuator according to a third embodiment. FIG. 13 is a cross-sectional view showing a portion of a clutch actuator according to a fourth embodiment. FIG. 13 is a cross-sectional view showing a portion of a clutch actuator according to a fifth embodiment. FIG. 13 is a cross-sectional view showing a planetary gear and its vicinity of a clutch actuator according to a fifth embodiment. FIG. 13 is a cross-sectional view showing a portion of a clutch actuator according to a sixth embodiment.

The clutch actuator according to several embodiments will be described below with reference to the drawings. Note that the same reference numerals are used to designate substantially the same components in several embodiments, and the description will be omitted.

First Embodiment
A clutch device to which a geared motor and a clutch actuator according to a first embodiment are applied is shown in Figures 1 and 2. The clutch device 1 is provided, for example, between an internal combustion engine and a transmission of a vehicle, and is used to allow or block the transmission of torque between the internal combustion engine and the transmission.

The clutch device 1 includes a clutch actuator 10, a clutch 70, an electronic control unit (hereinafter referred to as "ECU") 100 as a "control unit", an input shaft 61 as a "first transmission unit", an output shaft 62 as a "second transmission unit", etc.

The clutch actuator 10 includes a housing 12, an electric motor 20 as a "prime mover", a reduction gear 30, and a torque cam 2 as a "rotational translation part" or "rolling body cam". Here, the housing 12, the electric motor 20, the reduction gear 30, and a drive cam 40 that is part of the torque cam 2 described below constitute a geared motor 7.

The ECU 100 is a small computer having a CPU as a calculation means, a ROM, RAM, etc. as storage means, and an I/O as an input/output means. The ECU 100 executes calculations according to a program stored in the ROM, etc., based on information such as signals from various sensors provided in various parts of the vehicle, and controls the operation of various devices and equipment of the vehicle. In this way, the ECU 100 executes a program stored in a non-transitive physical recording medium. Execution of this program results in the execution of a method corresponding to the program.

The ECU 100 can control the operation of the internal combustion engine and the like based on information such as signals from various sensors. The ECU 100 can also control the operation of the electric motor 20, which will be described later.

The input shaft 61 is connected to, for example, the drive shaft of an internal combustion engine (not shown) and is rotatable together with the drive shaft. In other words, torque is input to the input shaft 61 from the drive shaft.

A fixed body 11 is provided in a vehicle equipped with an internal combustion engine (see FIG. 2). The fixed body 11 is formed, for example, in a cylindrical shape and is fixed to an engine room of the vehicle. A ball bearing 141 is provided between the inner peripheral wall of the fixed body 11 and the outer peripheral wall of the input shaft 61. As a result, the input shaft 61 is supported by the fixed body 11 via the ball bearing 141.

The housing 12 is provided between the inner peripheral wall of the fixed body 11 and the outer peripheral wall of the input shaft 61. The housing 12 has a housing inner cylindrical portion 121, a housing plate portion 122, a housing outer cylindrical portion 123, a seal groove portion 124, a housing step surface 125, a housing side spline groove portion 127, a housing hole portion 128, etc., which serve as a "housing cylindrical portion."

The housing inner cylinder portion 121 is formed in a generally cylindrical shape. The housing plate portion 122 is formed in an annular plate shape so as to extend radially outward from the end portion of the housing inner cylinder portion 121. The housing outer cylinder portion 123 is formed in a generally cylindrical shape so as to extend from the outer edge portion of the housing plate portion 122 to the same side as the housing inner cylinder portion 121. Here, the housing inner cylinder portion 121, the housing plate portion 122, and the housing outer cylinder portion 123 are integrally formed, for example, from a metal.

As described above, the housing 12 is generally hollow and flat.

The seal groove 124 is formed in an annular shape so as to be recessed radially inward from the outer peripheral wall of the housing inner cylinder 121. The housing step surface 125 is formed in an annular flat shape between the seal groove 124 and the housing plate 122 so as to face away from the housing plate 122.

The housing side spline grooves 127 are formed on the outer peripheral wall of the housing inner cylinder 121 so as to extend in the axial direction of the housing inner cylinder 121. A plurality of housing side spline grooves 127 are formed in the circumferential direction of the housing inner cylinder 121. The housing hole 128 is formed so as to penetrate the housing plate 122 in the plate thickness direction.

The housing 12 is fixed to the fixed body 11 so that a portion of the outer wall abuts against a portion of the wall surface of the fixed body 11 (see FIG. 2). Here, the housing 12 is provided coaxially with the fixed body 11 and the input shaft 61. Here, "coaxial" does not necessarily mean that the two axes are exactly coaxial, but also includes a state in which they are slightly eccentric or tilted (the same applies below).

The housing 12 has an accommodation space 120 as a "space." The accommodation space 120 is formed between the housing inner cylinder portion 121, the housing plate portion 122, and the housing outer cylinder portion 123.

The electric motor 20 is housed in the housing space 120. The electric motor 20 has a stator 21, a coil 22, a rotor 23, a magnet 230 as a "magnet", a magnet cover 24, etc.

The stator 21 has a stator yoke 211 and stator teeth 212. The stator 21 is formed, for example, from laminated steel plates. The stator yoke 211 is formed in a substantially cylindrical shape. The stator teeth 212 are formed integrally with the stator yoke 211 so as to protrude radially inward from the inner peripheral wall of the stator yoke 211. A plurality of stator teeth 212 are formed at equal intervals in the circumferential direction of the stator yoke 211. A coil 22 is provided on each of the plurality of stator teeth 212. The stator 21 is fixed to the housing 12 so that the outer peripheral wall of the stator yoke 211 fits into the inner peripheral wall of the housing outer cylinder portion 123.

The rotor 23 is formed of, for example, an iron-based metal. The rotor 23 has a rotor body 231 and a rotor cylinder portion 232. The rotor body 231 is formed in a substantially annular shape. The rotor cylinder portion 232 is formed to extend in a cylindrical shape from the outer edge of the rotor body 231.

The magnets 230 are provided on the outer peripheral wall of the rotor 23. Multiple magnets 230 are provided at equal intervals around the circumference of the rotor 23 so that their magnetic poles alternate.

The magnet cover 24 is provided on the rotor 23 so as to cover the radially outer surface of the magnet 230 of the rotor 23. More specifically, the magnet cover 24 is formed, for example, from a non-magnetic metal.

The clutch actuator 10 is equipped with a rotor bearing 15. The rotor bearing 15 is provided radially outside the housing inner cylinder portion 121 on the housing plate portion 122 side relative to the housing step surface 125. The rotor bearing 15 has an inner ring 151, an outer ring 152, bearing balls 153 as "bearing rolling elements", etc.

The inner ring 151 and the outer ring 152 are formed, for example, from metal in a cylindrical shape. The outer ring 152 is provided radially outside the inner ring 151. The bearing balls 153 are formed, for example, from metal in a spherical shape. The bearing balls 153 are provided so as to be able to roll between the inner ring 151 and the outer ring 152 in an annular groove formed in the outer peripheral wall of the inner ring 151 and an annular groove formed in the inner peripheral wall of the outer ring 152. A plurality of bearing balls 153 are provided in the circumferential direction of the inner ring 151 and the outer ring 152. The bearing balls 153 roll between the inner ring 151 and the outer ring 152, allowing the inner ring 151 and the outer ring 152 to rotate relative to each other. The bearing balls 153 restrict the relative movement of the inner ring 151 and the outer ring 152 in the axial direction.

The rotor bearing 15 is provided in the housing inner cylinder 121 with the inner peripheral wall of the inner ring 151 abutting against the outer peripheral wall of the housing inner cylinder 121 and one axial end face of the inner ring 151 spaced a predetermined distance from the housing plate 122. The rotor 23 is provided so that the inner peripheral wall of the rotor body 231 fits into the outer peripheral wall of the rotor bearing 15. As a result, the rotor bearing 15 supports the rotor 23 so that it can rotate relative to the housing 12.

The ECU 100 can control the operation of the electric motor 20 by controlling the power supplied to the coil 22. When power is supplied to the coil 22, a rotating magnetic field is generated in the stator 21, causing the rotor 23 to rotate. This causes torque to be output from the rotor 23. In this way, the electric motor 20 has the stator 21 and the rotor 23 that is arranged to be rotatable relative to the stator 21, and can output torque from the rotor 23 when power is supplied.

Here, the rotor 23 is arranged radially inside the stator 21 so as to be rotatable relative to the stator 21. The electric motor 20 is an inner rotor type brushless DC motor.

In this embodiment, the clutch actuator 10 is equipped with a rotation angle sensor 104. The rotation angle sensor 104 is provided on the electric motor 20 so as to be located on the housing plate portion 122 side relative to the coil 22.

The rotation angle sensor 104 detects the magnetic flux generated from the sensor magnet or magnet 230 that rotates integrally with the rotor 23, and outputs a signal corresponding to the detected magnetic flux to the ECU 100. This allows the ECU 100 to detect the rotation angle and rotation speed of the rotor 23 based on the signal from the rotation angle sensor 104. Furthermore, the ECU 100 can calculate the relative rotation angle of the drive cam 40 with respect to the housing 12 and a driven cam 50 (described later), and the axial relative position of the driven cam 50 with respect to the housing 12 and the drive cam 40, based on the rotation angle and rotation speed of the rotor 23, etc.

