JP2022119644A - clutch device - Google Patents

Abstract

To provide a clutch device having high durability and responsiveness.SOLUTION: A motor 20 having a stator 21 provided in a housing 12, and a rotor 23 provided rotatable relative to the stator 21 can output toque from the rotor 23 with power supplied. A bearing part 151 has a plurality of bearing rolling elements 173 for rolling in the peripheral direction of the rotor 23 to rotatably support the rotor 23, and lubricant 174 for lubricating areas around the bearing rolling elements 173. Only one bearing part 151 is provided for rotatably supporting the rotor 23. A speed reducer 30 has a sun gear tooth part 311 provided rotatable integrally with the rotor 23 and coaxial therewith for torque to be input thereto from the rotor 23.SELECTED DRAWING: Figure 1

Description

The present invention relates to clutch devices.

Conventionally, a clutch device is known that permits or blocks transmission of torque between a first transmission section and a second transmission section by changing the state of the clutch to an engaged state or a non-engaged state.

For example, the clutch device described in Patent Literature 1 includes a prime mover, a speed reducer, a rotary translator, a clutch, and a state changer. A prime mover outputs torque from a rotor when electric power is supplied. The reducer reduces the torque of the prime mover and outputs it. The torque output from the speed reducer is input to the rotation translation unit. The state changing portion receives an axial force from the rotational translation portion and is capable of changing the state of the clutch to an engaged state or a disengaged state.

WO2015/068822

By the way, in the clutch device of Patent Document 1, the speed reducer is a so-called eccentric cycloid speed reducer. Here, the input portion of the speed reducer is formed integrally with the rotor of the prime mover so as to be eccentric with respect to the rotation axis of the rotor. Therefore, a radial load is applied to the rotor as the input section swings. Therefore, in order to ensure the durability of the rotor bearing, a relatively large load capacity is required. In the clutch device of Patent Document 1, two bearings are provided: a first bearing that rotatably supports the end of the rotor opposite to the input portion, and a second bearing that rotatably supports the end of the rotor on the input portion side. It has one bearing.

On the other hand, the responsiveness of an actuator having an electric prime mover as a drive source, especially at low temperatures, greatly depends on the rotational torque of the bearing that supports the rotor. Therefore, it is thought that reducing the rotational torque of the bearing will lead to higher response at low temperatures. However, when employing an eccentric speed reducer that requires a bearing to withstand a relatively large load, if the number of rolling elements in the bearing is reduced in order to reduce the rotational torque of the bearing, there is a risk that the durability will decrease.

In addition, in a clutch device having an electric prime mover as a drive source, it is required to quickly remove the load transmitted to the clutch, that is, to release the clutch when the power supply fails due to disconnection of the power line of the prime mover. Sometimes. In this case, the driven torque, which is the torque applied to the rotor from the clutch side, such as the reaction force from the clutch or the minimum load of the return spring, must be greater than the starting torque of the bearing and the cogging torque of the prime mover.

The starting torque of a bearing consists of the radial load, axial load, bending moment, the term proportional to the magnetic force as the radial load generated between the rotor and stator of the prime mover, and the and a term that depends on the magnitude of the resistance to shear the lubricant. The latter increases in the low temperature range where the kinematic viscosity of the lubricant increases. Therefore, when power fails at low temperatures, it may be difficult to remove the load being transmitted to the clutch.

SUMMARY OF THE INVENTION An object of the present invention is to provide a clutch device with high durability and high responsiveness.

The clutch device according to the present invention includes a housing (12), a prime mover (20), a speed reducer (30), a rotation translation section (2), a clutch (70), state change sections (80, 90), and bearing sections (151, 151, 151, 151). 152). The prime mover has a stator (21) provided in the housing and a rotor (23) provided rotatably relative to the stator, and can be operated by energization to output torque from the rotor. A speed reducer (30) can reduce the torque of the prime mover and output it.

The rotation translation part includes a rotation part (40) that rotates relative to the housing when the torque output from the speed reducer is input, and a translation part that moves axially relative to the housing when the rotation part rotates relative to the housing. It has a part (50). The clutch is provided between a first transmission portion (61) and a second transmission portion (62) which are rotatable relative to the housing, and when engaged, the first transmission portion and the second transmission portion are engaged. and, in the disengaged state, interrupts torque transmission between the first transmission portion and the second transmission portion.

The state changing portion receives an axial force from the translating portion and is capable of changing the state of the clutch to an engaged state or a disengaged state depending on the relative axial position of the translating portion with respect to the housing. The bearing portion has a plurality of bearing rolling elements (173, 183) that roll in the circumferential direction of the rotor and rotatably support the rotor, and lubricants (174, 184) that lubricate the surroundings of the bearing rolling elements. Here, only one bearing portion is provided to rotatably support the rotor. The speed reducer has an input section (311) which is rotatable integrally with the rotor and is provided coaxially to receive torque from the rotor.

In the present invention, when torque is input from the prime mover to the input section, the input section rotates coaxially with the rotor. Therefore, it is possible to reduce the radial load acting on the input portion from gears or the like provided in the radial direction of the input portion. Therefore, the number of bearings that rotatably support the rotor can be reduced to one.

In addition, since the radial load acting on the input portion can be reduced, deterioration in durability can be suppressed even if the number of bearing rolling elements in the bearing portion is reduced. Therefore, the starting torque and rotation torque of the bearing can be reduced. As a result, responsiveness is improved particularly at low temperatures, and the minimum set load required to remove the load on the clutch during power failure can be reduced.

Also, by reducing the number of bearing rolling elements in the bearing portion, the moment of inertia of the rotor can be reduced, and the responsiveness can be further improved.

Sectional drawing which shows the clutch apparatus by 1st Embodiment. Sectional drawing which shows a part of clutch apparatus by 1st Embodiment. Schematic diagram of a 2-kh type paradox planetary gear reducer, and a table showing the relationship between input/output patterns, moment of inertia, and reduction ratio. Schematic diagram of a 3k-type paradox planetary gear reducer, and a table showing the relationship between the input/output pattern, the moment of inertia, and the reduction ratio. FIG. 4 is a diagram showing the relationship between the stroke of the translation portion and the load acting on the clutch; Sectional drawing which shows the bearing part of the clutch apparatus by 1st Embodiment. FIG. 10 is a diagram for explaining the effect of reducing the number of bearing rolling elements in a bearing, in which (A) is a diagram showing the bearing rolling elements rolling in the holding hole; The figure which shows the state which removed the rolling element. The figure which shows the relationship between the atmospheric temperature and the starting torque of a bearing part. FIG. 4 is a diagram showing the relationship between the number of bearing rolling elements and the starting torque of the bearing or the withstand load of the bearing; FIG. 4 is a diagram showing return spring load, ball cam load, and rotor detent torque acting on or occurring in the clutch device according to the first embodiment; FIG. 5 is a diagram showing the relationship between the stroke amount of the translation part and the magnitude of the load with respect to the return spring load, the ball cam load, and the clutch load. FIG. 4 is a diagram showing the balance of torque acting on or occurring in the clutch device according to the first embodiment; FIG. 5 is a diagram showing ACT load hiss characteristics, which is the relationship between the motor torque and the load during normal operation and reverse operation of the clutch device according to the first embodiment; The figure which shows the rotation speed at the time of starting of the prime mover of 1st Embodiment and a comparative form. The figure which shows the relationship between the rotation speed of the motor|power_engine of 1st Embodiment and a comparative form, and a rotational torque. FIG. 2 is a schematic cross-sectional view showing part of the clutch device according to the first embodiment; FIG. 4 is a diagram for explaining the resultant force acting on the input portion of the speed reducer; FIG. 5 is a diagram for explaining the torque sharing ratio of each planetary gear of the speed reducer and the resultant force acting on the input portion of the speed reducer; Sectional drawing which shows the clutch apparatus by 2nd Embodiment. FIG. 5 is a schematic cross-sectional view showing part of a clutch device according to a third embodiment; FIG. 11 is a schematic cross-sectional view showing part of a clutch device according to a fourth embodiment; Sectional drawing which shows a part of clutch apparatus by 5th Embodiment. Sectional drawing which shows a part of clutch apparatus by 6th Embodiment.

Hereinafter, clutch devices according to a plurality of embodiments will be described based on the drawings. In addition, the same code|symbol is attached|subjected to the substantially same structural part in several embodiment, and description is abbreviate|omitted.

(First embodiment)
A clutch device according to a first embodiment is shown in FIGS. A 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 transmission of torque between the internal combustion engine and the transmission.

The clutch device 1 includes a housing 12, a motor 20 as a "prime mover", a speed reducer 30, a ball cam 2 as a "rotational translation unit" or a "rolling cam", a clutch 70, a state changing unit 80, A bearing portion 151 is provided.

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

The ECU 10 is a small computer having a CPU as computing means, a ROM, a RAM, etc. as storage means, and an I/O as input/output means. The ECU 10 executes calculations according to programs stored in a ROM or the like based on information such as signals from various sensors provided in various parts of the vehicle, and controls operations of various devices and devices of the vehicle. Thus, the ECU 10 executes the program stored in the non-transitional substantive recording medium. By executing this program, the method corresponding to the program is executed.

The ECU 10 can control the operation of the internal combustion engine based on information such as signals from various sensors. The ECU 10 can also control the operation of a motor 20, which will be described later.

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

A vehicle equipped with an internal combustion engine is provided with a fixed body 11 (see FIG. 2). The fixed body 11 is formed, for example, in a tubular shape and fixed to the 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 . Thereby, the input shaft 61 is supported by the fixed body 11 via the ball bearings 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 cylinder portion 121, a housing plate portion 122, a housing outer cylinder portion 123, a housing small plate portion 124, a housing step surface 125, a housing small inner cylinder portion 126, a housing side spline groove portion 127, and the like. .

The housing inner cylindrical portion 121 is formed in a substantially cylindrical shape. The housing small plate portion 124 is formed in an annular plate shape so as to extend radially outward from the end portion of the housing inner cylindrical portion 121 . The housing small inner cylindrical portion 126 is formed in a substantially cylindrical shape so as to extend from the outer edge portion of the housing small plate portion 124 to the side opposite to the housing inner cylindrical portion 121 . The housing plate portion 122 is formed in an annular plate shape so as to extend radially outward from the end of the housing small inner cylindrical portion 126 opposite to the housing small plate portion 124 . Housing outer tubular portion 123 is formed in a substantially cylindrical shape so as to extend from the outer edge of housing plate portion 122 to the same side as housing small inner tubular portion 126 and housing inner tubular portion 121 . Here, the housing inner cylindrical portion 121, the housing small plate portion 124, the housing small inner cylindrical portion 126, the housing plate portion 122, and the housing outer cylindrical portion 123 are integrally formed of metal, for example.

As described above, the housing 12 is formed in a hollow and flat shape as a whole.

The housing stepped surface 125 is formed in an annular planar shape on the surface of the housing small plate portion 124 opposite to the housing small inner cylindrical portion 126 . The housing-side spline groove portion 127 is formed in the outer peripheral wall of the housing inner tubular portion 121 so as to extend in the axial direction of the housing inner tubular portion 121 . A plurality of housing-side spline groove portions 127 are formed in the circumferential direction of the housing inner tubular portion 121 .

The housing 12 is fixed to the fixed body 11 so that a part of the outer wall contacts a part of the wall surface of the fixed body 11 (see FIG. 2). The housing 12 is fixed to the fixed body 11 by bolts (not shown) or the like. Here, housing 12 is provided coaxially with fixed body 11 and input shaft 61 . A substantially cylindrical space is formed between the inner peripheral wall of the housing inner cylindrical portion 121 and the outer peripheral wall of the input shaft 61 .

The housing 12 has an accommodation space 120 . The accommodation space 120 is formed between the housing inner cylindrical portion 121 , the housing small plate portion 124 , the housing small inner cylindrical portion 126 , the housing plate portion 122 and the housing outer cylindrical portion 123 .

The motor 20 is housed in the housing space 120 . The motor 20 has a stator 21, a rotor 23, and the like. The stator 21 has a stator core 211 and coils 22 . The stator core 211 is formed, for example, from laminated steel plates in a substantially annular shape, and is fixed inside the housing outer tubular portion 123 . Coil 22 is provided for each of the plurality of salient poles of stator core 211 .

The motor 20 has a magnet 230 as a "permanent magnet". The rotor 23 is made of, for example, iron-based metal and has a substantially annular shape. More specifically, the rotor 23 is made of pure iron, which has relatively high magnetic properties, for example.

Magnet 230 is provided on the outer peripheral wall of rotor 23 . A plurality of magnets 230 are provided at equal intervals in the circumferential direction of the rotor 23 so that the magnetic poles alternate.

The bearing portion 151 is provided on the outer peripheral wall of the housing small inner cylindrical portion 126 . A sun gear 31 , which will be described later, is provided on the radially outer side of the bearing portion 151 . The rotor 23 is provided radially outside the sun gear 31 so as not to rotate relative to the sun gear 31 . Bearing portion 151 is provided in accommodation space 120 and rotatably supports sun gear 31 , rotor 23 and magnet 230 .

Here, the rotor 23 is provided radially inside the stator core 211 of the stator 21 so as to be relatively rotatable with respect to the stator 21 . The motor 20 is an inner rotor type brushless DC motor.

