JP2019152298A - Transmission device for motor - Google Patents

Abstract

To provide a transmission device for a motor capable of changing power of the motor with reduction gear ratios different from each other, and miniaturizing the entire device.SOLUTION: In a transmission device 1 for a motor, a planetary gear mechanism PG0 for input, first and second planetary gear mechanisms PG1, PG2 for output, and a clutch mechanism MCL are disposed in an internal space of a hollow type motor 2. When a sun gear S0 of the planetary gear mechanism PG0 is connected to a first sun gear S1 of the first planetary gear mechanism PG1 by the clutch mechanism MCL, a reduction gear ratio is determined to a first reduction gear ratio, and when the sun gear S0 is connected to a second sun gear S2 of the second planetary gear mechanism PG2, a reduction gear ratio is determined to a second reduction gear ratio different from the first reduction ratio. Switching of the connection of the first sun gear S1 or the second sun gear S2 by the clutch mechanism MCL is performed by a phase synchronization mechanism MCC.SELECTED DRAWING: Figure 1

Description

  The present invention relates to a transmission for an electric motor that shifts and outputs the power of an electric motor at different reduction ratios.

  As a conventional speed reducer for an electric motor, for example, one disclosed in Patent Document 1 is known. As shown in FIG. 9, the reduction gear includes an input shaft SIA to which the power of the motor MA is input, a first sun gear S1A directly coupled to the input shaft SIA, a second sun gear S2A integrated therewith, A ring gear member RA having a first ring gear R1A and a second ring gear R2A integrated with each other and a plurality of first pinion gears P1A meshing with the first sun gear S1A and the first ring gear R1A are disposed on the outer peripheral side of the second sun gears S1A and SA2. A first carrier C1A that rotatably supports a second carrier C2A that rotatably supports a plurality of second pinion gears P2A that mesh with the second sun gear S2A and the second ring gear R2A.

  The first carrier C1A is fixed to the case CAA so as not to rotate, and the second carrier C2A is directly connected to the output shaft SOA. The diameter of the first sun gear S1A is smaller than the diameter of the second sun gear S2A, and the diameters of the first and second ring gears R1A and R2A are equal to each other.

  In the above configuration, the power of the motor MA is input via the input shaft SIA and is output from the output shaft SOA while being decelerated at a predetermined reduction ratio. This reduction ratio is determined according to the gear ratio between the ring gear member RA and the first sun gear S1A and the gear ratio between the ring gear member RA and the second sun gear S2A.

JP 2015-145708 A

  As shown in FIG. 9, in this conventional speed reducer, the motor MA is disposed outside the gear unit including the first sun gear S1A in the axial direction of the input / output shafts SIA and SOA. For this reason, the axial direction length of a reduction gear becomes large, and a reduction gear cannot be constituted compactly. In addition, this reduction device only performs a reduction operation at a constant reduction ratio determined by the number of teeth of the ring gear member RA and the first and second sun gears S1A and S2A, and cannot perform a transmission operation at different reduction ratios.

  The present invention has been made to solve the above-described problems, and provides a transmission for an electric motor capable of shifting the power of the electric motor with different reduction ratios and reducing the size of the entire apparatus. The purpose is to do.

  In order to achieve the above object, the invention according to claim 1 is a transmission for an electric motor that shifts the power of the electric motor 2 and outputs it from the output shaft SO, and the electric motor 2 has an annular rotor 2b. A planetary gear mechanism PG0 for input and a first planetary gear for output, which are constituted by a hollow type motor and are arranged in the inner space of the motor 2 coaxially with the output shaft SO and aligned with each other in the axial direction. A mechanism PG1 and a second planetary gear mechanism PG2 are provided. An input planetary gear mechanism PG0 includes a ring gear R integrated with the rotor 2b of the electric motor 2 and a non-moving portion (case CA in the embodiment (hereinafter the same in this section)). Meshed with the carrier C coupled to the ring gear R and meshed with the plurality of pinion gears P and the plurality of pinion gears P rotatably supported by the carrier C, and coaxial with the output shaft SO. The first planetary gear mechanism PG1 includes a first ring gear R1 integrated with the rotor 2b, a first carrier C1 connected with the output shaft SO, and a sun gear S integrally connected to the input shaft SI. The second planetary gear mechanism PG2 has a plurality of first pinion gears P1 meshed with the first ring gear R1 and rotatably supported by the first carrier C1, and a first sun gear S1 meshed with the plurality of first pinion gears P1. Is a second ring gear R2 integral with the rotor 2b, a plurality of second pinion gears P2 meshed with the second ring gear R2 and rotatably supported by the first carrier C1, and a second sun gear meshed with the plurality of second pinion gears P2. S2, and the gear ratio of the ring gear R and the sun gear S (input side gear ratio γ0), the gear ratio of the first ring gear R1 and the first sun gear S2 (first gear ratio). γ1) and the gear ratio (second gear ratio γ2) between the second ring gear R2 and the second sun gear S2 are set to be different from each other, and are coaxial with the input shaft SI in the internal space of the motor 2. And a clutch mechanism MCL that selectively connects / disconnects between the input shaft SI and the first sun gear S1 or the second sun gear S2, and a switching mechanism (phase synchronization mechanism MCC) for switching connection / disconnection by the clutch mechanism MCL. ).

  The transmission for an electric motor includes an input planetary gear mechanism and an output first and second planetary gear mechanisms arranged coaxially with an output shaft. These three planetary gear mechanisms are configured as described above, and in particular, the tooth number ratio between the ring gear and the sun gear is set to be different between the three.