As shown in FIG. 3, the reducer 30 has a sun gear 31, a planetary gear 32, a carrier 33, a first ring gear 34, a second ring gear 35, etc.

The sun gear 31 is arranged coaxially with the rotor 23 so that they can rotate together. In other words, the rotor 23 and the sun gear 31 are formed separately from different materials and are arranged coaxially so that they can rotate together.

More specifically, the sun gear 31 has a sun gear base 310, a sun gear tooth portion 311 as a "tooth portion" and an "external tooth", a sun gear tube portion 312, and a sun gear extension portion 315. The sun gear 31 is formed, for example, from a metal. The sun gear base 310 is formed in a substantially circular ring shape. The sun gear tube portion 312 is formed integrally with the sun gear base 310 so as to extend in a cylindrical shape from the outer edge of the sun gear base 310. The sun gear tooth portion 311 is formed on the outer peripheral wall of the end portion of the sun gear tube portion 312 opposite the sun gear base 310. The sun gear extension portion 315 is formed in a ring shape so as to extend radially outward from the outer peripheral wall of the end portion of the sun gear tube portion 312 on the sun gear base 310 side.

The sun gear 31 is arranged so that the outer peripheral wall of the sun gear base 310 fits into the inner peripheral wall of the rotor cylinder 232. As a result, the sun gear 31, together with the rotor 23, is supported by the rotor bearing 15 so as to be rotatable relative to the housing 12.

The torque of the electric motor 20 is input to the sun gear 31, which rotates integrally with the rotor 23. Here, the sun gear 31 corresponds to the "input part" of the reducer 30.

Planetary gears 32 are provided along the circumferential direction of sun gear 31, and can revolve around the circumferential direction of sun gear 31 while rotating on their own axis while meshing with sun gear 31. More specifically, planetary gears 32 are formed, for example, from metal in a substantially cylindrical shape, and are provided radially outside sun gear 31 at equal intervals in the circumferential direction of sun gear 31. In this embodiment, four planetary gears 32 are provided. Planetary gear 32 has planetary gear tooth portion 321 as a "tooth portion" and an "external tooth". Planetary gear tooth portion 321 is formed on the outer peripheral wall of planetary gear 32 so as to be able to mesh with sun gear tooth portion 311.

The carrier 33 rotatably supports the planetary gear 32 and is rotatable relative to the sun gear 31.

More specifically, the carrier 33 has a carrier body 331. The carrier body 331 is formed, for example, from a metal into a substantially annular plate shape. The carrier body 331 is located between the coil 22 and the planetary gear 32 in the axial direction.

The reducer 30 has a pin 335 and a planetary gear bearing 36. The pin 335 is formed, for example, from a metal into a generally cylindrical shape. The pin 335 is provided so that its axial end is fixed to the carrier body 331.

The planetary gear bearing 36 is provided between the outer peripheral wall of the pin 335 and the inner peripheral wall of the planetary gear 32. As a result, the planetary gear 32 is rotatably supported by the pin 335 via the planetary gear bearing 36. That is, the pin 335 is provided at the center of rotation of the planetary gear 32 and rotatably supports the planetary gear 32. Furthermore, the planetary gear 32 and the pin 335 can move relative to each other in the axial direction within a predetermined range via the planetary gear bearing 36. In other words, the planetary gear 32 and the pin 335 have their relative axial movement range restricted to a predetermined range by the planetary gear bearing 36.

The first ring gear 34 has a first ring gear tooth portion 341 that is a tooth portion that can mesh with the planetary gear 32, and is fixed to the housing 12. More specifically, the first ring gear 34 is formed in a substantially cylindrical shape, for example, from metal. The first ring gear 34 is fixed to the housing 12 on the opposite side of the stator 21 from the housing plate portion 122, so that the outer edge portion fits into the inner peripheral wall of the housing outer cylinder portion 123. Therefore, the first ring gear 34 cannot rotate relative to the housing 12.

Here, the first ring gear 34 is arranged coaxially with the housing 12, the rotor 23, and the sun gear 31. The first ring gear tooth portion 341, which serves as a "tooth portion" and an "internal tooth", is formed on the inner peripheral wall of the first ring gear 34 so as to be able to mesh with one axial end side of the planetary gear tooth portion 321 of the planetary gear 32.

The second ring gear 35 has a second ring gear tooth portion 351 that is a tooth portion that can mesh with the planetary gear 32 and has a different number of teeth than the first ring gear tooth portion 341, and is arranged to be able to rotate together with the drive cam 40 described later. More specifically, the second ring gear 35 is formed into a cylindrical shape, for example, from metal.

Here, the second ring gear 35 is provided coaxially with the housing 12, the rotor 23, and the sun gear 31. The second ring gear tooth portion 351 as the "tooth portion" and the "internal teeth" is formed on the inner peripheral wall of the end portion of the second ring gear 35 on the first ring gear 34 side in the axial direction so as to be able to mesh with the other end portion of the planetary gear tooth portion 321 of the planetary gear 32 in the axial direction. In this embodiment, the number of teeth of the second ring gear tooth portion 351 is greater than the number of teeth of the first ring gear tooth portion 341. More specifically, the number of teeth of the second ring gear tooth portion 351 is greater than the number of teeth of the first ring gear tooth portion 341 by the number of planetary gears 32 multiplied by an integer.

In addition, since the planetary gear 32 must mesh normally without interference with the first ring gear 34 and the second ring gear 35, which have two different specifications at the same location, the design is such that one or both of the first ring gear 34 and the second ring gear 35 are shifted to keep the center distance of each gear pair constant.

With the above configuration, when the rotor 23 of the electric motor 20 rotates, the sun gear 31 rotates, and the planetary gear teeth 321 of the planetary gear 32 rotates around the sun gear 31 while meshing with the sun gear teeth 311, the first ring gear teeth 341, and the second ring gear teeth 351. Here, since the number of teeth of the second ring gear teeth 351 is greater than the number of teeth of the first ring gear teeth 341, the second ring gear 35 rotates relative to the first ring gear 34. Therefore, between the first ring gear 34 and the second ring gear 35, a minute difference rotation corresponding to the difference in the number of teeth between the first ring gear teeth 341 and the second ring gear teeth 351 is output as the rotation of the second ring gear 35. As a result, the torque from the electric motor 20 is decelerated by the reducer 30 and output from the second ring gear 35. In this way, the reducer 30 is capable of reducing and outputting the torque of the electric motor 20. In this embodiment, the reducer 30 constitutes a 3k type paradox planetary gear reducer.

The second ring gear 35 is formed separately from the drive cam 40 described below, and is arranged to be rotatable together with the drive cam 40. The second ring gear 35 reduces the torque from the electric motor 20 and outputs it to the drive cam 40. Here, the second ring gear 35 corresponds to the "output section" of the reducer 30.

The torque cam 2 has a driving cam 40 as a "rotating part", a driven cam 50 as a "translating part", and a cam ball 3 as a "cam rolling body".

The drive cam 40 has a drive cam body 41, a drive cam specific shape portion 42, a drive cam plate portion 43, a drive cam outer cylinder portion 44, a drive cam groove 400, etc. The drive cam body 41 is formed in a substantially circular plate shape. The drive cam specific shape portion 42 is formed so as to extend from the outer edge of the drive cam body 41 at an angle to the axis of the drive cam body 41. The drive cam plate portion 43 is formed in a substantially circular plate shape so as to extend radially outward from the end of the drive cam specific shape portion 42 opposite the drive cam body 41. The drive cam outer cylinder portion 44 is formed in a substantially cylindrical shape so as to extend from the outer edge of the drive cam plate portion 43 to the opposite side of the drive cam specific shape portion 42. Here, the drive cam body 41, the drive cam specific shape portion 42, the drive cam plate portion 43, and the drive cam outer cylinder portion 44 are integrally formed, for example, from metal.

The drive cam groove 400 is formed to extend in the circumferential direction of the drive cam body 41 while being recessed from one end face of the drive cam body 41 to the other end face. The drive cam groove 400 is formed so that the depth from one end face changes in the circumferential direction of the drive cam body 41. For example, three drive cam grooves 400 are formed at equal intervals in the circumferential direction of the drive cam body 41.

The drive cam 40 is provided between the housing inner cylinder portion 121 and the housing outer cylinder portion 123 so that the drive cam body 41 is located between the outer peripheral wall of the housing inner cylinder portion 121 and the inner peripheral wall of the sun gear cylinder portion 312 of the sun gear 31, and the drive cam plate portion 43 is located on the opposite side of the planetary gear 32 from the carrier body 331. The drive cam 40 is rotatable relative to the housing 12.