The configuration and the like of the bearing portion 151 will be described in detail later.

The ECU 10 can control the operation of the motor 20 by controlling the electric power supplied to the coil 22 . When power is supplied to the coil 22, a rotating magnetic field is generated in the stator core 211, causing the rotor 23 to rotate. As a result, torque is output from the rotor 23 . As described above, the motor 20 has a stator 21 and a rotor 23 that is rotatable relative to the stator 21, and can output torque from the rotor 23 when electric power is supplied.

In this embodiment, the clutch device 1 has a rotation angle sensor 104 . The rotation angle sensor 104 is provided in the accommodation space 120 .

The rotation angle sensor 104 detects magnetic flux generated from a sensor magnet that rotates together with the rotor 23 and outputs a signal corresponding to the detected magnetic flux to the ECU 10 . Accordingly, the ECU 10 can detect the rotation angle, rotation speed, etc. of the rotor 23 based on the signal from the rotation angle sensor 104 . The ECU 10 also determines the relative rotation angle of the drive cam 40 with respect to the housing 12 and a driven cam 50 described later, and the relative rotation angle of the driven cam 50 and the state changer 80 with respect to the housing 12 and the drive cam 40, based on the rotation angle and rotation speed of the rotor 23 . A relative position in the axial direction and the like can be calculated.

The speed reducer 30 is housed in the housing space 120 . The reduction gear 30 has a sun gear 31, a planetary gear 32, a carrier 33, a first ring gear 34, a second ring gear 35, and the like.

The sun gear 31 is provided so as to be coaxial with the rotor 23 and integrally rotatable. That is, the rotor 23 and the sun gear 31 are formed separately and arranged coaxially so as to be rotatable together.

More specifically, the sun gear 31 has a sun gear main body 310 , sun gear tooth portions 311 as “tooth portions” and “external teeth”, and gear side groove portions 315 . The sun gear main body 310 is made of metal, for example, and has a substantially cylindrical shape. Gear-side groove portion 315 is formed so as to extend in the axial direction in the outer peripheral wall of sun gear main body 310 on one end side. A plurality of gear-side grooves 315 are formed in the circumferential direction of sun gear body 310 . One end of sun gear body 310 is supported by bearing 151 .

A groove corresponding to the gear side groove 315 is formed in the inner peripheral wall of the rotor 23 . The rotor 23 is positioned radially outward of one end of the sun gear 31 , and is provided such that the groove portion is coupled with the gear-side groove portion 315 . As a result, the rotor 23 cannot rotate relative to the sun gear 31 .

The sun gear tooth portion 311 is formed on the outer peripheral wall of the sun gear 31 on the other end side. The torque of the motor 20 is input to the sun gear 31 that rotates integrally with the rotor 23 . Here, the sun gear tooth portion 311 of the sun gear 31 corresponds to the “input portion” of the speed reducer 30 . In this embodiment, the sun gear 31 is made of steel, for example.

A plurality of planetary gears 32 are provided along the circumferential direction of the sun gear 31 , and can revolve in the circumferential direction of the sun gear 31 while rotating while meshing with the sun gear 31 . More specifically, the planetary gears 32 are formed of metal, for example, in a substantially cylindrical shape, and are provided on the radially outer side of the sun gear 31 at equal intervals in the circumferential direction of the sun gear 31 . The planetary gear 32 has planetary gear teeth 321 as "teeth" and "external teeth". The planetary gear tooth portion 321 is formed on the outer peripheral wall of the planetary gear 32 so as to mesh with the sun gear tooth portion 311 .

Carrier 33 rotatably supports planetary gear 32 and is rotatable relative to sun gear 31 . More specifically, the carrier 33 is provided radially outside the sun gear 31 . Carrier 33 is rotatable relative to rotor 23 and sun gear 31 .

The carrier 33 has a carrier body 330 and pins 331 . The carrier main body 330 is formed of metal, for example, in a substantially annular shape. Carrier body 330 is positioned between sun gear 31 and coil 22 in the radial direction, and is positioned between rotor 23 and magnet 230 and planetary gear 32 in the axial direction. The planetary gear 32 is located on the side opposite to the housing plate portion 122 with respect to the carrier body 330 and the coil 22 .

The pin 331 has a connection portion 335 and a support portion 336 . The connection portion 335 and the support portion 336 are each formed of, for example, metal in a columnar shape. The connection portion 335 and the support portion 336 are integrally formed so that their axes are shifted and parallel to each other. Therefore, the connecting portion 335 and the supporting portion 336 have a crank-like cross-sectional shape on a virtual plane containing each axis (see FIG. 1).

The pin 331 is fixed to the carrier body 330 so that a connection portion 335, which is a part on one end side, is connected to the carrier body 330. As shown in FIG. Here, the support portion 336 is provided on the opposite side of the carrier body 330 from the rotor 23 and the magnets 230 so that the axis thereof is positioned radially outward of the carrier body 330 with respect to the axis of the connection portion 335 (FIG. 1). reference). A total of four pins 331 are provided corresponding to the number of planetary gears 32 .

The speed reducer 30 has a planetary gear bearing 36 . The planetary gear bearing 36 is, for example, a needle bearing, and is provided between the outer peripheral wall of the support portion 336 of the pin 331 and the inner peripheral wall of the planetary gear 32 . Thereby, the planetary gear 32 is rotatably supported by the support portion 336 of the pin 331 via the planetary gear bearing 36 .

The first ring gear 34 has a first ring gear tooth portion 341 which 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 made of metal, for example, and has a substantially annular shape. The first ring gear 34 is fixed to the housing 12 on the side opposite to the housing plate portion 122 with respect to the coil 22 so that the outer edge thereof 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 provided coaxially with the housing 12 , the rotor 23 and the sun gear 31 . A first ring gear tooth portion 341 as a “tooth portion” and an “internal tooth” is formed on the inner edge portion 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 . ing.

The second ring gear 35 has a second ring gear tooth portion 351 which is a tooth portion that can mesh with the planetary gear 32 and has a number of teeth different from that of the first ring gear tooth portion 341. ing. More specifically, the second ring gear 35 is made of metal, for example, and has a substantially annular shape. The second ring gear 35 has a gear inner tubular portion 355 , a gear plate portion 356 and a gear outer tubular portion 357 . The gear inner cylindrical portion 355 is formed in a substantially cylindrical shape. The gear plate portion 356 is formed in an annular plate shape extending radially outward from one end of the gear inner cylindrical portion 355 . The gear outer cylindrical portion 357 is formed in a substantially cylindrical shape so as to extend from the outer edge portion of the gear plate portion 356 to the side opposite to the gear inner cylindrical portion 355 .

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 a "tooth portion" and an "internal tooth" is formed on the inner peripheral wall of the gear outer cylindrical portion 357 so as to be able to mesh with the other axial end side of the planetary gear tooth portion 321 of the planetary gear 32. It is 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 obtained by multiplying 4, which is the number of the planetary gears 32, by an integer.

In addition, the planetary gear 32 needs to 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. It is designed to shift and keep the center distance of each gear pair constant.

With the above configuration, when the rotor 23 of the motor 20 rotates, the sun gear 31 rotates, and the planetary gear tooth portion 321 of the planetary gear 32 rotates while meshing with the sun gear tooth portion 311, the first ring gear tooth portion 341, and the second ring gear tooth portion 351. While rotating, it revolves in the circumferential direction of the sun gear 31 . Here, since 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 , 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 differential rotation corresponding to the difference in the number of teeth between the first ring gear tooth portion 341 and the second ring gear tooth portion 351 is output as the rotation of the second ring gear 35. . As a result, the torque from the motor 20 is reduced by the speed reducer 30 and output from the second ring gear 35 . Thus, the speed reducer 30 can reduce the torque of the motor 20 and output it. In this embodiment, the speed reducer 30 constitutes a 3k paradox planetary gear speed reducer.

The second ring gear 35 is formed separately from a drive cam 40 to be described later, and is provided so as to be rotatable together with the drive cam 40 . The second ring gear 35 reduces torque from the motor 20 and outputs it to the drive cam 40 . Here, the second ring gear 35 corresponds to the “output section” of the speed reducer 30 .

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

The drive cam 40 has a drive cam main body 41, a drive cam inner cylindrical portion 42, a drive cam plate portion 43, a drive cam outer cylindrical portion 44, a drive cam groove 400, and the like. The drive cam main body 41 is formed in a substantially annular plate shape. The drive cam inner cylindrical portion 42 is formed in a substantially cylindrical shape so as to extend axially from the outer edge portion of the drive cam main body 41 . The drive cam plate portion 43 is formed in a substantially annular plate shape so as to extend radially outward from the end portion of the drive cam inner cylindrical portion 42 opposite to the drive cam main body 41 . The drive cam outer cylinder portion 44 is formed in a substantially cylindrical shape so as to extend from the outer edge portion of the drive cam plate portion 43 to the side opposite to the drive cam inner cylinder portion 42 . Here, the drive cam main body 41, the drive cam inner cylindrical portion 42, the drive cam plate portion 43, and the drive cam outer cylindrical portion 44 are integrally formed of metal, for example.

The drive cam groove 400 is formed to extend in the circumferential direction while being recessed from the surface of the drive cam main body 41 on the side of the drive cam inner cylindrical portion 42 . For example, five drive cam grooves 400 are formed at equal intervals in the circumferential direction of the drive cam main body 41 . The drive cam groove 400 is formed such that the groove bottom is inclined with respect to the surface of the drive cam body 41 on the side of the drive cam inner cylindrical portion 42 so that the depth becomes shallower from one end to the other end in the circumferential direction of the drive cam body 41 . It is

In the drive cam 40, the drive cam body 41 is positioned between the outer peripheral wall of the housing inner cylindrical portion 121 and the inner peripheral wall of the sun gear 31, and the drive cam plate portion 43 is positioned on the side opposite to the carrier body 330 with respect to the planetary gear 32. It is provided between the housing inner tubular portion 121 and the housing outer tubular portion 123 so as to be able to do so. The drive cam 40 is rotatable relative to the housing 12 .

The second ring gear 35 is provided integrally with the drive cam 40 so that the inner peripheral wall of the gear inner cylindrical portion 355 fits into the outer peripheral wall of the drive cam outer cylindrical portion 44 . The second ring gear 35 is non-rotatable relative to the drive cam 40 . That is, the second ring gear 35 is provided so as to be integrally rotatable with the drive cam 40 as a "rotating portion". Therefore, when the torque from the motor 20 is reduced by the speed reducer 30 and output from the second ring gear 35 , the drive cam 40 rotates relative to the housing 12 . That is, the drive cam 40 rotates relative to the housing 12 when the torque output from the speed reducer 30 is input.

The driven cam 50 has a driven cam main body 51, a driven cam cylindrical portion 52, a cam-side spline groove portion 54, a driven cam groove 500, and the like. The driven cam main body 51 is formed in a substantially annular plate shape. The driven cam tubular portion 52 is formed in a substantially cylindrical shape so as to extend axially from the outer edge of the driven cam main body 51 . Here, the driven cam main body 51 and the driven cam cylindrical portion 52 are integrally formed of metal, for example.

The cam-side spline groove portion 54 is formed on the inner peripheral wall of the driven cam main body 51 so as to extend in the axial direction. A plurality of cam-side spline groove portions 54 are formed in the circumferential direction of the driven cam main body 51 .

The driven cam 50 has a driven cam main body 51 located on the opposite side of the drive cam main body 41 from the housing stepped surface 125 and radially inside the drive cam inner cylindrical portion 42 and the drive cam plate portion 43 . is provided to be spline-coupled with the housing-side spline groove portion 127 . As a result, the driven cam 50 is non-rotatable relative to the housing 12 and is axially movable relative to the housing 12 .

The driven cam groove 500 is formed to extend in the circumferential direction while being recessed from the surface of the driven cam body 51 on the drive cam body 41 side. For example, five driven cam grooves 500 are formed at equal intervals in the circumferential direction of the driven cam main body 51 . The driven cam groove 500 is formed such that the groove bottom is inclined with respect to the surface of the driven cam body 51 on the drive cam body 41 side so that the depth becomes shallower from one end to the other end in the circumferential direction of the driven cam body 51 . there is

When the drive cam groove 400 and the driven cam groove 500 are viewed from the surface of the drive cam main body 41 on the side of the driven cam main body 51 or from the surface of the driven cam main body 51 on the side of the drive cam main body 41, They are formed to have the same shape.

The ball 3 is made of metal, for example, and has a spherical shape. Balls 3 are provided so as to be able to roll in each of five drive cam grooves 400 and five driven cam grooves 500 . That is, a total of five balls 3 are provided.

Thus, the drive cam 40, the driven cam 50 and the balls 3 constitute the ball cam 2 as a "rolling cam". When the driving cam 40 rotates relative to the housing 12 and the driven cam 50, the balls 3 roll along the groove bottoms of the driving cam groove 400 and the driven cam groove 500, respectively.

As shown in FIG. 1 , the balls 3 are provided radially inside the first ring gear 34 and the second ring gear 35 . More specifically, most of the balls 3 are provided within the axial range of the first ring gear 34 and the second ring gear 35 .