  According to this configuration, when the motor operates and the rotor rotates, the power of the motor is input from the rotor to the ring gear, the first ring gear, and the second ring gear of each planetary gear mechanism integrated therewith, and for input. It is input to the sun gear of the planetary gear mechanism and the input shaft integrated therewith. Further, the connection / disconnection by the clutch mechanism is switched by the switching mechanism, whereby the clutch mechanism selectively selects between the input shaft and the first sun gear of the first planetary gear mechanism or the second sun gear of the second planetary gear mechanism. Connected / disconnected.

  When the input shaft is connected to the first sun gear, the power of the electric motor is decelerated at a predetermined first reduction gear ratio determined according to a gear ratio of the first ring gear and the first sun gear of the first planetary gear mechanism. Then, it is output to the output shaft via the first carrier. On the other hand, when the input shaft is connected to the second sun gear, the motive power of the electric motor is determined according to the gear ratio of the second ring gear and the second sun gear of the second planetary gear mechanism and the like, which is different from the first reduction ratio. And then output to the output shaft via the first carrier. As described above, the power of the electric motor can be changed at two different reduction ratios.

  The electric motor is a hollow electric motor, and the planetary gear mechanism for input, the first and second planetary gear mechanisms and the clutch mechanism are arranged and accommodated in the internal space. With this configuration, the length in the axial direction can be shortened as compared with a conventional reduction gear in which a motor and two rows of planetary gear mechanisms are arranged in parallel in the axial direction, and the transmission can be downsized accordingly. Furthermore, since the three planetary gear mechanisms, the input shaft, and the clutch mechanism are coaxially arranged, the meshing reaction force generated in the meshing portion is canceled out, and the load acting on the support portion such as the bearing Is reduced, it is possible to reduce the size of the support portion, thereby further reducing the size of the entire apparatus.

  According to a second aspect of the present invention, in the transmission for an electric motor according to the first aspect, the carrier C is connected to the non-moving portion via an intermittent hydraulic brake 21.

  According to this configuration, in a normal state, the carrier is fixed to the non-moving portion by connecting the hydraulic brake, and the impact torque generated when the clutch mechanism is engaged can be released via the hydraulic brake. Thereby, the load acting on the support portion such as a bearing is reduced, so that the support portion can be reduced in size, and the entire apparatus can be further reduced in size.

  Further, for example, when the electric motor fails, the hydraulic brake is cut off to cause the carrier to idle and stop the output of power from the transmission. This makes it possible to easily ensure the safety of the vehicle in the event of a failure of the electric motor, particularly when the electric motor is provided as an in-wheel motor in the vehicle, without requiring a dedicated disconnecting mechanism.

  According to a third aspect of the present invention, in the transmission for an electric motor according to the first or second aspect, the input shaft SI is formed in a hollow shape, extends coaxially in the input shaft SI, and has a circumference with respect to the input shaft SI. A control shaft (drum SC) configured to be capable of changing a relative phase, which is an angle of a direction, is further provided, and the clutch mechanism MCL is provided between the input shaft SI and the first sun gear S1 when the relative phase is a predetermined first phase. And a second clutch CL2 that connects between the input shaft SI and the second sun gear S2 when the relative phase is a predetermined second phase different from the first phase, and a switching mechanism Is configured by a phase synchronization mechanism MCC that is arranged outside the electric motor 2 and rotates the input shaft SI and the control shaft at the same rotational speed and controls the relative phase to the first phase or the second phase. Features .

  According to this configuration, the control shaft is provided in the hollow input shaft so as to extend coaxially, and this control shaft is configured such that the relative phase that is the angle in the circumferential direction with respect to the input shaft can be changed. . The clutch mechanism includes a first clutch that connects the input shaft and the first sun gear when the relative phase is a predetermined first phase, and an input shaft and a second sun gear that are connected when the relative phase is a predetermined second phase. A second clutch that connects between the two. Then, the input shaft and the control shaft are rotated at the same rotational speed by the phase synchronization mechanism arranged outside the electric motor, and the relative phase is controlled to the first phase or the second phase. Thus, the shift operation can be performed by selectively connecting the first clutch or the second clutch.

  According to a fourth aspect of the present invention, in the transmission for an electric motor according to the third aspect, the phase synchronization mechanism MCC is arranged coaxially with the input shaft SI and the control shaft and aligned in the axial direction. The third planetary gear mechanism PG3 has a third planetary gear mechanism PG3 and a fourth planetary gear mechanism PG4. The third planetary gear mechanism PG3 includes a third sun gear S3 connected to the input shaft SI, a rotatable third carrier C3, and a third sun gear S3. And a plurality of third pinion gears P3 rotatably supported by the third carrier C3 and a non-rotatable third ring gear R3 meshing with the plurality of third pinion gears P3, and the fourth planetary gear mechanism PG4 , A fourth sun gear S4 connected to the control shaft, a plurality of fourth pinion gears P4 meshed with the fourth sun gear S4 and rotatably supported by the third carrier C3, and a plurality of fourth pinion gears A fourth ring gear R4 meshed with P4 and connected to the operating portion 51. The third sun gear S3 and the fourth sun gear S4 have the same number of teeth, and the third ring gear R3 and the fourth ring gear R4 are mutually connected. It has the same number of teeth, and the operation unit 51 is configured to fix the fourth ring gear R4 when stopped and to rotate the fourth ring gear R4 according to the operation amount when operating. To do.

  According to this configuration, the phase synchronization mechanism includes the third planetary gear mechanism and the fourth planetary gear mechanism configured as described above, and the operation unit. When the operation unit is stopped, the fourth ring gear is fixed, so that the input shaft and the control shaft rotate at the same rotational speed while maintaining a relative phase. On the other hand, when the operation unit is operated, the fourth ring gear is rotated according to the operation amount, and the relative phase is changed according to the rotation amount. The input shaft and the control shaft rotate at the same rotational speed while maintaining the changed relative phase. Therefore, by operating the operation unit outside the electric motor, the relative phase can be controlled to the first phase or the second phase, and the speed change operation can be performed.