The second ring gear 35 is integral with the drive cam 40 so that the inner peripheral wall of the end opposite to the end where the second ring gear teeth portion 351 is formed fits into the outer edge of the drive cam plate portion 43. The second ring gear 35 cannot rotate relative to the drive cam 40. In other words, the second ring gear 35 is provided so as to be rotatable integrally with the drive cam 40 as the "rotating portion." Therefore, when the torque from the electric motor 20 is reduced by the reduction gear 30 and output from the second ring gear 35, the drive cam 40 rotates relative to the housing 12. In other words, the drive cam 40 rotates relative to the housing 12 when the torque output from the reduction gear 30 is input.

The driven cam 50 has a driven cam body 51, a driven cam specific shape portion 52, a driven cam plate portion 53, a cam side spline groove portion 54, a driven cam groove 500, etc. The driven cam body 51 is formed in a substantially circular plate shape. The driven cam specific shape portion 52 is formed so as to extend from the outer edge portion of the driven cam body 51 at an angle to the axis of the driven cam body 51. The driven cam plate portion 53 is formed in a substantially circular plate shape so as to extend radially outward from the end portion of the driven cam specific shape portion 52 opposite the driven cam body 51. Here, the driven cam body 51, the driven cam specific shape portion 52, and the driven cam plate portion 53 are integrally formed, for example, from metal.

The cam side spline grooves 54 are formed to extend in the axial direction on the inner peripheral wall of the driven cam body 51. Multiple cam side spline grooves 54 are formed in the circumferential direction of the driven cam body 51.

The driven cam 50 is arranged so that the driven cam body 51 is located on the opposite side of the rotor bearing 15 from the drive cam body 41 and radially inward of the drive cam specific shape portion 42 and the drive cam plate portion 43, and the cam side spline groove portion 54 is spline-engaged with the housing side spline groove portion 127. This makes the driven cam 50 non-rotatable relative to the housing 12, but movable axially relative to the housing 12.

The driven cam groove 500 is formed to extend in the circumferential direction of the driven cam body 51 while being recessed from one end face, which is the face of the driven cam body 51 facing the driving cam body 41, to the other end face. The driven cam groove 500 is formed so that the depth from one end face varies in the circumferential direction of the driven cam body 51. For example, three driven cam grooves 500 are formed at equal intervals in the circumferential direction of the driven cam body 51.

The driving cam groove 400 and the driven cam groove 500 are formed to have the same shape when viewed from the face side of the driving cam body 41 facing the driven cam body 51, or the face side of the driven cam body 51 facing the driving cam body 41.

The cam balls 3 are formed into a spherical shape, for example, from metal. The cam balls 3 are provided so as to be able to roll between the three driving cam grooves 400 and the three driven cam grooves 500. In other words, a total of three cam balls 3 are provided.

In this way, the driving cam 40, the driven cam 50, and the cam ball 3 constitute the torque cam 2 as a "rolling cam." When the driving cam 40 rotates relative to the housing 12 and the driven cam 50, the cam ball 3 rolls along the bottoms of the driving cam groove 400 and the driven cam groove 500.

As described above, the drive cam groove 400 and the driven cam groove 500 are formed so that their depths change in the circumferential direction of the drive cam 40 or the driven cam 50. Therefore, when the drive cam 40 rotates relative to the housing 12 and the driven cam 50 due to the torque output from the reducer 30, the cam ball 3 rolls in the drive cam groove 400 and the driven cam groove 500, and the driven cam 50 moves axially relative to the drive cam 40 and the housing 12, i.e., strokes.

In this way, the driven cam 50 has multiple driven cam grooves 500 formed on one end face so as to sandwich the cam ball 3 between the driven cam groove 400 and the driving cam groove 400, and constitutes the torque cam 2 together with the driving cam 40 and the cam ball 3. When the driving cam 40 rotates relative to the housing 12, the driven cam 50 moves relative to the driving cam 40 and the housing 12 in the axial direction. Here, the driven cam 50 does not rotate relative to the housing 12 because the cam side spline groove portion 54 is spline-coupled to the housing side spline groove portion 127. Furthermore, although the driving cam 40 rotates relative to the housing 12, it does not move relative to the housing 12 in the axial direction.

The torque cam 2 is provided on one axial side of the electric motor 20 and converts the rotational motion caused by the torque from the electric motor 20 into translational motion, which is axial relative movement with respect to the housing 12.

In this embodiment, the clutch actuator 10 is equipped with a return spring 55 and a return spring retainer 56 as "biasing members". The return spring 55 is, for example, a coil spring, and is provided radially outside the housing inner cylinder portion 121 on the side of the driven cam body 51 opposite the driving cam body 41. One end of the return spring 55 abuts against the surface of the driven cam body 51 opposite the driving cam body 41.

The return spring retainer 56 has a retainer inner cylinder portion 561, a retainer plate portion 562, and a retainer outer cylinder portion 563. The retainer inner cylinder portion 561 is formed in a substantially cylindrical shape. The retainer plate portion 562 is formed in an annular plate shape so as to extend radially outward from one end of the retainer inner cylinder portion 561. The retainer outer cylinder portion 563 is formed in a substantially cylindrical shape so as to extend from the outer edge of the retainer plate portion 562 to the same side as the retainer inner cylinder portion 561. The retainer inner cylinder portion 561, the retainer plate portion 562, and the retainer outer cylinder portion 563 are integrally formed, for example, from a metal.

The return spring retainer 56 is fixed to the housing inner cylinder 121 so that the inner peripheral wall of the retainer inner cylinder 561 fits into the outer peripheral wall of the housing inner cylinder 121. The other end of the return spring 55 abuts against the retainer plate 562 between the retainer inner cylinder 561 and the retainer outer cylinder 563.

The return spring 55 has a force that stretches in the axial direction. Therefore, the driven cam 50 is biased toward the drive cam body 41 by the return spring 55 with the cam ball 3 sandwiched between the driven cam 50 and the drive cam 40.

The output shaft 62 has a shaft portion 621, a plate portion 622, a tube portion 623, and a friction plate 624 (see FIG. 2). The shaft portion 621 is formed in a substantially cylindrical shape. The plate portion 622 is formed integrally with the shaft portion 621 so as to extend radially outward from one end of the shaft portion 621 in an annular plate shape. The tube portion 623 is formed integrally with the plate portion 622 so as to extend in a substantially cylindrical shape from the outer edge of the plate portion 622 to the opposite side of the shaft portion 621. The friction plate 624 is formed in a substantially annular plate shape and is provided on the end surface of the plate portion 622 on the tube portion 623 side. Here, the friction plate 624 cannot rotate relative to the plate portion 622. A clutch space 620 is formed inside the tube portion 623.

The end of the input shaft 61 passes through the inside of the housing inner cylinder portion 121 and is located on the opposite side of the driven cam 50 from the drive cam 40. The output shaft 62 is provided coaxially with the input shaft 61 on the opposite side of the driven cam 50 from the drive cam 40. A ball bearing 142 is provided between the inner peripheral wall of the shaft portion 621 and the outer peripheral wall of the end of the input shaft 61. As a result, the output shaft 62 is supported by the input shaft 61 via the ball bearing 142. The input shaft 61 and the output shaft 62 are rotatable relative to the housing 12.

The clutch 70 is provided between the input shaft 61 and the output shaft 62 in the clutch space 620. The clutch 70 has an inner friction plate 71, an outer friction plate 72, and a locking portion 701. The inner friction plate 71 is formed in a substantially annular plate shape, and a plurality of inner friction plates 71 are provided so as to be aligned in the axial direction between the input shaft 61 and the cylindrical portion 623 of the output shaft 62. The inner friction plate 71 is provided so that its inner edge is spline-connected to the outer peripheral wall of the input shaft 61. Therefore, the inner friction plate 71 cannot rotate relative to the input shaft 61, but can move relative to the input shaft 61 in the axial direction.

The outer friction plates 72 are formed in a substantially annular plate shape, and multiple outer friction plates 72 are provided between the input shaft 61 and the tubular portion 623 of the output shaft 62, lined up in the axial direction. Here, the inner friction plates 71 and the outer friction plates 72 are arranged alternately in the axial direction of the input shaft 61. The outer edge of the outer friction plate 72 is spline-coupled to the inner peripheral wall of the tubular portion 623 of the output shaft 62. Therefore, the outer friction plate 72 cannot rotate relative to the output shaft 62, but can move relative to the output shaft 62 in the axial direction. Of the multiple outer friction plates 72, the outer friction plate 72 located closest to the friction plate 624 can come into contact with the friction plate 624.

The locking portion 701 is formed in a substantially circular ring shape, and is provided so that its outer edge fits into the inner circumferential wall of the tubular portion 623 of the output shaft 62. The locking portion 701 can lock the outer edge of the outer friction plate 72 that is located closest to the driven cam 50 among the multiple outer friction plates 72. Therefore, the multiple outer friction plates 72 and the multiple inner friction plates 71 are prevented from falling off from the inside of the tubular portion 623. The distance between the locking portion 701 and the friction plate 624 is greater than the total plate thickness of the multiple outer friction plates 72 and the multiple inner friction plates 71.

