JP6546300B1 - Transmission for motor - Google Patents

Abstract

When a differential gear is used as a differential mechanism, a transmission for an electric motor capable of reducing the length in the axial direction and reducing the size of the entire apparatus is provided. A transmission for an electric motor according to the present invention shifts the power of an electric motor and transmits it to a differential gear. The double pinion type first planetary gear mechanism PG1 has a sun gear S1 to which the motive power of the electric motor 2 is input, a carrier C1 supporting a plurality of first and second pinion gears P1 and P2, and a ring gear R1. The ring gear R1 is braked by the first and second brakes B2, respectively. Each second pinion gear P2 has a base portion P2B meshing with the first pinion gear P1 and the ring gear R1, and an extension portion P2E extending from the base portion P2B to the opposite side to the motor 2, and the differential gear 4 meshes with the extension portion P2E It has two sun gears S2, and is arranged in an inner space defined by a plurality of extension parts P2E. [Selected figure] Figure 1

Description

  The present invention relates to a transmission for a motor for shifting the power of a motor and distributing the power to two output shafts by a differential mechanism.

  As a conventional transmission for this type of motor, for example, one disclosed in Patent Document 1 is known. The electric motor transmission (hereinafter simply referred to as "transmission") is applied to an electric vehicle, and in the example shown in FIG. 6, a transmission connected to an electric motor (hereinafter referred to as "motor") A mechanism and a differential mechanism are provided, and these three mechanisms are arranged in this order in the axial direction of the output shaft of the electric vehicle and accommodated in the housing.

  The transmission mechanism has first and second clutches and a planetary gear mechanism, which are arranged in order from the motor side in the axial direction. The planetary gear mechanism has a double pinion gear integrally including first and second pinion gears having different numbers of teeth, and a carrier rotatably supporting the double pinion gear. The planetary gear mechanism has a first sun gear connected to the rotor of the motor via a first clutch and meshing with the first pinion gear, and a first ring gear meshing with the first pinion gear and fixed to the housing. Furthermore, the number of teeth is greater than that of the first ring gear, which is connected to the rotor of the motor via the second clutch, meshes with the second pinion gear, and meshes with the second sun gear having a larger number of teeth than the first sun gear and the second pinion gear. Has a small second ring gear.

  The differential mechanism is formed of a double pinion type planetary gear mechanism, and includes a ring gear connected to a second ring gear of the transmission mechanism, first and second planetary gears meshing with the ring gear in order, and first and second planetary gears. It has a carrier that is rotatably supported, and a sun gear that meshes with the second planetary gear. The sun gear of the differential mechanism is coupled to one output shaft of the electric vehicle, and the carrier is coupled to the other output shaft of the electric vehicle.

  In this transmission, when the first clutch is turned on and the second clutch is turned off, the power of the motor is input to the first sun gear of the transmission mechanism, and after being shifted at a predetermined speed reduction ratio, the second ring gear is used. , Is output to the ring gear of the differential mechanism. On the other hand, when the first clutch is turned off and the second clutch is turned on, the power of the motor is input to the second sun gear, and after shifting at a predetermined reduction ratio smaller than the above, the difference is generated via the second ring gear. It is output to the ring gear of the moving mechanism. The power of the motor input to the differential mechanism is distributed to the sun gear and the carrier, and transmitted to the two output shafts.

JP, 2002-104001, A

  As described above, in the conventional transmission, a double pinion type planetary gear mechanism is used instead of a normal differential gear as a differential mechanism for distributing the power of a shifted motor. This is because the first and second clutches of the transmission mechanism and the two gear trains of the planetary gear mechanism and the differential mechanism are arranged side by side in the axial direction of the output shaft, so the axial length is longer than that of the differential gear. The use of a small double pinion planetary gear mechanism is advantageous in suppressing the axial length.