  Further, since the third and fourth planetary gear mechanisms are arranged in the axial direction, the axial length of the phase synchronization mechanism can be suppressed, and the phase synchronization mechanism can be configured in a compact manner.

1 is a diagram schematically illustrating an overall configuration of a transmission for an electric motor according to an embodiment of the present invention. It is a collinear diagram which shows the relationship of the rotation speed between the rotation elements in a transmission. It is a figure which expands and shows the clutch mechanism of a transmission partially. It is a figure which shows the operation state of a clutch mechanism when a transmission is 1st speed level. It is a figure which shows the operation state of the clutch mechanism in the 1st step in the upshift from the 1st gear to the 2nd gear, and the 3rd step in the downshift from the 2nd gear to the 1st gear. It is a figure which shows the operation state of the clutch mechanism in the 2nd stage in the upshift from 1st speed to 2nd speed, and the 2nd stage in the downshift from 2nd speed to 1st speed. It is a figure which shows the operation state of the clutch mechanism in the 3rd stage in the upshift from the 1st speed stage to the 2nd speed stage, and the 1st stage in the downshift from the 2nd speed stage to the 1st speed stage. It is a figure which shows the operation state of a clutch mechanism when a transmission is 2nd speed. It is a figure which shows schematically the conventional speed reducer for electric motors.

  Hereinafter, preferred embodiments of the present invention will be described in detail with reference to the drawings. A transmission 1 for an electric motor according to an embodiment of the present invention shown in FIG. 1 uses, for example, the power of an electric motor (hereinafter referred to as “motor”) 2 mounted as a power source in a four-wheeled vehicle (not shown) as 1 The gear is output from the output shaft SO after shifting at a reduction ratio of high speed (low speed) or second speed (high speed). The output shaft SO is connected to left and right drive wheels (both not shown) via a differential device or the like.

  The motor 2 is configured by a so-called flat hollow electric motor, and has a large internal space in the radial direction. The motor 2 is arranged coaxially with the output shaft SO, and includes an annular casing 2a (only a part is shown), a stator (not shown), and a rotor 2b. The stator is composed of a plurality of iron cores, coils, and the like, and is fixed to the casing 2a. The rotor 2b is composed of a plurality of magnets and the like, is disposed so as to face the stator, and is driven to rotate by supplying electric power to the stator.

  The transmission 1 includes a planetary gear mechanism PG0 for inputting power of a motor 2, a first planetary gear mechanism PG1 for first speed, and a second planetary gear mechanism PG2 for second speed, and the planetary gear mechanism. A clutch mechanism MCL and a phase synchronization mechanism MCC for selectively connecting / disconnecting between PG0 and the first or second planetary gear mechanism PG1, PG2 and setting the gear position to the first gear or the second gear are provided. Yes.

  The planetary gear mechanisms PG0, PG1, and PG2 are all of a single pinion type, are accommodated in the internal space of the motor 2, are coaxial with the output shaft SO, and are in the axial direction of the output shaft SO (hereinafter, simply referred to as “pindle gear mechanism”). They are arranged side by side in the “axial direction”.

  The planetary gear mechanism PG0 includes a ring gear R0, a carrier C0, a plurality of pinion gears P0, a sun gear S0, and the like. The ring gear R0 is provided integrally with the rotor 2b of the motor 2. The carrier C0 has a plurality of support shafts 11 integrally, and one end portion of these support shafts 11 is connected to a stationary case CA via a hydraulic brake 21.

  The hydraulic brake 21 is connected to an actuator (not shown), and normally, the actuator is maintained in a stopped state so that the carrier C0 is fixed to the case CA. When the actuator is operated, the carrier C0 is separated from the case CA.

  The plurality of pinion gears P0 are rotatably supported by the respective support shafts 11 and mesh with the ring gear R0. Sun gear S0 meshes with a plurality of pinion gears P0. The sun gear S0 is integrally connected to the hollow input shaft SI. The input shaft SI is disposed coaxially with the output shaft SO and is rotatably supported via a bearing (not shown). The input shaft SI extends on both sides of the sun gear S0 along the axial direction. On one side, the input shaft SI passes through first and second sun gears S1 and S2 described later of the first and second planetary gear mechanisms PG1 and PG2. It extends to the outside of the motor 2 on the other side.

  The first planetary gear mechanism PG1 includes a first ring gear R1, a first carrier C1, a plurality of first pinion gears P1, a first sun gear S1, and the like. The first ring gear R <b> 1 is provided integrally with the rotor 2 b of the motor 2. The first carrier C1 integrally has a plurality of support shafts 12 and is connected to the output shaft SO via the support shafts 12. The plurality of first pinion gears P1 are rotatably supported by the respective support shafts 12 and mesh with the first ring gear R1. The first sun gear S1 meshes with the plurality of first pinion gears P1. As described above, the input shaft SI integral with the sun gear S0 is passed through the first sun gear S1, and both S1 and SI are connected / disconnected by the first clutch CL1 of the clutch mechanism MCL.

  Further, the ratio γ1 between the number of teeth ZR1 of the first ring gear R1 and the number of teeth ZS1 of the first sun gear S1 (= ZR1 / ZS1, hereinafter referred to as “first gear ratio”) is the tooth of the ring gear R0 of the planetary gear mechanism PG0. The relationship is smaller than the ratio γ0 between the number ZR0 and the number of teeth ZS0 of the sun gear S0 (= ZR0 / ZS0, hereinafter referred to as “input side tooth number ratio”).