In an engaged state in which the multiple inner friction plates 71 and the multiple outer friction plates 72 are in contact with each other, i.e. engaged, a frictional force is generated between the inner friction plates 71 and the outer friction plates 72, and the relative rotation between the inner friction plates 71 and the outer friction plates 72 is restricted according to the magnitude of the frictional force. On the other hand, in a non-engaged state in which the multiple inner friction plates 71 and the multiple outer friction plates 72 are separated from each other, i.e. not engaged, no frictional force is generated between the inner friction plates 71 and the outer friction plates 72, and the relative rotation between the inner friction plates 71 and the outer friction plates 72 is not restricted.

When the clutch 70 is engaged, the torque input to the input shaft 61 is transmitted to the output shaft 62 via the clutch 70. On the other hand, when the clutch 70 is disengaged, the torque input to the input shaft 61 is not transmitted to the output shaft 62.

In this way, the clutch 70 transmits torque between the input shaft 61 and the output shaft 62. When the clutch 70 is in an engaged state, it allows the transmission of torque between the input shaft 61 and the output shaft 62, and when the clutch 70 is in a disengaged state, it blocks the transmission of torque between the input shaft 61 and the output shaft 62.

In this embodiment, the clutch device 1 is a normally open type clutch device that is normally in a disengaged state.

The clutch actuator 10 includes a state change unit 80. The state change unit 80 includes a disc spring 81, a disc spring retainer 82, and a disc spring thrust bearing 83 as "elastic deformation units". The disc spring retainer 82 includes a retainer tube portion 821 and a retainer flange portion 822. The retainer tube portion 821 is formed in a substantially cylindrical shape. The retainer flange portion 822 is formed in an annular plate shape so as to extend radially outward from one end of the retainer tube portion 821. The retainer tube portion 821 and the retainer flange portion 822 are integrally formed, for example, from a metal. The disc spring retainer 82 is provided on the driven cam 50 so that the other end of the retainer tube portion 821 is connected to the end face of the driven cam plate portion 53 opposite to the drive cam 40. Here, the retainer tube portion 821 and the driven cam plate portion 53 are connected, for example, by welding.

The disc spring 81 is arranged so that its inner edge is located between the driven cam plate portion 53 and the retainer flange portion 822 on the radial outside of the retainer cylinder portion 821. The disc spring thrust bearing 83 is formed in an annular shape and is arranged between the driven cam plate portion 53 and the inner edge of the disc spring 81 on the radial outside of the retainer cylinder portion 821.

The disc spring retainer 82 is fixed to the driven cam 50 so that the retainer flange portion 822 can engage one axial end, i.e., the inner edge, of the disc spring 81. Therefore, the disc spring 81 and the disc spring thrust bearing 83 are prevented from falling off the disc spring retainer 82 by the retainer flange portion 822. The disc spring 81 is elastically deformable in the axial direction.

Figure 3 is a cross-sectional view showing the clutch actuator 10 without the state change unit 80 attached.

As shown in Figures 1 and 2, when the cam ball 3 is located at a position (origin) corresponding to the deepest part of the drive cam groove 400, which is the furthest part in the axial direction of the drive cam body 41, i.e., the depth direction, from one end face of the drive cam body 41, and at a position (origin) corresponding to the deepest part of the drive cam groove 500, which is the furthest part in the axial direction of the drive cam body 41, i.e., the depth direction, from one end face of the driven cam body 51, the driven cam groove 500, which is the furthest part in the axial direction of the driven cam body 51, the distance between the drive cam 40 and the driven cam 50 is relatively small, and a gap Sp1 is formed between the other end of the axial direction of the disc spring 81, i.e., the outer edge, and the clutch 70 (see Figure 1). Therefore, the clutch 70 is in a disengaged state, and the transmission of torque between the input shaft 61 and the output shaft 62 is interrupted.

Here, during normal operation to change the state of the clutch 70, when power is supplied to the coil 22 of the electric motor 20 under the control of the ECU 100, the electric motor 20 rotates, torque is output from the reducer 30, and the drive cam 40 rotates relative to the housing 12. As a result, the cam ball 3 rolls from a position corresponding to the deepest part to one side in the circumferential direction of the drive cam groove 400 and the driven cam groove 500. As a result, the driven cam 50 moves axially relative to the housing 12 while compressing the return spring 55, that is, moves toward the clutch 70 side. As a result, the disc spring 81 moves toward the clutch 70 side.

When the disc spring 81 moves toward the clutch 70 due to the axial movement of the driven cam 50, the gap Sp1 becomes smaller and the other axial end of the disc spring 81 comes into contact with the outer friction plate 72 of the clutch 70. When the driven cam 50 moves further in the axial direction after the disc spring 81 comes into contact with the clutch 70, the disc spring 81 elastically deforms in the axial direction and pushes the outer friction plate 72 toward the friction plate 624. As a result, the multiple inner friction plates 71 and the multiple outer friction plates 72 engage with each other, and the clutch 70 enters an engaged state. This allows the transmission of torque between the input shaft 61 and the output shaft 62.

At this time, the disc spring 81 rotates relative to the driven cam 50 and the disc spring retainer 82 while being supported by the disc spring thrust bearing 83. In this way, the disc spring thrust bearing 83 supports the disc spring 81 while receiving a load from the disc spring 81 in the thrust direction.

When the clutch transmission torque reaches the clutch required torque capacity, the ECU 100 stops the rotation of the electric motor 20. This causes the clutch 70 to enter an engaged state in which the clutch transmission torque is maintained at the clutch required torque capacity. In this way, the disc spring 81 of the state change unit 80 receives an axial force from the driven cam 50, and can change the state of the clutch 70 to an engaged state or a disengaged state depending on the axial position of the driven cam 50 relative to the housing 12 and the drive cam 40.

The torque cam 2 also converts the rotational motion caused by the torque from the electric motor 20 into translational motion, which is axial relative movement with respect to the housing 12, and can change the state of the clutch 70 between an engaged state and a disengaged state.

The end of the output shaft 62 opposite the plate portion 622 of the shaft portion 621 is connected to an input shaft of a transmission (not shown) and can rotate together with the input shaft. In other words, the torque output from the output shaft 62 is input to the input shaft of the transmission. The torque input to the transmission is changed in speed by the transmission and output as drive torque to the drive wheels of the vehicle. This causes the vehicle to run.

In this embodiment, the clutch device 1 includes an oil supply unit 5 (see FIGS. 1 and 2). The oil supply unit 5 is formed in a passage shape on the output shaft 62 so that one end is exposed to the clutch space 620. The other end of the oil supply unit 5 is connected to an oil supply source (not shown). As a result, oil is supplied from one end of the oil supply unit 5 to the clutch 70 in the clutch space 620.

The ECU 100 controls the amount of oil supplied from the oil supply unit 5 to the clutch 70. The oil supplied to the clutch 70 can lubricate and cool the clutch 70. Thus, in this embodiment, the clutch 70 is a wet clutch and can be cooled by oil.

In this embodiment, the torque cam 2 as the "rotation translation part" forms an accommodation space 120 between the drive cam 40 and the second ring gear 35 as the "rotation part" and the housing 12. Here, the accommodation space 120 is formed inside the housing 12 on the opposite side of the drive cam 40 and the second ring gear 35 from the clutch 70. The electric motor 20 and the reducer 30 are provided in the accommodation space 120. The clutch 70 is provided in a clutch space 620, which is a space on the opposite side of the drive cam 40 from the accommodation space 120.

The clutch actuator 10 is equipped with a thrust bearing 16. As shown in FIG. 1, the thrust bearing 16 has rollers 161, a race 162, and a backup plate 163 as "thrust bearing rolling elements". The race 162 is formed into an annular plate shape, for example, from metal. The rollers 161 are formed into a roughly cylindrical shape, for example, from metal, and are provided so as to be able to roll in the circumferential direction of the race 162 while contacting one end face of the race 162. A plurality of rollers 161 are provided in the circumferential direction of the race 162.

The backup plate 163 has a plate body 164 and a plate protrusion 165. The plate body 164 is formed in a substantially circular ring shape. The plate protrusion 165 is formed in a substantially circular ring shape so as to protrude in the axial direction from the inner edge of the plate body 164. The plate body 164 and the plate protrusion 165 are integrally formed, for example, from a metal.

The backup plate 163 is provided radially outside the housing inner cylinder 121 so that the plate protrusion 165 abuts against the housing step surface 125. The race 162 is provided radially outside the housing inner cylinder 121 so that the other end face abuts against the end face of the plate body 164 opposite the plate protrusion 165. The roller 161 is provided between the race 162 and the drive cam body 41, and can roll in the circumferential direction of the race 162 while contacting the end face of the race 162 on the drive cam body 41 side and the face of the drive cam body 41 on the race 162 side.

The thrust bearing 16 bears the drive cam 40 while receiving a load in the thrust direction, i.e., the axial direction, from the drive cam 40. In this embodiment, the axial load from the clutch 70 acts on the thrust bearing 16 via the disc spring 81, disc spring thrust bearing 83, driven cam 50, cam ball 3, and drive cam 40.