  However, the double pinion type planetary gear mechanism is complicated in configuration as compared with the differential gear, resulting in the complication of the entire configuration of the device and an increase in cost. From this point of view, although it is preferable to use a differential gear as the differential mechanism, in that case, as a result of the differential gear having a large axial length being juxtaposed with the transmission mechanism, the axial length becomes large. It leads to the enlargement of the whole.

  The present invention has been made to solve such a problem, and in the case of using a differential gear as a differential mechanism, for an electric motor capable of shortening the axial length and achieving the downsizing of the entire device. The purpose is to provide a transmission.

  In order to achieve this object, the invention according to claim 1 changes the power of the motor 2 while shifting the power of the motor 2 and by means of a differential mechanism two output shafts (in the embodiment (same in the present embodiment) drive shafts SL and SR). ), And includes a double pinion type planetary gear mechanism (first planetary gear mechanism PG1) coaxially arranged with the output shaft. A carrier C1 rotatably supporting a sun gear S1 to be input, a plurality of first pinion gears P1 meshing with the sun gear S1, and a plurality of second pinion gears P2 meshing with the plurality of first pinion gears P1, and a plurality of second pinion gears P2 First gear for braking the carrier C1 (first brake B1, one-way clutch OWC) , A second brake B2 for braking the ring gear R1, and a plurality of second pinion gears P2 each have a base P2B meshing with the first pinion gear P1 and the ring gear R1, and a base P2B opposite to the motor 2 The differential mechanism includes the extending portion P2E, and the differential mechanism includes the differential gear 4 having the second sun gear S2 meshing with the extending portion P2E of the second pinion gear P2 as an input gear, and extends the plurality of second pinion gears P2 It is characterized in that it is disposed in the inner space defined by the part P2E.

  As described above, the electric motor transmission according to the present invention (hereinafter, appropriately referred to as the "transmission") includes a double pinion type planetary gear mechanism, and the differential mechanism is configured by a differential gear. The second pinion gear of the planetary gear mechanism has a base and an extension extending from the base to the opposite side of the motor, the base meshes with the first pinion gear and the ring gear, and the extension is an input gear of the differential gear It is engaged with the second sun gear. Also, first and second brakes are provided to brake the carrier and ring gear of the planetary gear mechanism, respectively.

  According to this configuration, the motive power of the motor is input to the sun gear of the planetary gear mechanism, and is input to the carrier and the ring gear via the first pinion gear and the base of the second pinion gear. As will be described later, the power of the electric motor is determined by the ratio of the number of teeth between the rotating elements of the planetary gear mechanism, which are different from each other when the carrier is braked by the first brake and when the ring gear is braked by the second brake. It is decelerated at the reduction ratio. Therefore, by switching the operation / stop of the first and second brakes, it is possible to shift the power of the electric motor in two-speed shift. The shifted power is input to the differential gear through the extension of the second pinion gear and the second sun gear, and is distributed to the two output shafts by the differential gear.

  Further, according to the present invention, a space is defined on the inner side by the extension portion extending from the base of the second pinion gear to the opposite side to the electric motor, and the differential gear is disposed and accommodated in the inner space. As a result, the axial length can be shortened as compared with the case where the differential gears are provided in parallel with the transmission mechanism in the axial direction, and the overall size of the device can be reduced. Further, since a differential gear having a simpler configuration than that of the double pinion type planetary gear mechanism is used as the differential mechanism, simplification of the overall configuration of the apparatus and cost reduction can be achieved.

  The invention according to claim 2 is the electric motor transmission as set forth in claim 1, wherein each of the plurality of second pinion gears P2 is smaller than the large gear P21 having the base portion P2B and the extension portion P2E and the large gear P21. It further comprises: a second ring gear R2 meshing with the small gear P22, and a third brake B3 for braking the second ring gear R2. The second ring gear R2 is composed of a double pinion gear integrally having a small gear P22 having a number of teeth. It features.