As described above, the drive cam groove 400 is formed such that the groove bottom is inclined from one end to the other end. Also, the driven cam groove 500 is formed such that the groove bottom is inclined from one end to the other end. Therefore, when the drive cam 40 rotates relative to the housing 12 and the driven cam 50 due to the torque output from the speed reducer 30, the balls 3 roll in the drive cam groove 400 and the driven cam groove 500, and the driven cam 50 rotates. Axial relative movement or stroke relative to cam 40 and housing 12 .

Thus, driven cam 50 moves axially relative to drive cam 40 and housing 12 as drive cam 40 rotates relative to housing 12 . Here, the driven cam 50 does not rotate relative to the housing 12 because the cam-side spline groove portion 54 is spline-connected to the housing-side spline groove portion 127 . Further, although the drive cam 40 rotates relative to the housing 12, it does not move relative to the housing 12 in the axial direction.

In this embodiment, the clutch device 1 includes a return spring 55 , a return spring retainer 56 and a C ring 57 . The return spring 55 is, for example, a coil spring. It is One end of the return spring 55 is in contact with the surface of the driven cam body 51 opposite to the drive cam body 41 .

The return spring retainer 56 is formed of metal, for example, in a substantially annular shape, and is in contact with the other end of the return spring 55 on the radially outer side of the housing inner cylindrical portion 121 . The C-ring 57 is fixed to the outer peripheral wall of the housing inner tubular portion 121 so as to engage the surface of the inner edge portion of the return spring retainer 56 opposite to the driven cam main body 51 .

The return spring 55 has force extending in the axial direction. Therefore, the driven cam 50 is urged toward the drive cam body 41 by the return spring 55 with the 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 cylindrical 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 in an annular plate shape from one end of the shaft portion 621 . The cylindrical portion 623 is formed integrally with the plate portion 622 so as to extend in a substantially cylindrical shape from the outer edge portion of the plate portion 622 to the side opposite to 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 cylinder portion 623 side. Here, the friction plate 624 cannot rotate relative to the plate portion 622 . A clutch space 620 is formed inside the cylindrical portion 623 .

The end of the input shaft 61 passes through the housing inner cylindrical portion 121 and is located on the opposite side of the driven cam 50 to the driving 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 at the end of the input shaft 61 . Thereby, the output shaft 62 is supported by the input shaft 61 via the ball bearings 142 . The input shaft 61 and the output shaft 62 are rotatable relative to the housing 12 .

Clutch 70 is provided between input shaft 61 and output shaft 62 in 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 plates 71 are formed in a substantially annular plate shape, and are provided in plurality 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 such that the inner edge thereof is spline-connected to the outer peripheral wall of the input shaft 61 . Therefore, the inner friction plate 71 is non-rotatable relative to the input shaft 61 and is axially movable relative to the input shaft 61 .

The outer friction plates 72 are formed in a substantially annular plate shape, and are provided in plurality so as to be aligned in the axial direction between the input shaft 61 and the cylindrical portion 623 of the output shaft 62 . Here, the inner friction plates 71 and the outer friction plates 72 are alternately arranged in the axial direction of the input shaft 61 . The outer friction plate 72 is provided so that its outer edge is spline-connected to the inner peripheral wall of the cylindrical portion 623 of the output shaft 62 . Therefore, the outer friction plate 72 is non-rotatable relative to the output shaft 62 and is axially movable relative to the output shaft 62 . The outer friction plate 72 located closest to the friction plate 624 among the plurality of outer friction plates 72 can contact the friction plate 624 .

The engaging portion 701 is formed in a substantially annular shape, and is provided so that the outer edge thereof fits into the inner peripheral wall of the tubular portion 623 of the output shaft 62 . The locking portion 701 can lock the outer edge portion of the outer friction plate 72 positioned closest to the driven cam 50 among the plurality of outer friction plates 72 . Therefore, the plurality of outer friction plates 72 and the plurality of inner friction plates 71 are prevented from coming off from the inside of the tubular portion 623 . The distance between locking portion 701 and friction plate 624 is greater than the total thickness of outer friction plates 72 and inner friction plates 71 .

In the engaged state in which the plurality of inner friction plates 71 and the plurality of outer friction plates 72 are in contact with each other, that is, in the engaged state, a friction force is generated between the inner friction plates 71 and the outer friction plates 72, and the friction force is generated. Relative rotation between the inner friction plate 71 and the outer friction plate 72 is regulated according to the size of . On the other hand, in the non-engaged state in which the plurality of inner friction plates 71 and the plurality of outer friction plates 72 are separated from each other, that is, the plurality of outer friction plates 72 are not engaged with each other, the frictional force between the inner friction plates 71 and the outer friction plates 72 is The relative rotation between the inner friction plate 71 and the outer friction plate 72 is not restricted.

When the clutch 70 is engaged, 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 in the disengaged state, the torque input to the input shaft 61 is not transmitted to the output shaft 62 .

Thus, clutch 70 transmits torque between input shaft 61 and output shaft 62 . Clutch 70 allows transmission of torque between input shaft 61 and output shaft 62 when in the engaged state, and allows transmission of torque between input shaft 61 and output shaft 62 when in the disengaged state. It interrupts transmission of torque to and from shaft 62 .

In this embodiment, the clutch device 1 is a so-called normally open type clutch device that is normally in a non-engaged state.

The state changing portion 80 has a disc spring 81, a disc spring retainer 82, and a thrust bearing 83 as an "elastic deformation portion". The disc spring retainer 82 has a retainer tubular portion 821 and a retainer flange portion 822 . The retainer tubular portion 821 is formed in a substantially cylindrical shape. The retainer flange portion 822 is formed in an annular plate shape extending radially outward from one end of the retainer cylinder portion 821 . The retainer tubular portion 821 and the retainer flange portion 822 are integrally formed of metal, for example. The disk spring retainer 82 is fixed to the driven cam 50 so that the outer peripheral wall of the other end of the retainer tubular portion 821 fits into the inner peripheral wall of the driven cam tubular portion 52 .

The disk spring 81 is provided so that the inner edge portion is positioned between the driven cam tubular portion 52 and the retainer flange portion 822 on the radially outer side of the retainer tubular portion 821 . The thrust bearing 83 is provided between the driven cam cylinder portion 52 and the disc spring 81 .

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

When the ball 3 is positioned at one end of the drive cam groove 400 and the driven cam groove 500, the distance between the drive cam 40 and the driven cam 50 is relatively small, and the other end of the disc spring 81 in the axial direction, that is, the outer edge portion, and the clutch 70 are separated from each other. A gap Sp1 is formed between (see FIG. 1). Therefore, the clutch 70 is in a disengaged state, and transmission of torque between the input shaft 61 and the output shaft 62 is interrupted.

Here, when power is supplied to the coil 22 of the motor 20 under the control of the ECU 10 , the motor 20 rotates, torque is output from the speed reducer 30 , and the drive cam 40 rotates relative to the housing 12 . As a result, the ball 3 rolls from one end of the driving cam groove 400 and the driven cam groove 500 to the other end. Therefore, the driven cam 50 moves axially relative to the housing 12 while compressing the return spring 55 , that is, moves toward the clutch 70 . As a result, the disk spring 81 moves toward the clutch 70 side.

When the disk spring 81 moves toward the clutch 70 due to the axial movement of the driven cam 50 , the gap Sp<b>1 becomes smaller and the other axial end of the disk spring 81 contacts the outer friction plate 72 of the clutch 70 . When the driven cam 50 further moves in the axial direction after the disc spring 81 contacts the clutch 70 , the disc spring 81 pushes the outer friction plate 72 toward the friction plate 624 while elastically deforming in the axial direction. As a result, the plurality of inner friction plates 71 and the plurality of outer friction plates 72 are engaged with each other, and the clutch 70 is engaged. Therefore, transmission of torque between the input shaft 61 and the output shaft 62 is allowed.

At this time, the disk spring 81 rotates relative to the driven cam 50 and the disk spring retainer 82 while being supported by the thrust bearing 83 . In this manner, the thrust bearing 83 bears the disk spring 81 while receiving a thrust-direction load from the disk spring 81 .

The ECU 10 stops rotation of the motor 20 when the clutch transmission torque reaches the clutch required torque capacity. As a result, the clutch 70 enters an engagement holding state in which the clutch transmission torque is maintained at the clutch required torque capacity. In this manner, the disc spring 81 of the state changing portion 80 receives an axial force from the driven cam 50 and engages the clutch 70 according to the axial relative position of the driven cam 50 with respect to the housing 12 and the drive cam 40 . It can be changed to engaged or disengaged.

The end of the shaft portion 621 opposite to the plate portion 622 is connected to an input shaft of a transmission (not shown), and the output shaft 62 is rotatable together with the input shaft. That is, 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 by the transmission and output as drive torque to the drive wheels of the vehicle. This allows the vehicle to run.

Next, the 3k-type paradox planetary gear reducer employed by the reducer 30 of the present embodiment will be described.

In the electric clutch device of the present embodiment, it is required to shorten the time required for the initial response to close the initial gap (corresponding to the gap Sp1) between the clutch and the actuator. In order to speed up the initial response, it can be seen from the rotational motion equation that the moment of inertia around the input axis should be reduced. When the input shaft is a solid cylindrical member, the moment of inertia increases in proportion to the fourth power of the outer diameter when the length and density are constant. In the clutch device 1 of this embodiment, the sun gear 31 corresponding to the "input shaft" here is a hollow cylindrical member, but this tendency does not change.

The upper part of FIG. 3 shows a schematic diagram of a 2-kh type paradox planetary gear reducer. The upper part of FIG. 4 shows a schematic diagram of a 3k type paradox planetary gear reducer. Here, A is the sun gear, B is the planetary gear, C is the first ring gear, D is the second ring gear, and S is the carrier. Comparing the 2kh type and the 3k type, the 3k type has a configuration in which a sun gear A is added to the 2kh type.

In the case of the 2kh type, the moment of inertia around the input shaft is the smallest when the carrier S, which is positioned most radially inward among the constituent elements, is used as the input element (see the table at the bottom of FIG. 3).

On the other hand, in the case of the 3k type, the moment of inertia around the input shaft is the smallest when the sun gear A, which is the radially innermost component, is used as the input element (see the lower table in FIG. 4). ).

The magnitude of the moment of inertia is greater in the case of the 2kh type with the carrier S as the input element than in the case of the 3k type with the sun gear A as the input element. Therefore, when a paradox planetary gear speed reducer is used as a speed reducer in an electric clutch device that requires a quick initial response, it is desirable to use the 3k type and the sun gear A as an input element.

In addition, an electric clutch device requires a very large load of several thousand to ten and several thousand N, and in order to achieve both a high response and a high load, it is necessary to increase the speed reduction ratio of the speed reducer. Comparing the maximum reduction ratios of the 2kh type and the 3k type with the same gear specifications, the maximum reduction ratio of the 3k type is about twice that of the 2kh type, which is large. Also, in the 3k type, a large reduction ratio can be obtained when the sun gear A, which has the smallest moment of inertia, is used as the input element (see the table at the bottom of FIG. 4). Therefore, it can be said that the optimum configuration for achieving both high response and high load is a 3k type configuration using the sun gear A as an input element.

In this embodiment, the speed reducer 30 is a 3k paradox planetary gear speed reducer having a sun gear 31 (A) as an input element, a second ring gear 35 (D) as an output element, and a first ring gear 34 (C) as a fixed element. is. Therefore, the moment of inertia around the sun gear 31 can be reduced, and the speed reduction ratio of the speed reducer 30 can be increased. Therefore, both high response and high load can be achieved in the clutch device 1 .

In the case of the 2 kh type, the carrier S directly contributes to power transmission. A large bending moment may act between it and the main body (see the schematic diagram at the top of FIG. 3).

On the other hand, in the case of the 3k type, the carrier S has only the function of holding the planetary gear B in an appropriate position with respect to the sun gear A, the first ring gear C, and the second ring gear D. The bending moment acting between the pin) and the main body of the carrier S is small (see the upper schematic diagram of FIG. 4).

Therefore, in the present embodiment, the speed reducer 30 is a high-response, high-load paradoxical planetary gear speed reducer of 3k type. , the planetary gear 32 can be supported from one side in the axial direction, that is, cantilevered.

Next, the effect of having the disc spring 81 as the elastically deformable portion of the state changing portion 80 will be described.

As shown in FIG. 5, regarding the axial movement of the driven cam 50, that is, the relationship between the stroke and the load acting on the clutch 70, the configuration in which the clutch 70 is pushed by a rigid body that is difficult to elastically deform in the axial direction (one point in FIG. (see the chain line) and the structure of this embodiment that pushes the clutch 70 with the disk spring 81 that can be elastically deformed in the axial direction (see the solid line in FIG. 5). It can be seen that the variation in the load acting on the clutch 70 is smaller in the structure that pushes the clutch 70 than in the structure that pushes the clutch 70 with a rigid body. Compared to a configuration in which a rigid body presses the clutch 70, the use of the disk spring 81 can reduce the combined spring constant, thereby reducing variations in the load caused by variations in the stroke of the driven cam 50 caused by the actuator. It is from. In this embodiment, since the state changing portion 80 has the disc spring 81 as an elastically deforming portion, it is possible to reduce variations in load due to variations in the stroke of the driven cam 50, and to easily apply a target load to the clutch 70. .