In this embodiment, the clutch actuator 10 includes an inner seal member 191 and an outer seal member 192 as "seal members." The inner seal member 191 is an oil seal formed into a ring shape from an elastic material such as rubber. The outer seal member 192 is an oil seal formed into a ring shape from an elastic material such as rubber and a metal ring.

The inner seal member 191 is provided in a seal groove portion 124 formed in the housing inner cylinder portion 121. The inner seal member 191 is provided in the seal groove portion 124 so that the outer edge portion can slide against the inner peripheral wall of the drive cam body 41.

The outer seal member 192 is provided between the housing outer cylinder portion 123 and the drive cam outer cylinder portion 44 on the opposite side of the second ring gear 35 from the first ring gear 34. The outer seal member 192 is provided on the housing outer cylinder portion 123 so that the seal lip portion of the inner edge can slide against the outer peripheral wall of the drive cam outer cylinder portion 44.

Here, the outer seal member 192 is positioned radially outward of the inner seal member 191 when viewed from the axial direction of the inner seal member 191 (see Figures 1 and 2).

As described above, the inner peripheral wall of the drive cam body 41 can slide against the inner seal member 191. In other words, the inner seal member 191 is provided so as to contact the drive cam 40 as the "rotating part." The inner seal member 191 provides an airtight or liquidtight seal between the drive cam body 41 and the housing inner cylinder portion 121.

The outer peripheral wall of the drive cam outer cylinder 44 can slide against the seal lip portion, which is the inner edge of the outer seal member 192. In other words, the outer seal member 192 is provided so as to come into contact with the drive cam 40, which serves as the "rotating part." The outer seal member 192 provides an airtight or liquidtight seal between the outer peripheral wall of the drive cam outer cylinder 44 and the inner peripheral wall of the housing outer cylinder 123.

The inner seal member 191 and the outer seal member 192 provided as described above can keep the accommodation space 120 that houses the electric motor 20 and the reducer 30 airtight or liquidtight, and can keep the space between the accommodation space 120 and the clutch space 620 in which the clutch 70 is provided airtight or liquidtight. As a result, even if foreign matter such as wear powder is generated in the clutch 70, the foreign matter can be prevented from entering the accommodation space 120 from the clutch space 620. Therefore, malfunction of the electric motor 20 or the reducer 30 due to foreign matter can be prevented.

The configuration of each part of this embodiment will be explained in more detail below.

As shown in FIG. 3, <1> the planetary gear 32, the pin 335, the planetary gear bearing 36, and the carrier 33 constitute the carrier subassembly 330.

At least one of the multiple planetary gears 32 has a first planetary gear annular surface 901, which is an annular surface, at one axial end, and a second planetary gear annular surface 902, which is an annular surface, at the other axial end.

More specifically, the planetary gear 32 has a planetary gear tooth portion 321, a planetary gear body 322, a first convex portion 323, and a second convex portion 324. The planetary gear body 322 is formed in a cylindrical shape. The first convex portion 323 is formed so as to protrude in a cylindrical shape from one end face in the axial direction of the planetary gear 32. The outer diameter of the first convex portion 323 is smaller than the outer diameter of the planetary gear body 322. The end face of the first convex portion 323 opposite the planetary gear body 322 is formed in an annular and flat shape. The second convex portion 324 is formed so as to protrude in a cylindrical shape from the other end face in the axial direction of the planetary gear 32. The outer diameter of the second convex portion 324 is smaller than the outer diameter of the planetary gear body 322. The end face of the second convex portion 324 opposite the planetary gear body 322 is formed in an annular and flat shape. The planetary gear teeth 321 are formed on the outer peripheral wall of the planetary gear body 322.

The first planetary gear annular surface 901 and the second planetary gear annular surface 902 are formed on the end faces of the first convex portion 323 and the second convex portion 324, respectively. Therefore, the first planetary gear annular surface 901 and the second planetary gear annular surface 902 are formed in an annular and flat shape. Each of the four planetary gears 32 has a first planetary gear annular surface 901 and a second planetary gear annular surface 902.

The sun gear 31 has a sun gear annular surface 911, a portion of which abuts against and faces a portion of the first planetary gear annular surface 901 in a slidable manner.

More specifically, the sun gear cylinder portion 312 has a sun gear large diameter portion 313 and a sun gear small diameter portion 314. The sun gear large diameter portion 313 is formed in a cylindrical shape on the sun gear base portion 310 side. The sun gear small diameter portion 314 is formed in a cylindrical shape integrally with the sun gear large diameter portion 313 on the opposite side of the sun gear base portion 310 with respect to the sun gear large diameter portion 313. The outer diameter of the sun gear small diameter portion 314 is smaller than the outer diameter of the sun gear large diameter portion 313. Therefore, an annular and planar step surface is formed between the outer peripheral wall of the sun gear large diameter portion 313 and the outer peripheral wall of the sun gear small diameter portion 314. The sun gear annular surface 911 is formed on this step surface. Therefore, the sun gear annular surface 911 is formed in an annular and planar shape.

The sun gear teeth portion 311 are formed on the outer peripheral wall of the sun gear small diameter portion 314. A portion of the sun gear annular surface 911 abuts against and faces a portion of the first planetary gear annular surface 901 so as to be capable of sliding against it. Therefore, when viewed from the direction of axis Ax1 of the sun gear 31 and the direction of axis Ax2 of the planetary gear 32, a portion of the sun gear annular surface 911 overlaps with a portion of the first planetary gear annular surface 901 (see the shaded portion in Figure 4).

The second ring gear 35 has an output side annular surface 921, a portion of which abuts against and faces a portion of the second planetary gear annular surface 902 in a slidable manner.

More specifically, the second ring gear 35 has a second ring gear tooth portion 351, a second ring gear body 352, and a second ring gear extension portion 353. The second ring gear body 352 is formed in a cylindrical shape. The second ring gear extension portion 353 is formed in an annular shape so as to extend radially inward from the inner peripheral wall of the second ring gear body 352 on the side opposite to the first ring gear 34. The end face of the second ring gear extension portion 353 on the first ring gear 34 side is formed in an annular and flat shape. The output side annular surface 921 is formed on the end face of the second ring gear extension portion 353 on the first ring gear 34 side. Therefore, the output side annular surface 921 is formed in an annular and flat shape.

The second ring gear tooth portion 351 is formed on the inner peripheral wall of the second ring gear body 352 on the first ring gear 34 side relative to the second ring gear extension portion 353. A portion of the output side annular surface 921 abuts against and faces a portion of the second planetary gear annular surface 902 so as to be capable of sliding. Therefore, when viewed from the axial direction of the planetary gear 32, a portion of the output side annular surface 921 and a portion of the second planetary gear annular surface 902 overlap.

When the first planetary gear annular surface 901 abuts against the sun gear annular surface 911, or when the second planetary gear annular surface 902 abuts against the output side annular surface 921, the carrier subassembly 330 is restricted from moving relative to the housing 12 in the axial direction.

This allows the axial position of the carrier subassembly 330 to be regulated by the sun gear annular surface 911 of the sun gear 31 and the output side annular surface 921 of the second ring gear 35.

<2> The outer diameters of the first planetary gear annular surface 901 and the second planetary gear annular surface 902 are set to be equal to or smaller than the diameter of the tooth root circle of the planetary gear 32. The outer diameter of the sun gear annular surface 911 is set to be equal to or larger than the diameter of the tooth tip circle of the sun gear 31. The inner diameter of the output side annular surface 921 is set to be equal to or smaller than the diameter of the tooth tip circle of the second ring gear 35.

More specifically, the outer diameters of the first planetary gear annular surface 901 and the second planetary gear annular surface 902 are set to be smaller than the diameter of the tooth root circle of the planetary gear tooth portion 321. The outer diameter of the sun gear annular surface 911 is set to be larger than the diameter of the tooth tip circle of the sun gear tooth portion 311. The inner diameter of the output side annular surface 921 is set to be smaller than the diameter of the tooth tip circle of the second ring gear tooth portion 351.

<4> The outer diameter of the sun gear annular surface 911 and the inner diameter of the output side annular surface 921 are set to a size that maintains the opposition between the first planetary gear annular surface 901 and the sun gear annular surface 911, and the opposition between the second planetary gear annular surface 902 and the output side annular surface 921, even if an error occurs in the center-to-center distance between the sun gear 31, the planetary gear 32, the first ring gear 34, and the second ring gear 35.

As shown in Figures 3 and 4, in this embodiment, when the sun gear 31 rotates due to the torque from the electric motor 20, the planetary gear 32 revolves in the circumferential direction of the sun gear 31 while meshing with the sun gear 31 and rotating on its axis. At this time, a part of the first planetary gear annular surface 901 and a part of the sun gear annular surface 911 can slide in the rotation direction of the planetary gear 32 (see the shaded area in Figure 4). The sliding speed at this time is extremely low. Also, at this time, a part of the second planetary gear annular surface 902 and a part of the output side annular surface 921 can slide in the rotation direction of the planetary gear 32. The sliding speed at this time is also extremely low.

As shown in Figures 5 and 6, in the comparative geared motor, the planetary gear 32 does not have a first convex portion 323, a second convex portion 324, a first planetary gear annular surface 901, or a second planetary gear annular surface 902. In addition, the sun gear annular surface 911 and the output side annular surface 921 are not formed.