  According to this configuration, when the second ring gear is braked by the third brake, the motive power of the electric motor is, as described later, when the carrier is braked by the first brake and when the ring gear is braked by the second brake. The vehicle is decelerated at an intermediate predetermined speed reduction ratio. Therefore, by switching the operation / stop of the first to third brakes, it is possible to shift the power of the electric motor in three-speed shift.

  According to a third aspect of the present invention, in the transmission for a motor according to the second aspect, the small gear P22 of the second pinion gear P2 is disposed on the opposite side to the electric motor 2 with respect to the large gear P21. Are disposed in an inner space defined by the extension P2E of the large gear P21 and the small gear P22.

  According to this configuration, a large inner space, which is longer in the axial direction than in the case of claim 1, is formed by the extension of the large gear of the second pinion gear and the small gear disposed on the opposite side of the large gear to the motor. A differential gear is disposed in the inner space defined. This allows the inner space to accommodate more or less of the differential gear.

  According to a fourth aspect of the invention, in the electric motor transmission of the second or third aspect, the electric motor 2, the planetary gear mechanism, and the differential gear 4 are accommodated in a case CA, and the first to third brakes B1 to B1 are provided. B3 is characterized in that it is disposed on the peripheral wall portion of the case CA.

  According to this configuration, since the first to third brakes are arranged on the peripheral wall portion of the case accommodating the motor, the planetary gear mechanism and the differential gear, the clutch for the same purpose is in the axial direction with respect to the planetary gear mechanism etc. As compared with the conventional case of juxtaposed in parallel, the axial direction length can be shortened, and the further miniaturization of the entire device can be achieved.

  The invention according to claim 5 is the transmission for a motor according to any one of claims 1 to 4, wherein the single pinion type is disposed between the motor 2 and the planetary gear mechanism coaxially with the output shafts SL and SR. The second planetary gear mechanism PG0 rotatably supports the sun gear S0 connected to the motor 2, the non-rotatable ring gear R0, and the pinion gear P0 meshing with the sun gear S0 and the ring gear R0. And a carrier C0 connected to the sun gear S1 of the planetary gear mechanism.

  According to this configuration, the power of the motor is input to the sun gear of the second planetary gear mechanism, decelerated at a predetermined reduction ratio, and then input to the sun gear of the planetary gear mechanism through the carrier. As a whole, a larger reduction ratio can be obtained as the entire transmission.

FIG. 1 schematically shows a transmission for an electric motor according to a first embodiment of the present invention. It is a figure which shows the 1st planetary gear mechanism etc. of the transmission for electric motors. It is a table | surface which shows the example of a setting of the number of teeth of the main rotating element in the transmission for electric motors of FIG. It is an operation table of the transmission for electric motors of FIG. FIG. 7 is a velocity diagram showing the relationship between the rotational speeds of the rotary elements in the forward high speed (Hi) mode of the electric motor transmission. FIG. 6 is a velocity diagram showing the relationship between the rotational speeds of the rotary elements in the forward low speed mode of the electric motor transmission. FIG. 7 is a velocity diagram showing the operation of the electric motor transmission of FIG. 1 together; It is a figure showing roughly the transmission for electric motors by a 2nd embodiment of the present invention. It is a table | surface which shows the example of a setting of the number of teeth of the main rotating element in the transmission for electric motors of FIG. It is an operation | movement table of the transmission for electric motors of FIG. FIG. 9 is a velocity diagram showing the relationship between the rotational speeds of the rotary elements in the forward middle speed (Mid) mode of the electric motor transmission of FIG. 8; FIG. 9 is a velocity diagram showing the operation of the electric motor transmission of FIG.