Next, the configuration and the like of the bearing portion 151 will be described in detail.

As shown in FIG. 6, the bearing portion 151 includes a plurality of bearing rolling elements 173 that roll in the circumferential direction of the rotor 23 and rotatably support the rotor 23, and a lubricant 174 that lubricates the periphery of the bearing rolling elements 173. have The bearing portion 151 rotatably supports the rotor 23 via the sun gear 31 . Here, only one bearing portion 151 that rotatably supports the rotor 23 is provided.

More specifically, the bearing portion 151 has an inner ring 171, an outer ring 172, bearing rolling elements 173, a lubricant 174, a retainer 177, and the like.

The inner ring 171 is made of metal, for example, and has a substantially cylindrical shape. The outer ring 172 is made of metal, for example, and has a substantially cylindrical shape. The inner diameter of outer ring 172 is larger than the outer diameter of inner ring 171 . An inner peripheral wall of the inner ring 171 is fitted to an outer peripheral wall of the housing small inner cylindrical portion 126 . The outer ring 172 has an outer peripheral wall fitted to one end of the sun gear body 310 , that is, the inner peripheral wall of the end opposite to the sun gear tooth portion 311 .

An annular inner ring groove portion 175 that is recessed radially inward is formed in the outer peripheral wall of the inner ring 171 . The inner peripheral wall of the outer ring 172 is formed with an annular outer ring groove portion 176 that is recessed radially outward.

The bearing rolling elements 173 are spherically formed of metal, for example. A plurality of bearing rolling elements 173 are provided so as to be able to roll between the inner ring groove portion 175 of the inner ring 171 and the outer ring groove portion 176 of the outer ring 172 .

<3> The retainer 177 is formed in an annular or cylindrical shape. A retainer 177 is provided between the inner ring 171 and the outer ring 172 . A plurality of holding holes 178 are formed in the retainer 177 . For example, 24 holding holes 178 are formed at equal intervals in the circumferential direction of the retainer 177 .

The bearing rolling elements 173 are provided so as to be held in the holding holes 178 . A total of eight bearing rolling elements 173 are provided at equal intervals in the circumferential direction of the retainer 177, for example. That is, the bearing rolling elements 173 are provided in the holding holes 178 arranged in the circumferential direction of the retainer 177 so as to skip two. The holding hole portion 178 can hold the bearing rolling element 173 so that the bearing rolling element 173 can roll between the inner ring 171 and the outer ring 172 .

In this embodiment, the number of bearing rolling elements 173 (8) is less than the number of holding holes 178 (24).

Lubricant 174 is fluid such as grease. Lubricant 174 is provided around bearing rolling elements 173 , inner ring groove 175 , outer ring groove 176 , and holding hole 178 of retainer 177 to lubricate the surroundings of bearing rolling elements 173 . This allows the bearing rolling elements 173 to roll smoothly between the inner ring 171 and the outer ring 172 in the holding hole portion 178 .

The dynamic viscosity of the lubricant 174 changes depending on the environmental temperature. The dynamic viscosity of the lubricant 174 increases, for example, as the environmental temperature decreases.

As shown in FIG. 7A, when the bearing rolling elements 173 roll between the inner ring 171 and the outer ring 172, the bearing rolling elements 173 roll while repelling the lubricant 174 around them. Therefore, a resistance f that repels the lubricant 174 acts on the rolling bearing rolling element 173 . Here, the total resistance in the bearing portion 151 is a value obtained by multiplying f by the number of bearing rolling elements 173 .

On the other hand, as shown in FIG. 7B, when the bearing rolling element 173 is removed from the holding hole portion 178, the resistance f does not occur in the holding hole portion 178 from which the bearing rolling element 173 is removed. Therefore, by reducing the number of bearing rolling elements 173 provided in the bearing portion 151, the total value of the resistance in the bearing portion 151 can be reduced. Thereby, the starting torque of the bearing portion 151 can be reduced.

Next, the starting torque of the bearing portion 151 of this embodiment and the comparative example will be described.

In the bearing portion 151 of the comparative example, the bearing rolling elements 173 are provided in all of the 24 holding holes 178 formed in the retainer 177 . That is, in the comparative form, the bearing portion 151 has 24 bearing rolling elements 173 . As for the number of bearing rolling elements provided in a similar bearing portion, 24 is a general (standard) number.

As shown in FIG. 8, the starting torque (see the solid line in FIG. 8) of the bearing portion 151 of the present embodiment having eight bearing rolling elements 173 is 24% for the bearing rolling elements 173 regardless of the ambient temperature (environmental temperature). It is smaller than the starting torque (see one-dot chain line in FIG. 8) of the bearing portion 151 of the comparative embodiment. That is, by reducing the number of bearing rolling elements 173, the starting torque of the bearing portion 151 can be reduced.

Especially in the cryogenic region, by changing the number of bearing rolling elements 173 from 24 (comparative embodiment) to 8 (this embodiment), the starting torque of the bearing portion 151 can be reduced by 15.9 (mNm). See Figure 8).

Moreover, the starting torque of the bearing portion 151 of the present embodiment having eight bearing rolling elements 173 satisfies the required value regardless of the ambient temperature (see FIG. 8).

A solid line in FIG. 9 indicates the relationship between the number of bearing rolling elements 173 in the bearing portion 151 and the starting torque of the bearing portion 151 (vertical axis on the left side of FIG. 9). 9 indicates the relationship between the number of bearing rolling elements 173 in the bearing portion 151 and the withstand load of the bearing portion 151 (vertical axis on the right side of FIG. 9).

As shown in FIG. 9, S is the maximum value of the load (stress) applied to the bearing portion 151 . Also, if the number of bearing rolling elements 173 is less than the assembly limit (a range that satisfies the assembly conditions of bearing portion 151 ), bearing rolling elements 173 may drop out from between inner ring 171 and outer ring 172 .

<2> In the present embodiment, the number of bearing rolling elements 173 is as small as possible within a range that withstands the load (stress) applied to the bearing section 151 and within a range that satisfies the assembly conditions of the bearing section 151. It is set to a certain eight (see FIG. 9). As a result, the starting torque of the bearing portion 151 can be reduced as compared with the comparative embodiment in which the number of the bearing rolling elements 173 is 24 while suppressing the deterioration of the durability of the bearing portion 151 .

The inner peripheral wall of the bearing portion 151 , that is, the inner peripheral wall of the inner ring 171 is fitted to the outer peripheral wall of the housing inner tubular portion 121 . The sun gear 31 is provided so that the inner peripheral wall of the sun gear body 310 fits into the outer peripheral wall of the bearing portion 151 , that is, the outer peripheral wall of the outer ring 172 . Thereby, the rotor 23 is rotatably supported by the housing inner cylindrical portion 121 via the sun gear 31 and the bearing portion 151 . That is, the bearing portion 151 rotatably supports the rotor 23 .

Thus, in this embodiment, there is only one bearing portion 151 that rotatably supports the rotor 23 .

The speed reducer 30 has a sun gear 31 as an “input portion” that is rotatable and coaxial with the rotor 23 and receives torque from the rotor 23 . Thus, in this embodiment, the speed reducer 30 is a non-eccentric planetary speed reducer that does not have an eccentric portion that is eccentric with respect to the rotor 23 .

Next, disengagement of the clutch 70 at the time of power failure will be described.

In the clutch device 1 having the electric motor 20 as a drive source as in the present embodiment, the load transmitted to the clutch 70 is reduced when the power source fails due to disconnection of the power line (coil 22, etc.) of the motor 20. It may be required to pull out quickly, that is, to release the clutch 70 .

As shown in FIG. 10 , a return spring load (Fs), which is the load of the return spring 55 , acts on the driven cam 50 of the ball cam 2 . The return spring load is the same as the ball cam load (Fc) that acts on the ball cam 2 when the amount of movement of the driven cam 50 toward the clutch 70 side, that is, the stroke amount is zero.

When the return spring load (ball cam load) acts on the driven cam 50 , drive cam driven torque, which is the torque that rotates the drive cam 40 , is generated. When the drive cam driven torque is generated, the rotor driven torque, which is the torque that rotates the rotor 23 , is generated via the speed reducer 30 . Here, rotor detent torque (Td), which is torque opposing the rotor driven torque, is generated in the rotor 23 . The rotor detent torque consists of the cogging torque of the motor 20 and the bearing loss torque that is the starting torque of the bearing portion 151 .

FIG. 11 is a diagram showing the relationship between the stroke amount of the driven cam 50 toward the clutch 70, the ball cam load (solid line), the return spring load (chain line), and the clutch load (broken line) acting on the clutch 70. is. As shown in FIG. 11, when the stroke amount of the driven cam 50 toward the clutch 70 is 0, the lower limit of the return spring load and the lower limit of the ball cam load match. Here, the disengagement condition of the clutch 70 is that the rotor driven torque (lower limit) due to the drive cam driven torque due to the return spring load when the stroke amount of the driven cam 50 is 0 is greater than the rotor detent torque (upper limit). ”.

In FIG. 12, T1 is the drive cam driven torque. T2 is loss torque, which is torque lost in the inner seal member 401, the outer seal member 402, and the thrust bearing 161 when the drive cam 40 rotates. T3 is the torque obtained by subtracting T2 from T1. T4 is a torque obtained by dividing T3 by the speed reduction ratio of the speed reducer 30 . T6 is a torque obtained by multiplying T4 by the inverse efficiency of the speed reducer 30, and is a rotor driven torque. T7 is the torque obtained by dividing the cogging torque of the motor 20 by two. T8 is the bearing loss torque of the bearing portion 151; Rotor detent torque is the sum of T7 and T8. T9 is a margin torque that is the difference between the rotor driven torque (T6) and the rotor detent torque (T7+T8). In this embodiment, the marginal torque (T9) is set to a value that allows the clutch 70 to be released within a predetermined period of time in the event of a power failure.

A basic formula for the disengagement condition of the clutch 70 is represented by Formula 1 below.
Rotor driven torque (T6) lower limit - rotor detent torque (T7+T8) upper limit>0 Equation 1

The upper limit of the rotor detent torque (T7+T8) is expressed by Equation 2 below.
Rotor detent torque (T7+T8) upper limit = cogging torque upper limit/2 + bearing loss torque (T8) upper limit Equation 2

The lower limit of the rotor driven torque (T6) is expressed by Equation 3 below.
Rotor driven torque (T6) lower limit = {drive cam driven torque (T1) - loss torque (T2)}/reduction ratio x reverse efficiency lower limit (Equation 3)

The drive cam driven torque (T1) is represented by the following equation 4.
Drive cam driven torque (T1) = lower limit of ball cam load/upper limit of ball cam conversion ratio x lower limit of reverse efficiency of ball cam Formula 4
The ball-cam conversion ratio in Equation 4 is the load that can be output when 1 Nm is applied to the drive cam 40 without friction. The ball cam reverse efficiency is the reverse efficiency of the ball cam 2 .

As described above, in order to drive the rotor 23 only with the return spring load, it is essential to reduce the starting torque (bearing loss torque) of the bearing portion 151 .

Next, an operation example of disengagement of the clutch 70 at the time of power failure will be described.

In the clutch device 1 having the electric motor 20 as a drive source as in the present embodiment, motor torque is generated by energizing the motor 20, increased by the speed reducer 30, input to the ball cam 2, and converted into a load. output as Since power loss occurs due to friction in the speed reducer 30 and the ball cam 2, the motor is balanced with the same load when the clutch 70 is pushed in (forward action) and when it is pushed back from the clutch 70 (reverse action). A difference occurs in torque (ACT load hiss characteristics: see FIG. 13).

The ACT conversion ratio (when there is no friction), which is the conversion ratio of the actuator of the clutch device 1, is expressed by the following equation 5.
ACT conversion ratio (no friction) = speed reduction ratio × conversion ratio Equation 5

The ACT conversion ratio at the time of direct action, which is the ACT conversion ratio at the time of positive action (in the case of no friction), is expressed by Equation 6 below.
ACT conversion ratio at positive operation (no friction) = ACT conversion ratio (no friction) x speed reducer positive efficiency x ball cam positive efficiency Equation 6
The reduction gear positive efficiency in Equation 6 is the positive efficiency of the reduction gear 30 . The ball cam positive efficiency is the positive efficiency of the ball cam 2 .

The ACT conversion ratio at the time of reverse action (in the case of no friction), which is the ACT conversion ratio at the time of reverse action, is expressed by Equation 7 below.
ACT conversion ratio during reverse operation (no friction) = ACT conversion ratio (no friction) x speed reducer reverse efficiency x ball cam reverse efficiency Equation 7
The inverse efficiency of the speed reducer in Equation 7 is the inverse efficiency of the speed reducer 30 . The ball cam reverse efficiency is the reverse efficiency of the ball cam 2 .

Note that the slope of the graph of the ACT load hiss characteristics (see FIG. 13) corresponds to the ACT conversion ratio.