The inner edge of the end surface of the coil 22 side of the carrier body 331 is provided so as to be able to abut and slide on the surface of the sun gear extension 315 on the planetary gear 32 side. Therefore, in the comparative embodiment, when the sun gear 31 rotates due to the torque from the electric motor 20, the planetary gear 32 revolves in the circumferential direction of the sun gear 31 while rotating while meshing with the sun gear 31, and the carrier body 331 rotates relative to the sun gear 31 on the radial outside of the sun gear 31. At this time, the inner edge of the end surface of the coil 22 side of the carrier body 331 and the surface of the sun gear extension 315 on the planetary gear 32 side can slide in the rotation direction of the carrier body 331 (see the shaded portion in FIG. 6). The sliding speed at this time is relatively high. Also, at this time, the tip of the pin 335 and the drive cam plate portion 43 can slide in the rotation direction of the carrier body 331. The sliding speed at this time is also relatively high. As a result, sliding stress is large, and there is a risk of severe wear between the sun gear 31 and carrier 33, and between the pin 335 and drive cam plate portion 43.

In contrast, in this embodiment, the sliding speed between the first planetary gear annular surface 901 and the sun gear annular surface 911, and the sliding speed between the second planetary gear annular surface 902 and the output side annular surface 921 are extremely low, so the sliding stress is small and wear between the planetary gear 32 and the sun gear 31 or the second ring gear 35 can be suppressed.

In addition, in this embodiment and the comparative embodiment, the sun gear 31, planetary gear 32, second ring gear 35, and drive cam 40 are heat-treated to increase their strength. In addition, the carrier body 331 and pin 335 are not heat-treated, and are weaker than the sun gear 31, planetary gear 32, second ring gear 35, and drive cam 40. Therefore, in the comparative embodiment in which the high-strength sun gear 31 slides against the low-strength carrier body 331 and the high-strength drive cam 40 slides against the low-strength pin 335, there is a risk that the carrier body 331 and pin 335 will wear out severely. On the other hand, in this embodiment in which the high-strength planetary gear 32 slides against the high-strength sun gear 31 or the high-strength second ring gear 35, wear of each part can be suppressed.

For example, even if the sun gear 31, planetary gear 32, second ring gear 35, and drive cam 40 are not heat treated, the amount of wear between the planetary gear 32 and the sun gear 31 or the second ring gear 35 in the configuration of this embodiment is about 1/10 of the amount of wear between the sun gear 31 and the carrier body 331, and between the drive cam 40 and the pin 335 in the configuration of the comparative embodiment. If the planetary gear 32, second ring gear 35, and drive cam 40 are heat treated, the amount of wear between the planetary gear 32 and the sun gear 31 or the second ring gear 35 in the configuration of this embodiment is about 1/100 to 1/200 of the amount of wear between the sun gear 31 and the carrier body 331, and between the drive cam 40 and the pin 335 in the configuration of the comparative embodiment.

As shown in Figures 7 and 8, when the sun gear 31 is rotated with the planetary gear 32 tilted due to the rattle of the planetary gear bearing 36, the contact parts between the planetary gear teeth 321 and the first and second ring gear teeth 341 and 351 slide in the axial direction, i.e., the axial direction (the direction indicated by y in Figure 8). The direction is the direction of the clutch 70 relative to the planetary gear 32 when the clutch 70 changes from a disengaged state to an engaged state, and the direction of the electric motor 20 relative to the planetary gear 32 when the clutch 70 changes from an engaged state to a disengaged state. The axial load acting in the axial direction at this time can be calculated as the product of the load acting on the tooth surface and the friction coefficient (μ).

7 and 8 show the load WC acting on the tooth surface of the planetary gear tooth portion 321 from the first ring gear tooth portion 341, the load WD acting on the tooth surface of the planetary gear tooth portion 321 from the second ring gear tooth portion 351, the contact point y1 between the first ring gear tooth portion 341 and the planetary gear tooth portion 321 before the start of sliding in the y direction, the contact point y1' during sliding, the contact point y2 between the second ring gear tooth portion 351 and the planetary gear tooth portion 321 before the start of sliding in the y direction, and the contact point y2' during sliding. Friction forces F1 and F2, which are axial loads, are generated in the direction opposite to the direction in which the contact point of the planetary gear 32 moves. Here, F1 = μ WC, F2 = μ WD.

The relationship between the rotation angle (θ) of the sun gear 31 and the positions (y) of the contact points y1 and y2 is shown in Figure 9. The position of contact point y1 is shown by a solid line, and the position of contact point y2 is shown by a dashed line.

Figure 10 shows the relationship between the torque T_out, which is the torque (load) acting on the second ring gear 35, and the axial load F_axial acting on the carrier 33. The solid line shows the axial load when the inclination of the planetary gear bearing 36 is relatively small, and the dashed line shows the axial load when the inclination of the planetary gear bearing 36 is relatively large. As shown in Figure 10, the greater the inclination of the planetary gear bearing 36, the greater the absolute value of the axial load.

The relationship between the slip rate and the friction coefficient between two parts that rotate relative to each other is shown in FIG. 11. In the comparative embodiment, the slip rate r2 between the sun gear 31 and the carrier body 331 and between the drive cam 40 and the pin 335 is relatively large, so the friction coefficient μ2 between the sun gear 31 and the carrier body 331 and between the drive cam 40 and the pin 335 is also relatively large. On the other hand, in this embodiment, the slip rate r1 between the planetary gear 32 and the sun gear 31 or the second ring gear 35 is significantly smaller than r2, so the friction coefficient μ1 between the planetary gear 32 and the sun gear 31 or the second ring gear 35 is also significantly smaller than μ2. Therefore, in this embodiment, the amount of wear of parts can be significantly reduced compared to the comparative embodiment.

As described above, in the present embodiment, <1> the second ring gear 35 has an output side annular surface 921, a portion of which abuts against and slidably faces a portion of the second planetary gear annular surface 902. When the first planetary gear annular surface 901 abuts against the sun gear annular surface 911, or when the second planetary gear annular surface 902 abuts against the output side annular surface 921, the carrier subassembly 330 is restricted from moving relative to the housing 12 in the axial direction.

In this embodiment, the axial direction, i.e., the position of the carrier subassembly 330 can be regulated by the sun gear annular surface 911 of the sun gear 31 and the output side annular surface 921 of the second ring gear 35. Here, the sliding speed between the first planetary gear annular surface 901 and the sun gear annular surface 911, and the sliding speed between the second planetary gear annular surface 902 and the output side annular surface 921 are extremely low compared to the sliding speed between the carrier subassembly and the position regulation part in Patent Document 1. Therefore, the sliding distance between the first planetary gear annular surface 901 and the sun gear annular surface 911, and the sliding distance between the second planetary gear annular surface 902 and the output side annular surface 921 can be extremely small, and the sliding stress can be significantly reduced. Therefore, the amount of wear of the planetary gear 32, the sun gear 31, and the second ring gear 35 can be reduced, and the wear resistance can be improved.

In addition, torque loss between the carrier subassembly 330 and the sun gear 31 and second ring gear 35, which serve as position control components, can be reduced, improving the efficiency of the entire reducer 30.

<2> In this embodiment, the outer diameters of the first planetary gear annular surface 901 and the second planetary gear annular surface 902 are set to be equal to or smaller than the diameter of the tooth root circle of the planetary gear 32. The outer diameter of the sun gear annular surface 911 is set to be equal to or larger than the diameter of the tooth tip circle of the sun gear 31. The inner diameter of the output side annular surface 921 is set to be equal to or smaller than the diameter of the tooth tip circle of the second ring gear 35.

As a result, when assembling the planetary gear 32, interference between the sun gear tooth portion 311 and the first planetary gear annular surface 901, and between the second ring gear tooth portion 351 and the second planetary gear annular surface 902 can be avoided, improving ease of assembly.

<4> In this embodiment, the outer diameter of the sun gear annular surface 911 and the inner diameter of the output side annular surface 921 are set to a size that maintains the opposition between the first planetary gear annular surface 901 and the sun gear annular surface 911, and the opposition between the second planetary gear annular surface 902 and the output side annular surface 921, even if an error occurs in the center-to-center distance between the sun gear 31, the planetary gear 32, the first ring gear 34, and the second ring gear 35.

As a result, even if an error occurs in the center distance between the sun gear 31, the planetary gear 32, the first ring gear 34, and the second ring gear 35, the planetary gear 32 can be prevented from falling off from between the sun gear annular surface 911 and the output side annular surface 921, and the axial position of the carrier subassembly 330 can be effectively and continuously regulated by the sun gear annular surface 911 and the output side annular surface 921.

<9> This embodiment is a clutch actuator 10 for use in a clutch device 1 having a clutch 70 that changes between an engaged state that allows torque transmission between the input shaft 61 and the output shaft 62 and a disengaged state that blocks torque transmission between the input shaft 61 and the output shaft 62, between the input shaft 61 and the output shaft 62, which are rotatable relative to each other, and includes the above-mentioned geared motor 7 and torque cam 2. The torque cam 2 has a drive cam 40, and converts the rotational motion of the drive cam 40 into translational motion, which is a relative axial movement with respect to the housing 12, and is capable of changing the state of the clutch 70 between an engaged state and a disengaged state.