  The second planetary gear mechanism PG2 includes a second ring gear R2, the first carrier C1, a plurality of second pinion gears P2, a second sun gear S2, and the like. The second ring gear R2 is provided integrally with the rotor 2b of the motor 2. The plurality of second pinion gears P2 are rotatably supported by the support shaft 12 of the first carrier C1 and mesh with the second ring gear R2. The second sun gear S2 meshes with the plurality of second pinion gears P1. As described above, the input shaft SI is passed through the second sun gear S2, and both S2 and SI are connected / disconnected by the second clutch CL2 of the clutch mechanism MCL.

  The ratio γ2 of the number of teeth ZR2 of the second ring gear R2 and the number of teeth ZS2 of the second sun gear S2 (= ZR2 / ZS2, hereinafter referred to as “second tooth number ratio”) is the first planetary gear mechanism PG1. The relationship is smaller than the first gear ratio γ1, which is the gear ratio between the ring gear R1 and the first sun gear S1.

  From the configuration of the above three planetary gear mechanisms PG0, PG1, and PG2 and the connection relationship between the rotating elements, the relationship of the rotational speeds between the rotating elements when the motor 2 is operated can be expressed as shown in the alignment chart of FIG. Is done. That is, the rotor 2b (ring gear R0, first and second ring gears R1, R2) of the motor 2, the carrier C0, the output shaft SO (first carrier C1), and the input shaft SI (sun gear S0) (first sun gear S1 or first The two sun gears S2) constitute four rotating elements in a collinear relationship. The left and right points of the output shaft SO indicate the output shaft when the first sun gear S1 is connected to the input shaft SI (sun gear S0) and when the second sun gear S2 is connected to the input shaft SI, respectively. Represents the number of revolutions of SO.

In these cases, the speed ratios RD1 and RD2 (ratio between the rotational speed of the output shaft SO and the rotational speed of the motor 2) use the input side tooth number ratio γ0, the first and second tooth number ratios γ1 and γ2, and It represents with Formula (1) and (2), respectively.
RD1 = {(1 + γ0) / (1 + γ1)} − 1 (1)
RD2 = {(1 + γ0) / (1 + γ2)} − 1 (2)
As shown in these equations (1) and (2) and FIG. 2, the speed ratio RD approaches the value 0 and the reduction ratio becomes larger as the output-side tooth ratio is closer to the input-side tooth ratio γ0. . In the present embodiment, since the first tooth number ratio γ1 is set to a value closer to the input side tooth number ratio γ0 than the second tooth number ratio γ2, when the first sun gear S1 is connected to the input shaft SI. A larger speed reduction ratio is obtained and the gear position is set to a low first speed. On the other hand, when the second sun gear S1 is connected to the input shaft SI, a smaller speed reduction ratio is obtained and the speed stage is increased. The second speed is set.

  The clutch mechanism MCL and the phase synchronization mechanism MCC selectively connect / disconnect the input shaft SI and the first sun gear S1 or the second sun gear S2 in order to set such a gear position. Hereinafter, these configurations will be described.

  As shown in FIG. 1, the clutch mechanism MCL includes a first clutch CL1 that connects / disconnects the first sun gear S1 and a second clutch CL2 that connects / disconnects the second sun gear S2. A drum SC is provided to control the operation of both clutches CL1 and CL2.

  The drum SC has a solid shaft shape, is coaxially provided inside the hollow input shaft SI, and is rotatably supported by a bearing (not shown) at one end (the right end in FIG. 1). Has been. The other end of the drum SC is connected to the phase synchronization mechanism MCC as will be described later, so that when the motor 2 is operated, the drum SC is in the same direction as the input shaft SI (the direction of the arrow in FIG. 3 and the like). ) At the same speed.

  As shown in FIG. 4A, the first clutch CL1 is disposed between the input shaft SI and the inner peripheral surface of the first sun gear S1, and connects / disconnects between the input shaft SI and the first sun gear S1. It has a plurality of struts (31A, 31B) for blocking.

  These struts are composed of two pairs of acceleration side struts 31A and deceleration side struts 31B. Each pair of acceleration-side struts 31A and deceleration-side struts 31B have the same configuration, face each other in the circumferential direction, and are provided symmetrically with each other. Therefore, the acceleration side strut 31A and the related configuration will be described below first.

  As shown in FIG. 3, the input shaft SI is provided with a support shaft 35 extending in the axial direction. The acceleration side strut 31A has an arc shape, is engaged with the support shaft 35 from the inside in the radial direction, is rotatably attached, is provided substantially concentrically with the input shaft SI and the drum SC. It extends along the inner peripheral surface 34 of the sun gear S1.

  Further, the acceleration side strut 31A includes an engaging portion 31a extending from the support shaft 35 in the rotation direction of the input shaft SI (clockwise in FIG. 3) and a driven portion 31b extending in the opposite direction. The end surface is an engagement surface 31c. On the other hand, the inner circumferential surface 34 of the first sun gear S1 is provided with a plurality of inwardly projecting inner teeth 36 at equal intervals in the circumferential direction, and one of these inner teeth 36 has an acceleration side strut 31A. When the engaging surface 31c is engaged, the input shaft SI is connected to the first sun gear S1, and the acceleration side is brought into an in-gear state.

  The angle of the engagement surface 31c is such that the extension line of the reaction force (arrow X in the figure) acting from the first sun gear S1 acting via the inner teeth 36 is the center of the support shaft 35 or radially outward (in-gear side). ). With this setting, the acceleration side strut 31A is called into the in-gear side as the applied torque (reaction force) increases, and is thereby reliably held so as not to come out of the support shaft 35.