Suppose that the reduction ratio is 60, the reduction gear positive efficiency is 80%, the reduction gear reverse efficiency is 80%, the ball cam conversion ratio, which is the load that can be output when 1 Nm is applied to the drive cam 40 without friction, is 300 N / Nm, the ball cam positive Assuming that the efficiency is 90% and the ball cam reverse efficiency is 90%, the load during forward operation is expressed by Equation 8 below.
During direct operation: Load = motor torque MT1 x reduction ratio x reduction gear positive efficiency x conversion ratio x ball cam positive efficiency Equation 8

On the other hand, the load at the time of reverse operation is represented by Equation 9 below.
During reverse operation: load = motor torque MT2 × reduction ratio / reduction gear reverse efficiency × conversion ratio / ball cam reverse efficiency Equation 9

When formulas 8 and 9 are simultaneously applied to the load, formula 10 below is obtained.
Motor torque MT2/motor torque MT1=reducer positive efficiency×reducer reverse efficiency×ball cam positive efficiency×ball cam reverse efficiency≈0.52 Equation 10
As shown in Equation 10, the motor torque that balances the same load during reverse operation is approximately half that of normal operation.

In order to release the clutch 70 when the power fails, it is necessary to reversely drive the rotor 23 only with the return spring load. The return spring load is set significantly smaller than the engagement load of the clutch 70 . Therefore, the rotor driven torque by the return spring 55 is similarly small.

For example, if the motor 20 can output a maximum of 1 Nm in the above specifications, the maximum load that can be output is 1 x 60 x 0.8 x 300 x 0.9 = 13000 (N). is 10000N, the remaining 3000N can be set as the return spring load. However, it is desirable to set the return spring load to about 1000 to 2000N because it is actually necessary to give some margin to the motor torque.

The driven torque of the rotor 23 generated by the load of 1000 N is 1000/300×0.9/60×0.8=40 (mNm), which is the detent torque of the rotor 23 (cogging torque/2+bearing loss torque). If it wins, it becomes possible to open the clutch 70 only with the load of the return spring.

At low temperatures, the kinematic viscosity of the lubricant 174 increases, and the shear resistance of the lubricant 174 generated between the bearing rolling elements 173 and the retainer 177, inner ring 171, and outer ring 172 increases. Therefore, out of the detent torque, the bearing loss torque tends to increase in the low temperature range. It can be done, but it can't be released at low temperatures." In order to avoid this, in this embodiment, the number of bearing rolling elements 173 of the bearing portion 151 is reduced from 24 in the comparative embodiment to 8, and the bearing loss torque is kept low even at low temperatures. can be opened.

Further, in the present embodiment, by reducing the number of bearing rolling elements 173 of the bearing portion 151, the moment of inertia of the rotor 23 is reduced, thereby improving startability and responsiveness. Responsiveness, especially when closing the clearance (initial gap: Sp1) between the state changing unit 80 and the clutch 70, depends on the speed of rise of the rotation speed of the motor 20 (see FIG. 14), so the responsiveness is improved. Regarding , reducing the moment of inertia by reducing the number of bearing rolling elements 173 is effective.

Further, as shown in FIG. 15, by reducing the number of bearing rolling elements 173 of the bearing portion 151 from 24 in the comparative embodiment to 8, the rotational torque of the motor 20 is can be reduced. Therefore, the responsiveness of the motor 20 can be improved.

<4> The bearing portion 151 is a ball bearing, that is, a “ball bearing”. More specifically, the bearing portion 151 is a “single-row ball bearing” in which the bearing rolling elements 173 are arranged in one row in the axial direction of the inner ring 171 and the outer ring 172 (see FIG. 16).

<5> In the axial direction of the bearing portion 151, the bearing portion 151 is provided apart from the sun gear tooth portion 311 as the “input portion” (see FIG. 16).

More specifically, in the axial direction of bearing portion 151, a "sun gear load application position" is a position at the center of bearing portion 151 and a position at the center of sun gear tooth portion 311 where a load acts on sun gear 31. are separated by a distance d1 (see FIG. 16).

Next, the operation and the like of the speed reducer 30 and the bearing portion 151 configured as described above will be described.

As shown in the upper part of FIG. 17 , four planetary gears 32 are provided at equal intervals in the circumferential direction of the sun gear 31 on the radially outer side of the sun gear 31 . For the sake of explanation, the four planetary gears 32 are referred to as planetary gears Gp1, Gp2, Gp3, and Gp4 in the counterclockwise direction.

In an ideal gear shape, the torque sharing rate of each planetary gear 32 (Gp1 to Gp4) is constant. Therefore, the tooth surface acting forces acting on the sun gear 31 from the planetary gears 32 (Gp1 to Gp4) cancel each other out, and the resultant force becomes 0 (see the lower part of FIG. 17).

FIG. 18 shows an example in which the torque sharing ratios of the planetary gears 32 (Gp1 to Gp4) are uneven. Note that when there are four planetary gears 32 as in this embodiment, the average torque sharing rate is 25%. As shown in the upper part of FIG. 18, when only the planetary gear Gp1 has a torque sharing ratio higher than 25% and the torque sharing ratios of the planetary gears Gp2 to Gp4 are a constant value lower than 25%, the tooth surface acting force acting on the sun gear 31 is reduced. The resultant force does not become 0 (see the lower part of FIG. 18).

Therefore, in this embodiment, the carrier 33 is configured so that the inner peripheral wall does not come into contact with the outer peripheral wall of the sun gear 31 or the like, that is, it is of a floating type. As a result, theoretically, the torque distribution ratio of each planetary gear 32 (Gp1 to Gp4) can be brought close to constant.

Therefore, the resultant force of the tooth surface acting forces acting on the sun gear 31, that is, the sun gear resultant force becomes small. As a result, in the axial direction of the bearing portion 151, even if the center position of the bearing portion 151 and the sun gear load application position are separated by a distance d1 (see FIG. 16), the product of the arm length (d1) and the sun gear resultant force A certain bending moment becomes a local minimum. That is, in the speed reducer 30, which is a non-eccentric planetary speed reducer that does not have an eccentric portion as in the present embodiment, the tooth surface load generated at the torque transmission portion (between the sun gear tooth portion 311 and the planetary gear tooth portion 321) is zero or very small in the radial direction.

<7> The motor 20 has a magnet 230 as a “permanent magnet” provided on the rotor 23 (see FIG. 16). Here, the magnet 230 is provided on the outer peripheral wall of the rotor 23 . That is, the motor 20 is a surface magnet type (SPM) motor.

The configuration of each part of the present embodiment will be described in more detail below.

<13> In the present embodiment, the clutch device 1 includes an oil supply portion 5 (see FIGS. 1 and 2). The oil supply portion 5 is formed in the shape of a passage on the output shaft 62 so that one end thereof is exposed to the clutch space 620 . The other end of the oil supply portion 5 is connected to an oil supply source (not shown). As a result, oil is supplied from one end of the oil supply portion 5 to the clutch 70 in the clutch space 620 .

The ECU 10 controls the amount of oil supplied from the oil supply section 5 to the clutch 70 . The oil supplied to clutch 70 can lubricate and cool clutch 70 . Thus, in this embodiment, the clutch 70 is a wet clutch and can be cooled by oil.

In this embodiment, the ball cam 2 as the "rotational translation section" forms an accommodation space 120 between the housing 12 and the drive cam 40 and the second ring gear 35 as the "rotational section." Here, the accommodation space 120 is formed inside the housing 12 on the side opposite to the clutch 70 with respect to the drive cam 40 and the second ring gear 35 . Motor 20 and speed reducer 30 are provided in accommodation space 120 . The clutch 70 is provided in a clutch space 620 which is a space on the opposite side of the housing space 120 with respect to the drive cam 40 .

In this embodiment, the clutch device 1 has a thrust bearing 161 and a thrust bearing washer 162 . The thrust bearing washer 162 is formed of metal, for example, in a substantially annular plate shape, and is provided so that one surface thereof abuts on the housing stepped surface 125 . The thrust bearing 161 is provided between the other surface of the thrust bearing washer 162 and the surface of the drive cam main body 41 opposite to the driven cam 50 . The thrust bearing 161 bears the drive cam 40 while receiving a load in the thrust direction from the drive cam 40 . In this embodiment, a thrust-direction load acting on the driving cam 40 from the clutch 70 side via the driven cam 50 acts on the housing step surface 125 via the thrust bearing 161 and the thrust bearing washer 162 . Therefore, the drive cam 40 can be stably supported by the housing step surface 125 .

In this embodiment, the clutch device 1 includes an inner seal member 401 and an outer seal member 402 as "seal members".

<14> The inner seal member 401 and the outer seal member 402 are oil seals that are annularly formed of an elastic material such as rubber and a metal ring.

The inner and outer diameters of inner seal member 401 are smaller than the inner and outer diameters of outer seal member 402 .

The inner seal member 401 is positioned radially between the housing inner cylindrical portion 121 and the thrust bearing 161 and axially between the thrust bearing washer 162 and the drive cam body 41 . . The inner seal member 401 is fixed to the housing inner tubular portion 121 and is rotatable relative to the drive cam 40 .

The outer seal member 402 is provided between the gear inner tubular portion 355 of the second ring gear 35 and the end portion of the housing outer tubular portion 123 on the clutch 70 side. The outer seal member 402 is fixed to the housing outer cylindrical portion 123 and is rotatable relative to the second ring gear 35 .

Here, the outer seal member 402 is provided so as to be positioned radially outward of the inner seal member 401 when viewed from the axial direction of the inner seal member 401 (see FIGS. 1 and 2).

The thrust bearing washer 162 side surface of the drive cam body 41 is slidable on the seal lip portion of the inner seal member 401 . That is, the inner seal member 401 is provided so as to come into contact with the drive cam 40 as a "rotating portion". The inner seal member 401 hermetically or liquid-tightly seals between the drive cam body 41 and the thrust bearing washer 162 .

The outer peripheral wall of the gear inner cylindrical portion 355 of the second ring gear 35 is slidable on the seal lip portion, which is the inner edge portion of the outer seal member 402 . That is, the outer seal member 402 is provided so as to come into contact with the second ring gear 35 that rotates integrally with the drive cam 40 on the radially outer side of the drive cam 40 as a "rotating portion." The outer seal member 402 hermetically or liquid-tightly seals the outer peripheral wall of the gear inner tubular portion 355 and the inner peripheral wall of the housing outer tubular portion 123 .

By the inner seal member 401 and the outer seal member 402 provided as described above, the accommodation space 120 that accommodates the motor 20 and the speed reducer 30 and the clutch space 620 in which the clutch 70 is provided are airtight or liquid sealed. It can be kept dense. As a result, for example, even if foreign matter such as abrasion powder is generated in the clutch 70 , it is possible to prevent the foreign matter from entering the housing space 120 from the clutch space 620 . Therefore, malfunction of the motor 20 or the speed reducer 30 due to foreign matter can be suppressed.

In this embodiment, since the space between the accommodation space 120 and the clutch space 620 is kept airtight or liquid-tight by the inner seal member 401 and the outer seal member 402 , wear powder and the like are contained in the oil supplied to the clutch 70 . , the oil containing the foreign matter can be prevented from flowing from the clutch space 620 into the accommodation space 120 .

In this embodiment, the housing 12 is formed to have a closed shape from a portion corresponding to the radially outer side of the outer sealing member 402 to a portion corresponding to the radially inner side of the inner sealing member 401 (see FIGS. 1 and 2). reference).

In this embodiment, the drive cam 40 and the second ring gear 35 that form the accommodation space 120 with the housing 12 rotate relative to the housing 12 but do not move relative to the housing 12 in the axial direction. Therefore, when the clutch device 1 is actuated, a change in the volume of the accommodation space 120 can be suppressed, and generation of negative pressure in the accommodation space 120 can be suppressed. As a result, it is possible to prevent oil or the like containing foreign matter from being sucked into the housing space 120 from the clutch space 620 side.

In addition, the inner seal member 401 that contacts the inner edge of the drive cam 40 slides on the drive cam 40 in the circumferential direction, but does not slide in the axial direction. Also, the outer seal member 402 that contacts the outer peripheral wall of the gear inner cylindrical portion 355 of the second ring gear 35 slides on the second ring gear 35 in the circumferential direction, but does not slide in the axial direction.

<12> As shown in FIG. 1 , the drive cam main body 41 is located on the side opposite to the clutch 70 with respect to the drive cam outer cylindrical portion 44 . That is, the drive cam 40 as a “rotating portion” is bent in the axial direction to form a drive cam main body 41 that is the inner edge of the drive cam 40 and a drive cam outer cylindrical portion 44 that is the outer edge of the drive cam 40 . are formed at different positions in the axial direction.

The driven cam main body 51 is provided so as to be positioned radially inside the drive cam inner cylindrical portion 42 on the clutch 70 side of the drive cam main body 41 . That is, the driving cam 40 and the driven cam 50 are provided in a nested manner in the axial direction.

More specifically, the driven cam main body 51 is positioned radially inside the gear plate portion 356 of the second ring gear 35 , the gear outer cylinder portion 357 , the drive cam plate portion 43 and the drive cam inner cylinder portion 42 . Further, the sun gear tooth portion 311 of the sun gear 31 , the carrier 33 and the planetary gear 32 are positioned radially outside the drive cam body 41 and the driven cam body 51 . As a result, the axial size of the clutch device 1 including the speed reducer 30 and the ball cam 2 can be significantly reduced.