In a clutch actuator, a high load is applied to the planetary gear and the like when the clutch is engaged. Therefore, in the clutch actuator of Patent Document 1, there is a risk that the amount of wear on the parts increases, the current to the motor increases, and efficiency and reliability decrease. On the other hand, the clutch actuator 10 of this embodiment is equipped with the above-mentioned geared motor 7, so the amount of wear on the parts is reduced, the current to the electric motor 20 is reduced, and efficiency and reliability can be improved.

Second Embodiment
A part of a geared motor and a clutch actuator according to the second embodiment is shown in Fig. 12. The second embodiment differs from the first embodiment in the configurations of a second ring gear 35 and a drive cam 40, etc.

In this embodiment, the second ring gear 35 does not have a second ring gear extension 353. Therefore, the second ring gear 35 has a cylindrical second ring gear main body 352 as its main component, so it can have a simple shape and can be easily manufactured by pressing or the like.

The drive cam 40 has a drive cam outermost cylinder portion 45. The drive cam outermost cylinder portion 45 is formed in a cylindrical shape integral with the outer edge portion of the drive cam plate portion 43. Here, the end portion of the drive cam outermost cylinder portion 45 on the planetary gear 32 side protrudes toward the planetary gear 32 side relative to the end face of the drive cam plate portion 43 on the planetary gear 32 side. The end face of the drive cam outermost cylinder portion 45 on the planetary gear 32 side is formed in an annular and flat shape.

The second ring gear 35 is rotatable together with the drive cam 40 so that the inner peripheral wall of the second ring gear body 352 opposite the second ring gear teeth portion 351 fits into the outer peripheral wall of the drive cam outermost cylinder portion 45.

The drive cam 40 has an output side annular surface 921, a portion of which abuts against and faces a portion of the second planetary gear annular surface 902 in a slidable manner.

The output side annular surface 921 is formed on the end surface of the drive cam outermost cylinder portion 45 on the planetary gear 32 side. Therefore, the output side annular surface 921 is formed in an annular and planar shape. A portion of the output side annular surface 921 abuts against and faces a portion of the second planetary gear annular surface 902 so as to be able to slide. Therefore, when viewed from the axial direction of the planetary gear 32, a portion of the output side annular surface 921 and a portion of the second planetary gear annular surface 902 overlap.

When the first planetary gear annular surface 901 abuts against the sun gear annular surface 911, or when the second planetary gear annular surface 902 abuts against the output side annular surface 921, the carrier subassembly 330 is restricted from moving relative to the housing 12 in the axial direction.

This allows the axial position of the carrier subassembly 330 to be regulated by the sun gear annular surface 911 of the sun gear 31 and the output side annular surface 921 of the drive cam 40.

As described above, <1> in this embodiment, the drive cam 40 has an output side annular surface 921, a portion of which abuts against and slidably faces a portion of the second planetary gear annular surface 902. When the first planetary gear annular surface 901 abuts against the sun gear annular surface 911, or when the second planetary gear annular surface 902 abuts against the output side annular surface 921, the carrier subassembly 330 is restricted from moving relative to the housing 12 in the axial direction.

As a result, the same effects as in the first embodiment can be achieved.

Third Embodiment
A part of a geared motor and a clutch actuator according to the third embodiment is shown in Fig. 13. The third embodiment differs from the second embodiment in the configurations of the rotor 23 and the sun gear 31, etc.

In this embodiment, the rotor 23 and the sun gear 31 are integrally formed from the same material. The sun gear 31 does not have a sun gear base 310 or a sun gear extension 315. The rotor 23 and the sun gear 31 are integrally formed as a whole in a cylindrical shape by connecting the rotor tube 232 and the sun gear large diameter portion 313 of the sun gear tube 312.

In this embodiment, the sun gear extension 315, which was necessary in the comparative embodiment to restrict the axial movement of the carrier 33, is not necessary, so the configuration of the sun gear 31 can be simplified. In addition, by forming the rotor 23 and the sun gear 31 integrally, the number of parts can be reduced while simplifying the shape of the parts.

Fourth Embodiment
A part of a geared motor and a clutch actuator according to the fourth embodiment is shown in Fig. 14. The fourth embodiment differs from the third embodiment in the configuration of the carrier 33 and the like.

<5> In this embodiment, the carrier 33 is provided on the drive cam 40 side with respect to the axial center of the planetary gear 32.

More specifically, the carrier body 331 of the carrier 33 is disposed on the drive cam plate portion 43 side of the drive cam 40 with respect to the axial center of the planetary gear body 322, and between the drive cam specific shape portion 42 and the drive cam outermost cylinder portion 45.

As a result, compared to the third embodiment, the planetary gear 32 can be positioned closer to the coil 22 by, for example, the thickness of the carrier body 331, and the axial length of the clutch actuator 10 can be shortened by d1.

A predetermined gap is formed between the carrier body 331 and the pin 335 and the drive cam plate portion 43. Therefore, the carrier body 331 and the pin 335 cannot abut against or slide on the drive cam plate portion 43.

<6> In this embodiment, the carrier 33 is made of a magnetic material.

As described above, in this embodiment, <5> the carrier 33 is provided on the drive cam 40 side with respect to the axial center of the planetary gear 32. Therefore, the planetary gear 32 can be positioned closer to the coil 22 side, and the axial length of the clutch actuator 10 can be shortened.

<6> In this embodiment, the carrier 33 is made of a magnetic material.

If the carrier 33 is made of a magnetic material and is provided on the coil 22 side relative to the axial center of the planetary gear 32, the magnet 230 of the electric motor 20 and the carrier 33 may come close to each other, causing a load that attracts the carrier 33 to the electric motor 20 side. Therefore, in this embodiment, the carrier 33 is made of a magnetic material that is easy to process, and is provided on the drive cam 40 side relative to the axial center of the planetary gear 32, thereby reducing material costs and suppressing the generation of the above load.

Fifth Embodiment
A part of a geared motor and a clutch actuator according to the fifth embodiment is shown in Fig. 15. The fifth embodiment differs from the fourth embodiment in the configuration near the first ring gear 34 and the like.

<3> This embodiment further includes an annular plate 95 provided on the opposite side of the first ring gear 34 from the second ring gear 35.

More specifically, the plate 95 is formed into an annular plate shape, for example, from metal. The plate 95 is provided so that its outer edge is sandwiched between a stepped surface formed on the inner peripheral wall of the housing outer cylinder portion 123 and the first ring gear 34.

The plate 95 has a plate annular surface 931, a portion of which is an annular surface that abuts against and faces the first planetary gear annular surface 901 in a slidable manner.

More specifically, the plate annular surface 931 is formed in an annular, planar shape on the inner edge of the end face of the plate 95 that faces the first ring gear 34 and the planetary gear 32.

When the first planetary gear annular surface 901 abuts against the plate annular surface 931, or when the second planetary gear annular surface 902 abuts against the output side annular surface 921, the carrier subassembly 330 is restricted from moving relative to the housing 12 in the axial direction.

This allows the axial or axial position of the carrier subassembly 330 to be regulated by the sun gear annular surface 911 of the sun gear 31 and the plate annular surface 931 of the plate 95, and the output side annular surface 921 of the second ring gear 35.

16 , when an axial load F_axial1 acts on the planetary gear teeth 321 from the second ring gear teeth 351 or the first ring gear teeth 341, an axial load F_axial2 equivalent to F_axial1 acts on the output side annular surface 921 from the plate annular surface 931. The moment M2 acting on the planetary gear 32 is expressed as follows:
M2=R_axial×F_axial2−R_pitch×F_axial1
Here, the distance R_pitch between the position where the axial load F_axial1 acts and the axis Ax2 of the planetary gear 32 and the distance R_axial between the position where the axial load F_axial2 acts and the axis Ax2 of the planetary gear 32 are substantially the same (R_axial ≈ R_pitch), so
M2 ≒ 0
As a result, almost no moment is generated in the planetary gear 32.

As described above, <3> this embodiment further includes an annular plate 95 provided on the opposite side of the first ring gear 34 from the second ring gear 35. The plate 95 has a plate annular surface 931, a portion of which is an annular surface that abuts against and slidably faces the first planetary gear annular surface 901. When the first planetary gear annular surface 901 abuts against the plate annular surface 931, or when the second planetary gear annular surface 902 abuts against the output side annular surface 921, the carrier subassembly 330 is restricted from moving relative to the housing 12 in the axial direction.

In this embodiment, the axial direction, i.e., the position in the axial direction, of the planetary gear 32 can be regulated by sandwiching the first planetary gear annular surface 901 and the second planetary gear annular surface 902 between the plate annular surface 931 and the output side annular surface 921. This reduces the moment that can be generated in the planetary gear 32. This reduces the stress on the planetary gear bearing 36, improving reliability.