  In the input shaft SI, a spring accommodating groove 37 is formed at a position corresponding to the engaging portion 31a of the acceleration side strut 31A, and the set spring 32 is accommodated in the spring accommodating groove 37. The set spring 32 is composed of a coil spring, and both end portions thereof are in contact with the spring receiving portion 37a and the engaging portion 31a at the bottom of the spring accommodating groove 37, whereby the acceleration side strut 31A is made counterclockwise as shown in FIG. Always energized in the direction (in-gear side).

  Further, the input shaft SI is formed with a ball accommodation hole 38 penetrating in the radial direction at a position corresponding to the vicinity of the end of the driven portion 31b. The ball 33 is accommodated in the ball accommodation hole 38, and the ball 33 partially protrudes from the ball accommodation hole 38 toward the drum SC. A drum groove 39 is formed in the vicinity of the ball receiving hole 38 on the outer peripheral surface of the drum SC. The drum groove 39 extends in the circumferential direction by a predetermined length, and both end portions thereof are formed in a curved shape and serve as transition portions between adjacent outer peripheral surfaces.

  With the above configuration, when the drum groove 39 of the drum SC is positioned below the ball accommodation hole 38, the ball 33 accommodated in the ball accommodation hole 38 falls into the drum groove 39. In this state, the acceleration side strut 31A is rotated counterclockwise in FIG. 3 by the urging force of the set spring 32, and the engagement position where the engagement surface 31c can engage with the internal teeth 36 of the first sun gear S1. (Position of FIG. 3).

  On the other hand, when the drum groove 39 is not positioned below the ball receiving hole 38, the ball 33 is pushed outward by the outer peripheral surface of the drum SC against the urging force of the set spring 32, and accelerated. The driven part 31b of the side strut 31A is pressed. As a result, the acceleration side strut 31A rotates in the clockwise direction in the figure, and the engagement surface 31c is located at an engagement release position (not shown) incapable of engaging with the internal teeth 36 of the first sun gear S1.

  On the other hand, the deceleration side strut 31B has the same configuration as the acceleration side strut 31A described above, and is provided symmetrically with the acceleration side strut 31A including the other components such as the set spring 32 and the ball 33 in the circumferential direction. Yes. The drum groove 39 is shared between the acceleration side and deceleration side struts 31A and 31B, and two balls 33 and 33 for both struts 31A and 31B can be accommodated simultaneously.

  As shown in FIG. 4B, the second clutch CL2 has the same configuration as the first clutch CL1 described above, and includes two pairs of acceleration side struts 41A and deceleration side struts 41B, a set spring 32 and balls 33. A drum groove 39 is provided in the same manner.

  On the other hand, as is clear from the comparison of FIGS. 4A and 4B, the circumferential position (angle) of the drum groove 39 in the drum SC differs between the first and second clutches CL1 and CL2, and the drum The positional relationship between the groove 39 and the ball 33 is set to be different. With this configuration, one of the first and second clutches CL1 and CL2 is controlled to be in a connected state by changing a relative angle in the circumferential direction of the drum SC with respect to the input shaft SI (hereinafter referred to as “relative phase” as appropriate). At the same time, the other gear is controlled to be in a cut-off state, so that the gear position can be switched to the first gear or the second gear.

  The phase synchronization mechanism MCC synchronizes this relative phase and changes it during shifting. As shown in FIG. 1, the phase synchronization mechanism MCC is disposed outside the motor 2 and includes a third planetary gear mechanism PG3 and a fourth planetary gear mechanism PG4. These planetary gear mechanisms PG3 and PG4 are both of a single pinion type, and are arranged coaxially with the input shaft SI and the drum SC and aligned with each other in the axial direction.

  The third planetary gear mechanism PG3 includes a third sun gear S3 coupled to the input shaft SI, a plurality of third pinion gears P3 meshed with the third sun gear S3 and rotatably supported by the third carrier C3, and a plurality of first gears P3. A third ring gear R3 that meshes with the three-pinion gear P3 is provided. The third ring gear R3 is fixed to a stationary case CA.

  The fourth planetary gear mechanism PG4 includes a fourth sun gear S4 connected to the drum SC, and a plurality of rotatably supported by a third carrier C3 common to the third planetary gear mechanism PG3 in mesh with the fourth sun gear S4. A fourth pinion gear P4 and a fourth ring gear R4 that meshes with the plurality of fourth pinion gears P4 are provided. The fourth ring gear R4 is connected to an operation unit 51 configured by an actuator or the like, and is held in a fixed state when the operation unit 51 is stopped, and rotates according to the operation amount when the operation unit 51 is operated. Further, the tooth numbers ZS3 and ZS4 of the third sun gear S3 and the fourth sun gear S4 are equal to each other, and the tooth numbers ZR3 and ZR4 of the third ring gear R3 and the fourth ring gear R4 are equal to each other.

  From the above configuration, when the fourth ring gear R4 is fixed, the drum SC rotates at the same rotation speed while maintaining the relative phase with the input shaft SI. Further, when the fourth ring gear R4 is rotated by the operation of the operation unit 51, the relative phase is changed by rotating the drum SC with respect to the input shaft SI according to the rotation angle and the number of teeth. The drum SC and the input shaft SI rotate at the same rotational speed while maintaining the changed relative phase. In this manner, the relative phase is controlled by operating the operation unit 51 of the phase synchronization mechanism MCC.

  Hereinafter, the operation of the transmission 1 is upshifted from the first gear to the second gear and then downshifted from the second gear to the first gear with reference to FIGS. To do. FIG. 4 shows the first speed state, FIG. 8 shows the second speed state, and FIGS. 5 to 7 show the state during the upshift or the downshift between the two gear stages.

  In the first speed stage state (lower stage standby state) shown in FIG. 4, the relative phase is controlled to a predetermined first phase. In the first clutch CL1, the two balls 33, 33 on the acceleration side and the deceleration side are drums. The groove 39 is accommodated in the circumferential direction with a slight margin and falls into the groove 39. For this reason, the acceleration side and deceleration side struts 31A and 31B are located at the engaging positions where they engage with the internal teeth 36 of the first sun gear S1, and both the acceleration side and the reduction side are in the in-gear state. Is rotating at the same rotational speed as the input shaft SI.