In this embodiment, as shown in FIG. 1, the drive cam body 41, the sun gear 31, the carrier 33, and the coil 22 are arranged so as to partially overlap in the axial direction of the drive cam body 41. As shown in FIG. In other words, a portion of the coil 22 is positioned radially outward of a portion of the drive cam body 41 , the sun gear 31 and the carrier 33 in the axial direction. As a result, the size of the clutch device 1 in the axial direction can be further reduced.

Further, as shown in FIG. 1, in the present embodiment, the bearing portion 151 is provided radially inward of the sun gear tooth portion 311 as the "input portion" when viewed from the axial direction of the bearing portion 151 . More specifically, when viewed from the axial direction of the bearing portion 151, the outer edge portion (the outer peripheral wall of the outer ring 172) of the bearing portion 151 is positioned radially inward with respect to the outer edge portion (tooth tip) of the sun gear tooth portion 311. It is set up like this. Therefore, the radial size of the clutch device 1 can be reduced while securing the radial sizes of the stator 21 and the rotor 23 .

As described above, <1> in the present embodiment, the bearing portion 151 includes a plurality of bearing rolling elements 173 that roll in the circumferential direction of the rotor 23 and rotatably support the rotor 23 , and the bearing rolling elements 173 . It has a lubricant 174 that lubricates the surroundings. Here, only one bearing portion 151 that rotatably supports the rotor 23 is provided. The speed reducer 30 has a sun gear tooth portion 311 as an “input portion” which is provided coaxially and rotatably with the rotor 23 and receives torque from the rotor 23 .

In this embodiment, when torque is input from the motor 20 to the sun gear tooth portion 311 , the sun gear tooth portion 311 rotates coaxially with the rotor 23 . Therefore, the radial load acting on the sun gear tooth portion 311 from the gear or the like provided radially outside the sun gear tooth portion 311 can be reduced. Therefore, the number of bearing portions 151 that rotatably support the rotor 23 can be reduced to one.

Moreover, since the radial load acting on the sun gear tooth portion 311 can be reduced, even if the number of the bearing rolling elements 173 of the bearing portion 151 is reduced, deterioration in durability can be suppressed. Therefore, the starting torque and rotation torque of the bearing portion 151 can be reduced. As a result, the responsiveness is improved particularly at low temperatures, and the minimum set load required to remove the load on the clutch 70 in the event of a power failure can be reduced.

Also, by reducing the number of bearing rolling elements 173 of the bearing portion 151, the moment of inertia of the rotor 23 can be reduced, and the responsiveness can be further improved.

<2> In the present embodiment, the number of bearing rolling elements 173 is set to be as small as possible within a range that can withstand the load applied to the bearing portion 151 and within a range that satisfies the assembly conditions of the bearing portion 151. It is

Therefore, the starting torque of the bearing portion 151 can be reduced while suppressing deterioration of the durability of the bearing portion 151 . Thereby, responsiveness can be further improved while ensuring durability.

Moreover, <3> in the present embodiment, the bearing portion 151 has a retainer 177 formed with a plurality of retaining holes 178 capable of retaining the bearing rolling elements 173 . The number of bearing rolling elements 173 is less than the number of holding holes 178 .

Thus, by setting the number of bearing rolling elements 173 to be smaller than the number of holding holes 178 formed in cage 177, the starting torque of bearing 151 can be easily reduced, and responsiveness can be improved. .

Moreover, in <4> this embodiment, the bearing part 151 is a ball bearing.

Therefore, the durability and bearing accuracy of the bearing portion 151 can be improved. Moreover, the bearing portion 151 is a single row ball bearing. Therefore, the size of the bearing portion 151 in the axial direction can be reduced.

Further, <5> in the present embodiment, the bearing portion 151 is provided apart from the sun gear tooth portion 311 as the “input portion” in the axial direction of the bearing portion 151 .

Therefore, a part of the speed reducer 30 and a part of the ball cam 2 can be nested to secure a large space.

In the speed reducer 30, which is a non-eccentric planetary speed reducer with no eccentric portion, the tooth surface load generated in the torque transmission portion (between the sun gear tooth portion 311 and the planetary gear tooth portion 321) is zero or zero in the radial direction. extremely small. Therefore, even if the bearing portion 151 and the sun gear tooth portion 311 are separated from each other in the axial direction of the bearing portion 151, no bending moment is generated, and the impact on the durability of the bearing portion 151 is small. Moreover, a large load in the radial direction does not act on the sun gear tooth portion 311 , and the rotor 23 can be appropriately rotatably supported by the bearing portion 151 .

<7> In this embodiment, the motor 20 has a magnet 230 provided on the rotor 23 . That is, the motor 20 is a brushless DC motor using the magnet 230 as a "permanent magnet."

In addition, in this embodiment, the inner seal member 401 and the outer seal member 402 can keep the space between the accommodation space 120 and the clutch space 620 liquid-tight. As a result, even if the oil supplied to the clutch 70 for cooling the clutch 70 contains magnetic particles such as iron powder, the oil containing the magnetic particles is prevented from flowing from the clutch space 620 into the accommodation space 120. can. Therefore, it is possible to suppress magnetic particles from being attracted to the magnet 230 of the motor 20, thereby suppressing deterioration of rotation performance of the motor 20 and malfunction.

<8> In this embodiment, the speed reducer 30 has a sun gear 31 , a planetary gear 32 , a carrier 33 , a first ring gear 34 and a second ring gear 35 . The torque of the motor 20 is input to the sun gear tooth portion 311 as an "input portion". The planetary gear 32 can revolve in the circumferential direction of the sun gear tooth portion 311 (sun gear 31) while rotating while being meshed with the sun gear tooth portion 311 (sun gear 31).

The carrier 33 rotatably supports the planetary gear 32 and is rotatable relative to the sun gear teeth 311 (sun gear 31). The first ring gear 34 can mesh with the planetary gears 32 . The second ring gear 35 is meshable with the planetary gear 32 and has a different number of teeth than the first ring gear 34 , and outputs torque to the drive cam 40 of the ball cam 2 .

In this embodiment, the speed reducer 30 corresponds to the structure with the highest response and highest load among the many structures and input/output patterns of the paradox planetary gear speed reducer. Therefore, both high response and high load of the speed reducer 30 can be achieved.

Further, in the present embodiment, as described above, the inner seal member 401 and the outer seal member 402 can keep the space between the accommodation space 120 and the clutch space 620 liquid-tight. As a result, it is possible to suppress the effects of oil containing fine iron powder, such as damage, wear, and in principle a decrease in efficiency, on the speed reducer 30 as a "paradox planetary gear speed reducer" having a large number of meshing portions.

<9> In the present embodiment, the first ring gear 34 is fixed to the housing 12 . The second ring gear 35 is rotatably provided integrally with the drive cam 40 .

In this embodiment, the responsiveness of the clutch device 1 can be improved by connecting each part as described above so that the moment of inertia of the high-speed rotation portion of the speed reducer 30 as the "paradox planetary gear speed reducer" is small.

<11> In the present embodiment, the “rotating portion” of the “rotational translation portion” is the drive cam 40 having a plurality of drive cam grooves 400 formed on one surface in the axial direction. The “translation portion” is the driven cam 50 having a plurality of driven cam grooves 500 formed on one axial surface. The “rotational translation part” is the drive cam 40 , the driven cam 50 , and the ball cam 2 having the ball 3 provided to roll between the drive cam groove 400 and the driven cam groove 500 .

Therefore, the efficiency of the "rotational translation section" can be improved as compared with the case where the "rotational translational section" is composed of, for example, a "slide screw". Moreover, compared with the case where the "rotational translation part" is composed of, for example, a "ball screw", the cost can be reduced, the size of the "rotational translation part" in the axial direction can be reduced, and the clutch device can be made smaller.

<12> In the present embodiment, the drive cam 40 as the "rotating portion" is arranged such that the drive cam main body 41, which is the inner edge, and the drive cam outer cylindrical portion 44, which is the outer edge, are located at different positions in the axial direction. formed.

Therefore, the drive cam 40 and the driven cam 50 and the speed reducer 30 as the "translation part" can be arranged in a nested manner in the axial direction, and the size of the clutch device 1 in the axial direction can be reduced.

<13> In the present embodiment, the motor 20 and the speed reducer 30 are provided in the accommodation space 120 formed inside the housing 12 on the side opposite to the clutch 70 with respect to the drive cam 40 . The clutch 70 is provided in a clutch space 620 which is a space on the opposite side of the housing space 120 with respect to the drive cam 40 .

An inner seal member 401 and an outer seal member 402 as "seal members" are formed in an annular shape, and are in contact with the drive cam 40 as a "rotating part" or the second ring gear 35 that rotates integrally with the drive cam 40. The space between the housing space 120 and the clutch space 620 can be kept airtight or liquid-tight.

As a result, for example, even if foreign matter such as abrasion powder is generated in the clutch 70 , it is possible to prevent the foreign matter from entering the housing space 120 from the clutch space 620 . Therefore, malfunction of the motor 20 or the speed reducer 30 due to foreign matter can be suppressed. Therefore, malfunction of the clutch device 1 due to foreign matter can be suppressed.

In this embodiment, an inner seal member 401 as a "seal member" and an outer seal member 402 are arranged to contact the drive cam 40 as a "rotating part" or the second ring gear 35 that rotates integrally with the drive cam 40. The space between the housing space 120 and the clutch space 620 is kept airtight or liquid-tight. Therefore, it is possible to prevent oil or the like containing fine iron powder or the like from entering the accommodation space 120 that accommodates the motor 20 and the speed reducer 30, so that good performance of the clutch device 1 can be maintained for a long period of time.

Further, in this embodiment, the inner seal member 401 and the outer seal member 402 rotate integrally with the drive cam 40, which is a component after being decelerated by the speed reducer 30 and amplified to a large drive torque, or the drive cam 40. It is provided so as to contact the second ring gear 35 . Therefore, the percentage of torque loss associated with sealing by the "sealing member" to the total becomes small, which is advantageous in terms of efficiency. If the "seal member" is provided so as to come into contact with the rotor 23 or the like, which is a component on the input side of the speed reducer 30, the "seal member" deprives the loss torque against a small driving torque, so the efficiency is greatly improved. may decrease to

In this embodiment, the storage space 120 is formed on the upstream side of the drive cam 40 in the power flow path, and is sealed by the inner seal member 401 and the outer seal member 402 . Also, the inner seal member 401 and the outer seal member 402 do not move relative to the housing 12 in the axial direction. Therefore, even if the drive cam 40 rotates, the volume of the accommodation space 120 does not change. As a result, it is not affected by changes in the spatial volume due to the translational movement of the driven cam 50 as a "translational part", and a special sealing member such as a bellows-shaped sealing member described in, for example, US Patent Application Publication No. 2015/0144453 is used. No volume change absorbing means is required.

<14> In the present embodiment, the inner seal member 401 and the outer seal member 402 as "seal members" are oil seals.

Therefore, the contact area between the inner seal member 401 and the outer seal member 402 and the drive cam 40 or the second ring gear 35 can be reduced. As a result, the sliding resistance acting on the inner seal member 401 and the outer seal member 402 when the drive cam 40 rotates can be reduced. Therefore, it is possible to suppress a decrease in efficiency during operation of the clutch device 1 .

<15> In the present embodiment, the state changing portion 80 has a disc spring 81 as an “elastic deformation portion” capable of elastically deforming in the axial direction of the driven cam 50 as a “translation portion”.

By controlling the rotational angular position of the motor 20, the thrust force of the clutch 70 can be controlled with high precision based on the displacement and load characteristics of the disc spring 81. FIG. Therefore, variations in the load acting on the clutch 70 with respect to variations in the stroke of the driven cam 50 can be reduced. As a result, highly accurate load control becomes possible, and the clutch device 1 can be controlled with high accuracy.

(Second embodiment)
A clutch device according to the second embodiment is shown in FIG. The second embodiment differs from the first embodiment in the configuration of the clutch and the state changing unit, and the like.

In this embodiment, ball bearings 141 and 143 are provided between the inner peripheral wall of the fixed body 11 and the outer peripheral wall of the input shaft 61 . Thereby, the input shaft 61 is supported by the fixed body 11 via the ball bearings 141 and 143 .

The housing 12 is fixed to the fixed body 11 so that a part of the outer wall contacts the wall surface of the fixed body 11 . For example, the housing 12 is fixed so that the surface of the housing small plate portion 124 opposite to the ball 3 , the inner peripheral wall of the housing inner cylindrical portion 121 and the inner peripheral wall of the housing small inner cylindrical portion 126 are in contact with the outer wall of the fixed body 11 . It is fixed to body 11 . The housing 12 is fixed to the fixed body 11 by bolts (not shown) or the like. Here, housing 12 is provided coaxially with fixed body 11 and input shaft 61 .

The arrangement of the motor 20, speed reducer 30, ball cam 2, etc. with respect to the housing 12 is the same as in the first embodiment.