Sixth Embodiment
A part of a geared motor and a clutch actuator according to the sixth embodiment is shown in Fig. 17. The sixth embodiment differs from the first embodiment in the configuration of a sun gear annular surface 911 and the like.

<7> This embodiment further includes an annular recess 900. The annular recess 900 is provided on the sun gear annular surface 911. The annular recess 900 is formed in an annular shape so as to be recessed from the sun gear annular surface 911, i.e., the step surface between the outer peripheral wall of the sun gear large diameter portion 313 and the outer peripheral wall of the sun gear small diameter portion 314, toward the rotor 23.

By forming an annular recess 900 in the sun gear 31, the inertia of the sun gear 31 can be reduced and costs can be reduced.

Other Embodiments
In the above embodiment, an example is shown in which four planetary gears 32 are provided. However, in other embodiments, any number of planetary gears 32 may be provided as long as it is two or more. Here, it is preferable that the planetary gears 32 are provided at equal intervals in the circumferential direction of the carrier 33.

In the above embodiment, an example was shown in which all of the multiple planetary gears 32 have the first planetary gear annular surface 901 and the second planetary gear annular surface 902. In contrast, in <8> other embodiments, some of the multiple planetary gears 32 may have the first planetary gear annular surface 901 and the second planetary gear annular surface 902. In this case, the number of processing steps for the first planetary gear annular surface 901 and the second planetary gear annular surface 902 can be reduced. Here, for example, when six planetary gears 32 are provided, a form in which three of the six planetary gears 32 have the first planetary gear annular surface 901 and the second planetary gear annular surface 902 is considered. It is desirable that the three planetary gears 32 having the first planetary gear annular surface 901 and the second planetary gear annular surface 902 are arranged at equal intervals around the circumference of the carrier 33, with one planetary gear 32 not having the first planetary gear annular surface 901 and the second planetary gear annular surface 902 sandwiched between them.

In the sixth embodiment, an example was shown in which the annular recess 900 was provided on the sun gear annular surface 911. In contrast, in <7> other embodiments, the annular recess 900 may be provided on at least one of the first planetary gear annular surface 901, the second planetary gear annular surface 902, the sun gear annular surface 911, and the output side annular surface 921. This can reduce the inertia of the part in which the annular recess 900 is provided, and can also reduce costs.

In the above embodiment, the carrier is made of a magnetic material. In contrast, in other embodiments, the carrier may be made of a non-magnetic material, such as stainless steel.

In other embodiments, the number of drive cam grooves 400 and driven cam grooves 500 may be three or more. The number of cam balls 3 may also be set to match the number of drive cam grooves 400 and driven cam grooves 500.

Furthermore, the above-mentioned multiple embodiments may be combined in any way as long as there are no structural impediments.

The present invention is not limited to vehicles that run on driving torque from an internal combustion engine, but can also be applied to electric vehicles and hybrid vehicles that run on driving torque from a motor.

In another embodiment, torque may be input from the "second transmission unit" and output from the "first transmission unit" via the "clutch." For example, if one of the "first transmission unit" or the "second transmission unit" is fixed so that it cannot rotate, the rotation of the other of the "first transmission unit" or the "second transmission unit" can be stopped by engaging the "clutch." In this case, the clutch device can be used as a brake device.

The geared motor of the present invention can be used not only as the motor part of a clutch actuator, but also as the motor part of other devices or as a standalone motor.

As such, the present disclosure is not limited to the above-described embodiments, but can be implemented in various forms without departing from the spirit of the present disclosure.

The clutch device control unit and the method thereof described in the present disclosure may be realized by a dedicated computer provided by configuring a processor and a memory programmed to execute one or more functions embodied in a computer program. Alternatively, the clutch device control unit and the method thereof described in the present disclosure may be realized by a dedicated computer provided by configuring a processor with one or more dedicated hardware logic circuits. Alternatively, the clutch device control unit and the method thereof described in the present disclosure may be realized by one or more dedicated computers configured by combining a processor and a memory programmed to execute one or more functions and a processor configured with one or more hardware logic circuits. In addition, the computer program may be stored in a computer-readable non-transient tangible recording medium as instructions executed by the computer.

7 Geared motor, 12 Housing, 20 Electric motor, 31 Sun gear, 32 Planetary gear, 36 Planetary gear bearing, 33 Carrier, 34 First ring gear, 35 Second ring gear, 40 Drive cam (rotating part), 330 Carrier sub-assembly, 335 Pin, 901 First planetary gear annular surface, 902 Second planetary gear annular surface, 911 Sun gear annular surface, 921 Output side annular surface

Claims (9)

A housing (12);
an electric motor (20) provided in the housing and capable of outputting torque when energized;
A reducer (30) capable of reducing and outputting torque from the electric motor;
A rotating part (40) that rotates by torque from the reduction gear,
The reducer includes:
a sun gear (31) to which torque from the electric motor is input;
a plurality of planetary gears (32) that are meshed with the sun gear and can revolve around the circumferential direction of the sun gear while rotating;
A pin (335) provided at the rotation center of the planetary gear;
a planetary gear bearing (36) provided between the planetary gear and the pin;
an annular carrier (33) that rotatably supports the planetary gear by supporting one end of the pin and is rotatable relative to the sun gear;
a first ring gear (34) that is annular and can mesh with the planetary gear; and
a second ring gear (35) that is annular and can mesh with the planetary gear and is formed so that the number of teeth of the tooth portion is different from that of the first ring gear, and outputs torque to the rotating portion;
the planetary gear, the pin, the planetary gear bearing and the carrier constitute a carrier subassembly (330);
At least one of the plurality of planetary gears has a first planetary gear annular surface (901) which is an annular surface at one end in the axial direction, and a second planetary gear annular surface (902) which is an annular surface at the other end in the axial direction,
The sun gear has a sun gear annular surface (911) which is an annular surface a part of which abuts against and slidably faces a part of the first planetary gear annular surface,
The second ring gear or the rotating portion has an output side annular surface (921) which is an annular surface a part of which abuts against and slidably faces a part of the second planetary gear annular surface,
when the first planetary gear annular surface abuts against the sun gear annular surface or when the second planetary gear annular surface abuts against the output side annular surface, the carrier sub-assembly is restricted from moving relative to the housing in an axial direction ,
The planetary gear has a cylindrical planetary gear body (322), a first convex portion (323) protruding in a cylindrical shape from one end face in the axial direction of the planetary gear body, and a second convex portion (324) protruding in a cylindrical shape from the other end face in the axial direction of the planetary gear body,
The first planetary gear annular surface is formed on an end surface of the first convex portion,
A geared motor , wherein the second planetary gear annular surface is formed on an end surface of the second convex portion . an outer diameter of the first planetary gear annular surface and an outer diameter of the second planetary gear annular surface are set to be smaller than a diameter of a tooth root circle of the planetary gear;
An outer diameter of the annular surface of the sun gear is set to be equal to or larger than a diameter of a tip circle of the sun gear,
2. The geared motor according to claim 1, wherein an inner diameter of the output side annular surface is set to be equal to or smaller than a diameter of a tip circle of the second ring gear. The gearbox further includes an annular plate (95) provided on the opposite side of the first ring gear to the second ring gear,
The plate has a plate annular surface (931) which is an annular surface a part of which abuts against and slidably faces the first planetary gear annular surface,
3. A geared motor as described in claim 1 or 2, wherein the carrier subassembly is restricted in its axial movement relative to the housing when the first planetary gear annular surface abuts the plate annular surface or when the second planetary gear annular surface abuts the output side annular surface. The geared motor according to any one of claims 1 to 3, wherein the outer diameter of the sun gear annular surface and the inner diameter of the output side annular surface are set to a size that maintains the opposition between the first planetary gear annular surface and the sun gear annular surface, and the opposition between the second planetary gear annular surface and the output side annular surface, even if an error occurs in the center distance between the sun gear, the planetary gear, the first ring gear, and the second ring gear. 5. The geared motor according to claim 1, wherein the carrier is provided only on the rotating portion side with respect to the axial center of the planetary gear. The geared motor according to claim 5, wherein the carrier is made of a magnetic material. The geared motor according to any one of claims 1 to 6 further comprises an annular recess (900) that is an annular recess provided on at least one of the first planetary gear annular surface, the second planetary gear annular surface, the sun gear annular surface, and the output side annular surface. A geared motor according to any one of claims 1 to 7, wherein some of the multiple planetary gears have the first planetary gear annular surface and the second planetary gear annular surface. A clutch actuator (10) for use in a clutch device (1) including a clutch (70) between a first transmission part (61) and a second transmission part (62) that are rotatable relative to one another, the clutch (70) changing its state between an engaged state that allows transmission of torque between the first transmission part and the second transmission part and a disengaged state that interrupts transmission of torque between the first transmission part and the second transmission part,
A geared motor (7) according to any one of claims 1 to 8;
a rotation-translation unit (2) having the rotating unit and capable of converting a rotational motion of the rotating unit into a translational motion that is a relative movement in an axial direction with respect to the housing, and changing a state of the clutch into an engaged state or a disengaged state;
A clutch actuator comprising:

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