  On the other hand, in the second clutch CL2, the drum groove 39 is slightly deviated from the deceleration side ball 33 toward the deceleration side (counterclockwise in FIG. 4), and the two balls 33 and 33 are lifted by the drum SC. Yes. For this reason, the acceleration side and deceleration side struts 41A and 41B are located at the disengagement position disengaged from the inner teeth 36 of the second sun gear S2, and both the acceleration side and the reduction side are in the off-gear state. Further, since the lower first clutch CL1 is in the in-gear state, the rotation speed of the second sun gear S2 is lower than that of the input shaft SI.

  When performing an upshift from this lower standby state, the operation unit 51 of the phase synchronization mechanism MCC is driven in a predetermined direction by a predetermined amount to rotate the fourth ring gear R4. As a result, the drum SC is rotated by a predetermined angle clockwise from the position of FIG. 4 to the position of FIG. 8 with respect to the input shaft SI, and the relative phase is controlled from the first phase to the predetermined second phase. .

  FIG. 5 shows an initial state in which the drum SC is rotated by a small angle from the position of FIG. In the first clutch CL1, since the balls 33 and 33 still remain in the drum groove 39, the operation state is not changed from the standby state in FIG. 4, and the acceleration side and deceleration side struts 31A and 31B are held in the engaged positions. Thus, the in-gear state on the acceleration side and the deceleration side is maintained.

  At this time, in the second clutch CL2, the deceleration-side ball 33 falls into the drum groove 39, so that the deceleration-side strut 41B moves to the engagement position. In this case, since the lower stage side is in the in-gear state as described above, the rotation speed of the second sun gear S2 is lower than that of the input shaft SI, and the second sun gear S2 is overtaken by the input shaft SI. Torque is not transmitted via the side strut 41B.

  When the drum SC rotates from the position of FIG. 5 to the position of FIG. 6, in the first clutch CL1, the drum groove 39 is disengaged from the ball 33 on the deceleration side, whereby the ball 33 is pushed up by the drum SC. At that time, when a load is generated on the acceleration side, the deceleration side strut 31B is released (moved from the engagement position to the engagement release position) with almost no load, while the first sun gear S1 is Since it is driven by the acceleration side strut 31A in the in-gear state and rotates at the same rotational speed as the input shaft SI, no further upshift operation is performed.

  On the other hand, when a load is generated on the speed reduction side, the engagement surface 31c of the speed reduction strut 31B is pressed against the inner teeth 36 of the first sun gear S1, so that the speed reduction strut 31B can be released. The torque of the drum SC that overcomes the deceleration load is required, and when the torque of the drum SC overcomes the deceleration load, the deceleration side strut 31B is released. Along with this, the input shaft SI becomes free with respect to the first sun gear S1, and the rotational speed of the input shaft SI begins to decrease. On the other hand, in the second clutch CL2, when the reduced rotational speed of the input shaft SI coincides with the rotational speed of the second sun gear S2, the deceleration side strut 41B in the engagement position is engaged with the internal teeth 36, The in-gear on the deceleration side is completed, and the rotation speed of the second sun gear S2 matches the rotation speed of the input shaft SI.

  As described above, during upshifting, the upshift operation does not proceed during acceleration load, and the upshift operation proceeds during deceleration load. Thus, it is possible to reliably prevent the occurrence of the in-gear on the acceleration side at the same time.

  When the drum SC rotates from the position of FIG. 6 to the position of FIG. 7, in the second clutch CL2, the acceleration side strut 41A moves to the engagement position by the acceleration side ball 33 falling into the drum groove 39. At this time, when an acceleration load is generated, the engaging surface of the acceleration side strut 41A engages with the inner teeth 36 of the second sun gear S2 as soon as it matches. Thereby, the in-gear on the acceleration side is completed, and at this point, the upshift to the second gear is substantially completed.

  When the in-gear in the second clutch CL2 is completed in this way, in the first clutch CL1, the rotational speed of the input shaft SI is less than the rotational speed of the first sun gear S1, and the input shaft SI is overtaken by the first sun gear S1. Therefore, torque transmission through the acceleration side strut 31A is not performed.

  When the drum SC rotates from the position shown in FIG. 7 to the position shown in FIG. 8 and the relative phase reaches the second phase, in the first clutch CL1, the drum groove 39 is disengaged from the ball 33 on the acceleration side. 31A moves to the disengagement position with almost no load, thereby completing the upshift to the second gear.

  In the second gear state (upper standby state) shown in FIG. 8, in the second clutch CL <b> 2, the acceleration-side and deceleration-side balls 33 and 33 have fallen into the drum groove 39. For this reason, the acceleration side and deceleration side struts 41A and 41B are located at the engaged positions, the acceleration side and the deceleration side are in the in-gear state, and the second sun gear S2 rotates at the same rotational speed as the input shaft SI. .

  On the other hand, in the first clutch CL1, the drum groove 39 is slightly displaced from the acceleration side ball 33 to the acceleration side (clockwise in FIG. 8), and the balls 33 and 33 are lifted by the drum SC. For this reason, the acceleration side and deceleration side struts 31A and 31B are located at the disengagement position, and both the acceleration side and the deceleration side are in an off-gear state. Further, since the upper second clutch CL2 is in the in-gear state, the rotational speed of the first sun gear S1 is higher than that of the input shaft SI.