In this embodiment, the output shaft 62 has a shaft portion 621 , a plate portion 622 , a cylindrical portion 623 and a cover 625 . 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 in an annular plate shape from one end of the shaft portion 621 . The cylindrical portion 623 is formed integrally with the plate portion 622 so as to extend in a substantially cylindrical shape from the outer edge portion of the plate portion 622 to the side opposite to the shaft portion 621 . The output shaft 62 is supported by the input shaft 61 via ball bearings 142 . A clutch space 620 is formed inside the cylindrical portion 623 .

Clutch 70 is provided between input shaft 61 and output shaft 62 in clutch space 620 . The clutch 70 has a support portion 73 , a friction plate 74 , a friction plate 75 and a pressure plate 76 . The support portion 73 is formed in a substantially annular plate shape so as to extend radially outward from the outer peripheral wall of the end portion of the input shaft 61 on the driven cam 50 side with respect to the plate portion 622 of the output shaft 62 .

The friction plate 74 is formed in a substantially annular plate shape, and is provided on the outer edge portion of the support portion 73 on the side of the plate portion 622 of the output shaft 62 . The friction plate 74 is fixed to the support portion 73 . The friction plate 74 can come into contact with the plate portion 622 by deforming the outer edge portion of the support portion 73 toward the plate portion 622 .

The friction plate 75 is formed in a substantially annular plate shape, and is provided on the opposite side of the output shaft 62 to the plate portion 622 on the outer edge portion of the support portion 73 . The friction plate 75 is fixed to the support portion 73 .

The pressure plate 76 is formed in a substantially annular plate shape and is provided on the driven cam 50 side with respect to the friction plate 75 .

In the engaged state in which the friction plate 74 and the plate portion 622 are in contact with each other, that is, in the engaged state, a frictional force is generated between the friction plate 74 and the plate portion 622, and the frictional force increases depending on the magnitude of the frictional force. Relative rotation between the plate 74 and the plate portion 622 is restricted. On the other hand, in the non-engaged state where the friction plates 74 and the plate portion 622 are separated from each other, that is, they are not engaged with each other, no frictional force is generated between the friction plates 74 and the plate portions 622, and the friction plates 74 and 622 Relative rotation with the plate portion 622 is not restricted.

When the clutch 70 is engaged, 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 in the disengaged state, the torque input to the input shaft 61 is not transmitted to the output shaft 62 .

The cover 625 is formed in a substantially annular shape and is provided on the cylindrical portion 623 of the output shaft 62 so as to cover the side of the pressure plate 76 opposite to the friction plate 75 .

In this embodiment, the clutch device 1 includes a state changing section 90 instead of the state changing section 80 shown in the first embodiment. The state changing portion 90 has a diaphragm spring 91, a return spring 92, a release bearing 93, etc. as an "elastic deformation portion".

The diaphragm spring 91 is formed in the shape of a substantially circular disc spring, and is provided on the cover 625 so that one end in the axial direction, that is, the outer edge thereof, contacts the pressure plate 76 . Here, the diaphragm spring 91 is formed such that the outer edge thereof is located on the clutch 70 side with respect to the inner edge, and the portion between the inner edge and the outer edge is supported by the cover 625 . Further, the diaphragm spring 91 is elastically deformable in the axial direction. As a result, the diaphragm spring 91 urges the pressure plate 76 toward the friction plate 75 with one axial end, that is, the outer edge. As a result, the pressure plate 76 is pressed against the friction plate 75 and the friction plate 74 is pressed against the plate portion 622 . That is, the clutch 70 is normally in the engaged state.

In this embodiment, the clutch device 1 is a so-called normally closed type clutch device that is normally engaged.

The return spring 92 is, for example, a coil spring, and is provided so that one end abuts the end surface of the driven cam tubular portion 52 on the clutch 70 side.

A release bearing 93 is provided between the other end of the return spring 92 and the inner edge of the diaphragm spring 91 . The return spring 92 biases the release bearing 93 toward the diaphragm spring 91 side. The release bearing 93 bears the diaphragm spring 91 while receiving a thrust load from the diaphragm spring 91 . The biasing force of the return spring 92 is smaller than the biasing force of the diaphragm spring 91 .

As shown in FIG. 19, when the ball 3 is positioned at one end of the driving cam groove 400 and the driven cam groove 500, the distance between the driving cam 40 and the driven cam 50 is relatively small, and the distance between the release bearing 93 and the driven cam 50 is relatively small. A gap Sp<b>2 is formed between the driven cam barrel portion 52 and the end face thereof. Therefore, the friction plate 74 is pressed against the plate portion 622 by the biasing force of the diaphragm spring 91, the clutch 70 is engaged, and transmission of torque between the input shaft 61 and the output shaft 62 is allowed.

Here, when power is supplied to the coil 22 of the motor 20 under the control of the ECU 10 , the motor 20 rotates, torque is output from the speed reducer 30 , and the drive cam 40 rotates relative to the housing 12 . As a result, the ball 3 rolls from one end of the driving cam groove 400 and the driven cam groove 500 to the other end. Therefore, the driven cam 50 moves relative to the housing 12 and the drive cam 40 in the axial direction, that is, moves toward the clutch 70 . As a result, the gap Sp2 between the release bearing 93 and the end surface of the driven cam tubular portion 52 is reduced, and the return spring 92 is axially compressed between the driven cam 50 and the release bearing 93 .

When the driven cam 50 moves further toward the clutch 70 side, the return spring 92 is compressed to the maximum, and the release bearing 93 is pressed toward the clutch 70 side by the driven cam 50 . As a result, the release bearing 93 presses the inner edge of the diaphragm spring 91 and moves toward the clutch 70 against the reaction force from the diaphragm spring 91 .

When the release bearing 93 presses the inner edge of the diaphragm spring 91 and moves toward the clutch 70 , the inner edge of the diaphragm spring 91 moves toward the clutch 70 and the outer edge of the diaphragm spring 91 moves away from the clutch 70 . As a result, the friction plates 74 are separated from the plate portion 622, and the state of the clutch 70 is changed from the engaged state to the disengaged state. As a result, transmission of torque between the input shaft 61 and the output shaft 62 is interrupted.

The ECU 10 stops the rotation of the motor 20 when the clutch transmission torque becomes zero. Thereby, the state of the clutch 70 is maintained in the disengaged state. In this manner, the diaphragm spring 91 of the state changing portion 90 receives an axial force from the driven cam 50 and changes the state of the clutch 70 between the engaged state and the disengaged state according to the axial relative position of the driven cam 50 with respect to the housing 12 . It can be changed to an engaged state.

Also in this embodiment, the inner seal member 401 and the outer seal member 402 as "seal members" can keep the space between the housing space 120 and the clutch space 620 airtight or liquid-tight.

In this embodiment, the clutch device 1 does not include the oil supply section 5 shown in the first embodiment. That is, in this embodiment, the clutch 70 is a dry clutch.

Thus, the present invention can also be applied to a normally closed clutch device having a dry clutch.

As described above, <15> in the present embodiment, the state changing section 90 has the diaphragm spring 91 as the "elastic deformation section" capable of elastically deforming in the axial direction of the driven cam 50 as the "translation section". ing.

By controlling the rotational angular position of the motor 20, the thrust of the clutch 70 can be controlled with high accuracy based on the displacement and load characteristics of the diaphragm spring 91. FIG. Therefore, variations in the load acting on the clutch 70 with respect to variations in the stroke of the driven cam 50 can be reduced. Accordingly, similarly to the first embodiment, highly accurate load control becomes possible, and the clutch device 1 can be controlled with high accuracy.

(Third Embodiment)
A part of the clutch device according to the third embodiment is shown in FIG. The third embodiment differs from the first embodiment in the configuration of the bearing portion and the like.

This embodiment includes a bearing portion 152 instead of the bearing portion 151 shown in the first embodiment. The bearing portion 152 has a plurality of bearing rolling elements 183 that roll in the circumferential direction of the rotor 23 and rotatably support the rotor 23 , and a lubricant 184 that lubricates the surroundings of the bearing rolling elements 183 . The bearing portion 152 rotatably supports the rotor 23 via the sun gear 31 . Here, only one bearing portion 152 that rotatably supports the rotor 23 is provided.

More specifically, the bearing portion 152 has a support 181 , a support recess 182 , bearing rolling elements 183 and a lubricant 184 .

The support 181 is made of metal, for example, and has a substantially cylindrical shape. The support recess 182 is formed so as to be recessed radially outward from the inner peripheral wall of the support 181 .

The bearing rolling element 183 is, for example, a “roller” formed of metal in a substantially cylindrical shape. The bearing rolling element 183 is provided in the support recess 182 such that its axis is substantially parallel to the axis of the support 181 . The bearing rolling element 183 is rotatable around the axis within the support recess 182 . In this embodiment, a total of eight bearing rolling elements 183 are provided at equal intervals in the circumferential direction of the support 181, for example.

In the bearing portion 152, the outer peripheral wall of the support 181 is fitted to one end of the sun gear main body 310, that is, the inner peripheral wall of the end opposite to the sun gear tooth portion 311, and the bearing rolling elements 183 are fitted to the inner cylindrical portion of the housing. It is provided so as to contact the outer peripheral wall of 121 . Thereby, the rotor 23 is rotatably supported by the housing inner cylindrical portion 121 via the sun gear 31 and the bearing portion 152 . That is, the bearing portion 152 rotatably supports the rotor 23 .

Here, when the rotor 23 rotates relative to the housing inner cylindrical portion 121 , the bearing rolling elements 183 rotate within the support recess 182 .

Lubricant 184 is fluid such as grease. A lubricant 184 is provided around the bearing rolling elements 183 and in the support recesses 182 of the support 181 to lubricate the surroundings of the bearing rolling elements 183 . Thereby, the bearing rolling element 183 can roll smoothly between the support 181 and the housing 12 in the support recess 182 .

The dynamic viscosity of the lubricant 184 changes depending on the environmental temperature. The dynamic viscosity of the lubricant 184 increases, for example, as the environmental temperature decreases.

The outer diameter of the bearing portion 152, that is, the outer diameter of the support body 181, is smaller than the outer diameter of the bearing portion 151, that is, the outer diameter of the outer ring 172 shown in the first embodiment.

<4> The bearing portion 152 is a “roller bearing” having bearing rolling elements 183 as “rollers”. More specifically, the bearing portion 152 is a “single-row roller bearing” in which the bearing rolling elements 183 are arranged in one row in the axial direction of the support 181 (see FIG. 20).

Therefore, the size and cost of the bearing portion 152 can be reduced compared to the bearing portion 151 as the "ball bearing" shown in the first embodiment.

(Fourth embodiment)
A part of the clutch device according to the fourth embodiment is shown in FIG. The fourth embodiment differs from the first embodiment in the configuration of the bearing portion and the like.

This embodiment includes a bearing portion 152 instead of the bearing portion 151 shown in the first embodiment. Since the configuration of the bearing portion 152 is the same as that of the bearing portion 152 shown in the third embodiment, the description thereof is omitted.

The outer peripheral wall of the support 181 of the bearing 152 is fitted to the inner peripheral wall of the other end of the sun gear body 310 , that is, the end on the sun gear tooth 311 side. provided to come into contact with Thereby, the rotor 23 is rotatably supported by the drive cam body 41 via the sun gear 31 and the bearing portion 152 . That is, the bearing portion 152 rotatably supports the rotor 23 .

Thus, in this embodiment, there is only one bearing portion 152 that rotatably supports the rotor 23 .

The outer diameter of the bearing portion 152 is smaller than the outer diameter of the bearing portion 151 shown in the first embodiment, that is, the outer diameter of the outer ring 172 .

The magnet 230 is provided inside the outer peripheral wall of the rotor 23 instead of the outer peripheral wall of the rotor 23 . That is, the motor 20 is an interior magnet (IPM) motor.

The outer diameter of stator core 211 is the same as the outer diameter of stator core 211 of the first embodiment. Moreover, the radial length of the stator core 211 is larger than the radial length of the stator core 211 of the first embodiment. Therefore, the number of winding turns of the coil 22 can be increased as compared with the first embodiment.

In this embodiment, the bearing portion 152 having a small radial size is provided radially inward of the sun gear tooth portion 311 as an "input portion" to secure the radial space of the motor 20, thereby achieving the first embodiment. , the outer diameter of the rotor 23 is reduced, the radial length of the stator core 211 is increased, and the number of turns of the coil 22 is increased. As a result, the torque constant can be increased, and a high-output, high-torque motor can be realized.

Further, by using the motor 20 as an internal magnet type (IPM) motor, the processing cost of the magnet (permanent magnet) can be reduced, and the cost of the clutch device 1 as a whole can be reduced.

<6> As described above, the bearing portion 152 is provided radially inside the sun gear tooth portion 311 as the “input portion” and rotatably supports the rotor 23 . More specifically, bearing portion 152 rotatably supports rotor 23 via sun gear 31 provided integrally with rotor 23 .

In the present embodiment, by providing the bearing portion 152 having a small size in the radial direction inside the sun gear tooth portion 311 in the radial direction, compared to the first embodiment in which the bearing portion 151 is provided inside the rotor 23 in the radial direction, the motor 20 space in the radial direction can be secured. Thereby, the degree of freedom in designing the motor 20 can be improved.

(Fifth embodiment)
A portion of the clutch device according to the fifth embodiment is shown in FIG. The fifth embodiment differs from the first embodiment in the configuration of the sealing member and the like.

<14> In the present embodiment, an outer seal member 403 is provided instead of the outer seal member 402 shown in the first embodiment. The outer seal member 403 is formed in an annular shape from an elastic material such as rubber. The outer sealing member 403 is a so-called O-ring.

The outer seal member 403 is provided in an annular seal groove portion 358 formed in the outer peripheral wall of the gear outer cylindrical portion 357 . That is, the outer seal member 403 is provided so as to come into contact with the second ring gear 35 that rotates integrally with the drive cam 40 on the radially outer side of the drive cam 40 as a “rotating portion”.

The inner peripheral wall of housing outer cylindrical portion 123 is slidable with the outer edge portion of outer sealing member 403 . That is, the outer seal member 403 is provided so as to come into contact with the housing outer tubular portion 123 of the housing 12 . The outer seal member 403 is elastically deformed in the radial direction and seals between the gear outer cylindrical portion 357 and the inner peripheral wall of the housing outer cylindrical portion 123 in an air-tight or liquid-tight manner.

As described above, <14> in the present embodiment, the outer seal member 403 as a "seal member" is an O-ring.

Therefore, the configuration of the clutch device 1 can be simplified and the cost can be reduced.

(Sixth embodiment)
A part of the clutch device according to the sixth embodiment is shown in FIG. The sixth embodiment differs from the fifth embodiment in the configuration of the sealing member and the like.

In this embodiment, an outer seal member 404 is provided in place of the outer seal member 403 shown in the fifth embodiment. The outer seal member 404 is annularly formed of an elastic material such as rubber.

More specifically, the outer seal member 404 has a seal annular portion 940 , a first outer lip portion 941 , a second outer lip portion 942 , a first inner lip portion 943 and a second inner lip portion 944 . The seal annular portion 940, the first outer lip portion 941, the second outer lip portion 942, the first inner lip portion 943, and the second inner lip portion 944 are integrally formed.

The seal annular portion 940 is formed in a substantially annular shape. The first outer lip portion 941 is annularly formed over the entire circumferential range of the seal annular portion 940 so as to extend radially outward from the seal annular portion 940 and inclined to one side in the axial direction. The second outer lip portion 942 is annularly formed over the entire circumferential range of the seal annular portion 940 so as to extend obliquely from the seal annular portion 940 to the other side in the axial direction and radially outward. The first inner lip portion 943 is annularly formed over the entire circumferential range of the seal annular portion 940 so as to extend radially inward from the seal annular portion 940 and inclined to one side in the axial direction. The second inner lip portion 944 is annularly formed over the entire circumferential range of the seal annular portion 940 so as to extend radially inward from the seal annular portion 940 and inclined toward the other side in the axial direction. As a result, the outer seal member 404 is formed to have an X shape in cross section along a virtual plane including all the axes (see FIG. 23).

As shown in FIG. 23 , the outer seal member 404 is provided in an annular seal groove portion 358 formed in the outer peripheral wall of the gear outer cylindrical portion 357 . Here, the tip portions of the first inner lip portion 943 and the second inner lip portion 944 are in contact with the seal groove portion 358 . That is, the outer seal member 404 is provided so as to come into contact with the second ring gear 35 that rotates integrally with the drive cam 40 on the radially outer side of the drive cam 40 as a “rotating portion”.

The tip portions of the first outer lip portion 941 and the second outer lip portion 942 are in contact with the inner peripheral wall of the housing outer cylindrical portion 123 . Therefore, the contact area between the outer seal member 404 and the housing outer cylinder portion 123 is smaller than the contact area between the outer seal member 403 and the housing outer cylinder portion 123 in the fifth embodiment. As a result, the sliding resistance acting on the outer seal member 404 when the drive cam 40 rotates can be reduced.

The first outer lip portion 941 and the second outer lip portion 942 of the outer seal member 404 are elastically deformed in the radial direction, and airtight or liquid-tight the gear outer cylindrical portion 357 and the inner peripheral wall of the housing outer cylindrical portion 123 . is sealed to The outer seal member 404 is a so-called lip seal.

As described above, <14> in the present embodiment, the outer seal member 404 as a "seal member" is a lip seal.

Therefore, the contact area between the outer seal member 404 and the housing outer cylindrical portion 123 can be reduced. As a result, when the drive cam 40 rotates, the sliding resistance acting on the outer seal member 404 can be reduced. Therefore, it is possible to suppress a decrease in efficiency during operation of the clutch device 1 .

(Other embodiments)
In the above-described embodiment, the number of bearing rolling elements in the bearing is set to be as small as possible within a range that can withstand the load applied to the bearing and within a range that satisfies the assembly conditions of the bearing. Indicated. On the other hand, in other embodiments, the number of bearing rolling elements may be any number within a range that can withstand the load applied to the bearing portion and within a range that satisfies the assembly conditions of the bearing portion. .

Further, in the first embodiment described above, the number of bearing rolling elements is smaller than the number of holding holes. In contrast, in other embodiments, the number of bearing rolling elements may be the same as the number of retaining holes.

Moreover, in the above-mentioned 1st Embodiment, the bearing part showed the example which is a "single-row ball bearing." In contrast, in another embodiment, the bearing portion may be a "multi-row ball bearing" in which a plurality of rows of bearing rolling elements as "balls" are arranged in the axial direction of the inner ring and the outer ring.

Further, in the above-described third embodiment, an example in which the bearing portion is a "single-row roller bearing" is shown. On the other hand, in another embodiment, the bearing portion may be a "multi-row roller bearing" in which a plurality of rows of bearing rolling elements as "rollers" are arranged in the axial direction of the support.

Moreover, in the above-described fourth embodiment, the example in which the bearing portion is provided radially inward of the input portion of the speed reducer has been described. On the other hand, in <6> another embodiment, the bearing portion may be provided radially outside the input portion and rotatably support the rotor.

Also, in other embodiments, the motor 20 may not have the magnet 230 as a "permanent magnet."

Further, in the above-described embodiment, an example in which the drive cam 40 as the “rotating portion” is formed separately from the second ring gear 35 of the speed reducer 30 is shown. On the other hand, <10> In other embodiments, the drive cam 40 as the “rotating portion” may be formed integrally with the second ring gear 35 of the speed reducer 30 . In this case, it is possible to reduce the number of parts and the number of assembly man-hours, thereby achieving cost reduction.

In another embodiment, the drive cam 40 as a "rotating part" may be formed such that the inner edge and the outer edge are at the same position in the axial direction.

Moreover, in other embodiments, the inner seal member 401 as a "seal member" is not limited to an oil seal, and may be an O-ring or a lip seal.

Further, in other embodiments, the "sealing member" that maintains airtightness or liquidtightness between the housing space and the clutch space may not be provided.

Further, in the above-described embodiment, the inner rotor type motor 20 in which the rotor 23 is provided radially inside the stator 21 is shown. On the other hand, in another embodiment, the motor 20 may be an outer rotor type motor in which the rotor 23 is provided radially outside the stator 21 .

Further, in the above-described embodiments, examples were shown in which the rotation translation unit is a rolling-element cam having a driving cam, a driven cam, and rolling elements. Conversely, in other embodiments, the rotary translational portion may have a rotary portion that rotates relative to the housing and a translational portion that moves axially relative to the housing as the rotary portion rotates relative to the housing. For example, it may be composed of a "slide screw" or a "ball screw".

In another embodiment, the elastically deformable portion of the state changing portion may be, for example, a coil spring or rubber as long as it can be elastically deformed in the axial direction. Also, in another embodiment, the state changing section may be composed only of a rigid body without having an elastic deformation section.

In other embodiments, the number of driving cam grooves 400 and driven cam grooves 500 is not limited to five, and may be any number as long as the number is three or more. Also, the number of balls 3 may be adjusted according to the number of drive cam grooves 400 and driven cam grooves 500 .

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

Further, in another embodiment, torque may be input from the second transmission section and torque may be output from the first transmission section via the clutch. Further, for example, when one of the first transmission section and the second transmission section is fixed so as not to rotate, the rotation of the other of the first transmission section and the second transmission section can be stopped by engaging the clutch. can. In this case, the clutch device can be used as a braking device.

Thus, the present disclosure is not limited to the above embodiments, and can be embodied in various forms without departing from the scope of the present disclosure.

The clutch system control and techniques described in the present disclosure may be provided by configuring a processor and memory programmed to perform one or more functions embodied by a computer program. It may be realized by Alternatively, the clutch system controls and techniques described in this disclosure may be implemented by a dedicated computer provided by configuring the processor with one or more dedicated hardware logic circuits. Alternatively, the clutch device controller and techniques described in this disclosure may be a processor configured with one or more hardware logic circuits and a processor and memory programmed to perform one or more functions. may be implemented by one or more dedicated computers configured by a combination of The computer program may also be stored as computer-executable instructions on a computer-readable non-transitional tangible recording medium.

Reference Signs List 1 clutch device 2 ball cam (rotational translation part) 12 housing 20 motor (prime mover) 21 stator 23 rotor 30 reduction gear 40 drive cam (rotating part) 50 driven cam (translation part) 61 input shaft (first transmission portion), 62 output shaft (second transmission portion), 70 clutch, 80, 90 state changing portion, 151, 152 bearing portion, 173, 183 bearing rolling element, 174, 184 lubricant, 311 sun gear tooth portion (input part)

Claims (15)

a housing (12);
a prime mover (20) having a stator (21) provided in the housing and a rotor (23) provided rotatably relative to the stator, and capable of being operated by energization and outputting torque from the rotor;
a reducer (30) capable of reducing and outputting the torque of the prime mover;
A rotating part (40) that rotates relative to the housing when the torque output from the speed reducer is input, and translation that moves relative to the housing in an axial direction when the rotating part rotates relative to the housing. a rotary translator (2) having a portion (50);
provided between a first transmission portion (61) and a second transmission portion (62) which are rotatable relative to the housing, and when engaged, the first transmission portion and the second transmission portion; a clutch (70) that allows torque transmission between the
A state changer (80, 90) that receives an axial force from the translating portion and can change the state of the clutch to an engaged state or a disengaged state according to the axial relative position of the translating portion with respect to the housing. )When,
A plurality of bearing rolling elements (173, 183) that roll in the circumferential direction of the rotor and rotatably support the rotor, and a lubricant (174, 184) that lubricates the surroundings of the bearing rolling elements. bearing portions (151, 152),
A clutch device in which the speed reducer has an input portion (311) provided coaxially and rotatably integrally with the rotor and receiving torque from the rotor. 2. A clutch according to claim 1, wherein the number of said bearing rolling elements is set to be as small as possible within a range that can withstand the load applied to said bearing portion and within a range that satisfies assembly conditions of said bearing portion. Device. The bearing portion has a cage (177) formed with a plurality of holding holes (178) capable of holding the bearing rolling elements,
3. A clutch device according to claim 1, wherein the number of said bearing rolling elements is less than the number of said holding holes. The clutch device according to any one of claims 1 to 3, wherein the bearing portion is a ball bearing or a roller bearing. The clutch device according to any one of claims 1 to 4, wherein the bearing portion is spaced apart from the input portion in the axial direction of the bearing portion. The clutch device according to any one of claims 1 to 4, wherein the bearing portion is provided radially inside or radially outside the input portion and rotatably supports the rotor. A clutch device according to any one of claims 1 to 6, wherein the prime mover has permanent magnets (230) provided on the rotor. The speed reducer is
a planetary gear (32) capable of revolving in the circumferential direction of the input portion while rotating while meshing with the input portion;
a carrier (33) that rotatably supports the planetary gear and is rotatable relative to the input section;
a first ring gear (34) meshable with the planetary gear; and
8. A second ring gear (35) capable of meshing with said planetary gear, having a number of teeth different from that of said first ring gear, and outputting torque to said rotating part. Clutch device according to the paragraph. The first ring gear is fixed to the housing,
The clutch device according to claim 8, wherein the second ring gear is rotatably provided integrally with the rotating portion. The clutch device according to claim 8 or 9, wherein the rotating portion is formed integrally with the second ring gear. The rotating part is a driving cam (40) having a plurality of driving cam grooves (400) formed on one surface,
The translation part is a driven cam (50) having a plurality of driven cam grooves (500) formed on one surface,
The rotation translation part is a rolling body cam (2) having a rolling body (3) provided to roll between the drive cam, the driven cam, and the drive cam groove and the driven cam groove. A clutch device according to any one of claims 1 to 10. The clutch device according to any one of claims 1 to 11, wherein the rotating portion is formed such that the inner edge portion (41) and the outer edge portion (44) are located at different positions in the axial direction. The prime mover and the speed reducer are provided in a housing space (120) formed inside the housing on the opposite side of the rotating part from the clutch,
The clutch is provided in a clutch space (620), which is a space on the opposite side of the rotating part from the accommodation space,
An annular seal member (401, 402, 403, 404). 14. A clutch device according to claim 13, wherein said seal member is an O-ring, a lip seal, or an oil seal. The clutch device according to any one of claims 1 to 14, wherein the state changing portion has an elastically deformable portion (81, 91) elastically deformable in the axial direction of the translational portion.

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