  When downshifting is performed from the upper standby state, the operation unit 51 of the phase synchronization mechanism MCC is driven by the same predetermined amount in the opposite direction to that of the above-described upshift, and the fourth ring gear R4 is rotated. Thereby, the drum SC rotates counterclockwise by a predetermined angle from the position of FIG. 8 to the position of FIG. 4 with respect to the input shaft SI, and the relative phase is controlled from the second phase to the first phase.

  When the drum SC rotates from the position shown in FIG. 8 to the position shown in FIG. 7, in the second clutch CL2, since the balls 33 and 33 still remain in the drum groove 39, the operation state is different from the standby state shown in FIG. First, the acceleration side and deceleration side struts 41A and 41B are held in the engaged positions, and the in-gear states on the acceleration side and the deceleration side are maintained.

  At this time, in the first clutch CL1, when the acceleration side ball 33 falls into the drum groove 39, the acceleration side strut 31A moves to the engagement position. In this case, since the upper stage side is in the in-gear state as described above, the rotation speed of the first sun gear S1 is higher than that of the input shaft SI, and the input shaft SI is overtaken by the first sun gear S1. Torque is not transmitted via the side strut 31A.

  When the drum SC rotates from the position shown in FIG. 7 to the position shown in FIG. 6, in the second clutch CL2, the drum groove 39 is disengaged from the acceleration side ball 33, whereby the ball 33 is pushed up by the drum SC. At that time, when a load is generated on the speed reduction side, the acceleration side strut 41A is released with almost no load, while the second sun gear S2 is driven by the speed reduction side strut 41B in the in-gear state. Since it rotates at the same rotational speed as the axis SI, no further downshift operation is performed.

  On the other hand, when a load is generated on the acceleration side, the engagement surface of the acceleration side strut 41A is pressed against the inner teeth 36 of the second sun gear S2, and therefore, to release the acceleration side strut 41A, The torque of the drum SC that overcomes the acceleration load is necessary, and the acceleration side strut 41A is released when the torque of the drum SC overcomes the acceleration load. Along with this, the input shaft SI becomes free with respect to the second sun gear S2, and the rotational speed of the input shaft SI starts to increase. On the other hand, in the first clutch CL2, when the increased rotational speed of the input shaft SI coincides with the rotational speed of the first sun gear S1, the acceleration side strut 31A in the engagement position is engaged with the internal teeth 36, The in-gear on the acceleration side is completed, and the rotation speed of the first sun gear S1 matches the rotation speed of the input shaft SI.

  As described above, during downshifting, the downshift operation does not proceed during deceleration load, and the downshift operation proceeds during acceleration load. Thus, it is possible to reliably prevent the occurrence of the in-gear on the acceleration side at the same time.

  When the drum SC rotates from the position shown in FIG. 6 to the position shown in FIG. 5, in the first clutch CL1, the speed reducing ball 33 falls into the drum groove 39, whereby the speed reducing strut 31B moves to the engaging position. At this time, if a deceleration load is generated, the engagement surface of the deceleration side strut 31B is engaged with the inner teeth 36 of the first sun gear S1 as soon as it matches. Thereby, the in-gear on the deceleration side is completed, and at this point, the downshift to the first gear is substantially completed.

  Further, when the in-gear in the first clutch CL1 is completed in this way, in the second clutch CL2, the rotational speed of the input shaft SI exceeds the rotational speed of the second sun gear S2, and the second sun gear S2 is overtaken by the input shaft SI. Therefore, torque transmission via the deceleration side strut 41B is not performed.

  When the drum SC rotates from the position shown in FIG. 5 to the position shown in FIG. 4 and the relative phase reaches the first phase, the drum groove 39 is disengaged from the ball 33 on the deceleration side in the second clutch CL2. 41B moves to the non-engagement position with almost no load, thereby completing the downshift to the first gear.

  As described above, according to the transmission 1 for an electric motor of the present embodiment, the relative phase of the drum SC with respect to the input shaft SI is controlled to the first phase or the second phase by the phase synchronization mechanism MCC, and the first clutch CL1. Alternatively, the power of the motor 2 can be changed at two different reduction ratios by selectively controlling the second clutch CL2 to the connected state.

  In addition, when the shift stage is upshifted from the first speed stage to the second speed stage, the first and second clutches CL1 and CL2 are controlled according to the relative phase being controlled from the first phase to the second phase. Since the two struts, that is, the deceleration side strut 41B, the deceleration side strut 31B, the acceleration side strut 41A, and the acceleration side strut 31A operate sequentially one by one as described above, the upshift can be performed seamlessly. Similarly, when the gear position is downshifted from the second speed to the first speed, the acceleration side strut 31A, the acceleration side strut 41A, and the like are controlled in accordance with the relative phase being controlled from the second phase to the first phase. Since the deceleration side strut 31B and the deceleration side strut 41B operate one by one in order as described above, the downshift can be performed seamlessly.

  Further, the planetary gear mechanism PG0, the first and second planetary gear mechanisms PG1, PG2, and the clutch mechanism MCL are disposed in the internal space of the motor 2, so that the motor and two rows of planetary gear mechanisms are arranged in parallel in the axial direction. Compared with a conventional speed reducer, the length in the axial direction can be shortened, and the transmission can be downsized accordingly. Further, since the three planetary gear mechanisms PG0 to PG2, the input shaft SI, the drum SC, the clutch mechanism MCC, and the like are arranged coaxially, the meshing reaction force generated in their meshing portions is offset, and the bearings and the like By reducing the load acting on the support portion, it is possible to reduce the size of the support portion, thereby further reducing the size of the entire apparatus.

  Further, the hydraulic brake 21 provided between the carrier C and the case CA can release the impact torque generated when the clutch mechanism is intermittent. Thereby, the load acting on the support portion such as a bearing is reduced, so that the support portion can be reduced in size, and the entire apparatus can be further reduced in size. Further, when the motor 2 fails, the actuator can be operated and the hydraulic brake 21 can be shut off, whereby the carrier C can be idled and the output of power from the transmission 1 can be stopped. This makes it possible to easily ensure the safety of the vehicle when the motor 2 fails, particularly when the motor 2 is provided as an in-wheel motor in the vehicle without the need for a dedicated disconnecting mechanism.

  In addition, this invention can be implemented in various aspects, without being limited to the described embodiment. For example, in the embodiment, the first and second clutches having the acceleration-side and deceleration-side struts are used as the clutch mechanism. Other types of clutch mechanisms can be used as long as they can selectively connect / disconnect with the sun gear S1 or the second sun gear S2.

  Further, as the switching mechanism, a phase synchronization mechanism that controls the relative phase to the first phase or the second phase is used. However, the present invention is not limited to this, and depending on the configuration of the clutch mechanism, the connection / Any other suitable mechanism can be employed as long as the switching can be switched.

  Furthermore, although embodiment is the example which applied the transmission 1 to the motor 2 as a motive power source of a vehicle, this invention is not limited to this, It can apply widely to another use. In addition, it is possible to appropriately change the detailed configuration within the scope of the gist of the present invention.

1 Gearbox for electric motor 2 Motor (electric motor)
2b Rotor 21 Hydraulic brake 51 Operation part SO Output shaft SI Input shaft SC Drum (control shaft)
CA case (non-moving part)
PG0 planetary gear mechanism for input R ring gear C carrier P pinion gear S sun gear PG1 first planetary gear mechanism R1 first ring gear C1 first carrier P1 first pinion gear S1 first sun gear PG2 second planetary gear mechanism R2 second ring gear P2 second Pinion gear S2 Second sun gear MCL Clutch mechanism CL1 First clutch CL2 Second clutch MCC Phase synchronization mechanism (switching mechanism)
PG3 Third planetary gear mechanism R3 Third ring gear C3 Third carrier P3 Third pinion gear S3 Third sun gear PG4 Fourth planetary gear mechanism R4 Fourth ring gear P4 Fourth pinion gear S4 Fourth sun gear γ0 Input side gear ratio (ring gear and sun gear) Tooth ratio)
γ1 first gear ratio (tooth ratio between the first ring gear and the first sun gear)
γ2 Second tooth ratio (tooth ratio between second ring gear and second sun gear)

Claims (4)

A speed change device for an electric motor that changes the power of the electric motor and outputs it from an output shaft,
The electric motor is composed of a hollow electric motor having an annular rotor,
An input planetary gear mechanism, an output first planetary gear mechanism, and a second planetary gear mechanism are disposed in an internal space of the electric motor so as to be coaxial with the output shaft and aligned in the axial direction. ,
The planetary gear mechanism for input includes a ring gear integral with the rotor of the electric motor, a carrier coupled to a non-moving portion, a plurality of pinion gears meshed with the ring gear and rotatably supported by the carrier, And a sun gear integrally connected to an input shaft extending coaxially with the output shaft.
The first planetary gear mechanism includes a first ring gear integral with the rotor, a first carrier coupled to the output shaft, and a plurality of gears meshed with the first ring gear and rotatably supported by the first carrier. A first pinion gear and a first sun gear meshing with the plurality of first pinion gears;
The second planetary gear mechanism includes a second ring gear integral with the rotor, a plurality of second pinion gears meshed with the second ring gear and rotatably supported by the first carrier, and the plurality of second pinion gears. A second sun gear that meshes,
The gear ratio between the ring gear and the sun gear, the gear ratio between the first ring gear and the first sun gear, and the gear ratio between the second ring gear and the second sun gear are set to be different from each other. And
A clutch mechanism that is arranged coaxially with the input shaft in the internal space of the electric motor and selectively connects / disconnects between the input shaft and the first sun gear or the second sun gear;
And a switching mechanism for switching connection / disconnection by the clutch mechanism.   The transmission for an electric motor according to claim 1, wherein the carrier is connected to the non-moving portion via an intermittent hydraulic brake. The input shaft is formed in a hollow shape,
A control shaft extending coaxially into the input shaft and configured to change a relative phase that is a relative angle in the circumferential direction with respect to the input shaft;
The clutch mechanism includes: a first clutch that connects the input shaft and the first sun gear when the relative phase is a predetermined first phase; and a predetermined second phase that is different from the first phase in the relative phase. A second clutch connecting between the input shaft and the second sun gear at the time of
The switching mechanism is arranged outside the electric motor, and rotates the input shaft and the control shaft at the same rotational speed, and controls the relative phase to the first phase or the second phase. The transmission for an electric motor according to claim 1 or 2, wherein the transmission is configured. The phase synchronization mechanism includes a third planetary gear mechanism and a fourth planetary gear mechanism that are arranged coaxially with the input shaft and the control shaft and aligned in the axial direction.
The third planetary gear mechanism includes a third sun gear coupled to the input shaft, a rotatable third carrier, a plurality of third gears meshed with the third sun gear and rotatably supported by the third carrier. A pinion gear and a non-rotatable third ring gear meshing with the plurality of third pinion gears,
The fourth planetary gear mechanism includes a fourth sun gear coupled to the control shaft, a plurality of fourth pinion gears meshed with the fourth sun gear and rotatably supported by the third carrier, and the plurality of fourth gears. A fourth ring gear meshing with the pinion gear and connected to the operation portion,
The third sun gear and the fourth sun gear have the same number of teeth, and the third ring gear and the fourth ring gear have the same number of teeth.
4. The electric motor according to claim 3, wherein the operation unit is configured to fix the fourth ring gear when stopped, and to rotate the fourth ring gear according to an operation amount when operating. 5. Gearbox.

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