  Hereinafter, preferred embodiments of the present invention will be described in detail with reference to the drawings. A transmission for an electric motor (hereinafter referred to as "transmission" as appropriate) 1 according to a first embodiment of the present invention shown in FIG. 1 is, for example, an electric motor (hereinafter referred to as a power source) mounted on a four-wheel vehicle (not shown). The power of the motor 2) is changed. The shifted power is distributed by the differential gear 4 to the left and right drive shafts SL and SR coaxial with each other, and is further transmitted to the left and right drive wheels (not shown). The motor 2, the transmission 1, and the differential gear 4 are coaxially arranged with the drive shafts SL and SR in the axial direction, and are accommodated in the case CA.

  The motor 2 has an annular stator 2a and a rotor 2b. The stator 2a is configured of a plurality of iron cores, coils, etc., and is fixed to the case CA. The rotor 2b is composed of a plurality of magnets and the like, and is disposed to face the stator 2a, and is rotationally driven in the forward or reverse direction by the supply of power to the stator 2a.

  The transmission 1 includes a first planetary gear mechanism PG1, a second planetary gear mechanism PG0, a first brake B1, a second brake B2, and the like. These components are arranged coaxially with the drive shafts SL and SR, and the second planetary gear mechanism PG0 and the first planetary gear mechanism PG1 are arranged side by side in this order from the motor 2 side There is.

  The second planetary gear mechanism PG0 decelerates the power of the motor 2 at a constant reduction ratio and outputs it to the first planetary gear mechanism PG1. The second planetary gear mechanism PG0 is a single having a sun gear S0, a plurality of pinion gears P0, a carrier C0 and a ring gear R0. It is composed of a pinion type planetary gear mechanism.

  The sun gear S0 is integrally provided on the rotor 2b of the motor 2. The plurality of pinion gears P0 are rotatably supported by the carrier C0 and mesh with the sun gear S0 and the ring gear R0. The carrier C0 is connected to a sun gear S1 described later of the first planetary gear mechanism PG1. The ring gear R0 is fixed to the case CA.

  On the other hand, the first planetary gear mechanism PG1 is a double pinion type, and has a sun gear S1, a plurality of (three in this example) first pinion gears P1, a plurality of (three in this example) second pinion gears P2, a carrier It comprises C1 and a ring gear R1.

  As described above, the sun gear S1 is coupled to the carrier C0 of the second planetary gear mechanism PG0. The first and second pinion gears P1 and P2 are rotatably supported by the support shafts 11 and 12 of the carrier C1, respectively, and the first pinion gear P1 meshes with the sun gear S1. The second pinion gear P2 has a base P2B and an extension P2E extending from the base P2B to the differential gear 4 side. The second pinion gear P2 meshes with the first pinion gear P1 and the ring gear R1 at the base P2B, and the differential gear at the extension P2E. It meshes with the second sun gear S2 which is an input gear of No.4.

  As shown in FIG. 3, the number of teeth ZS1 and ZR1 of the sun gear S1 and the ring gear R1 is equal to the number of teeth ZS0 and ZR0 of the sun gear S0 of the second planetary gear mechanism PG0 and the ring gear R0, respectively. Further, the number of teeth ZS2 of the second sun gear S2 is smaller than the number of teeth ZS1 and ZS0 of the above two sun gears S1 and S0.

  The support shaft 12 of the carrier C1 further extends from the extension portion P2E of the second pinion gear P2 to the differential gear 4 side, and the first brake B1 is connected to this portion. The second brake B2 is connected to the ring gear R1.

  The first and second brakes B1 and B2 are provided on the peripheral wall portion of the case CA, and are formed of, for example, a multi-plate friction clutch. The first brake B1 includes a ring-shaped inner 13a having a large number of clutch plates, a ring-shaped outer 13b having a large number of clutch plates alternately arranged with the plurality of clutch plates, and an actuator for driving the outer 13b. (Not shown) and the like. The inner 13a is connected to the support shaft 12 of the carrier C1, and the outer 13b is provided on the case CA so as to be axially movable and non-rotatable. In addition, the operation of the actuator is controlled by a control device (not shown), whereby the operation / stop of the first brake B1 is switched.

  Specifically, when the actuator is stopped, the clutch plate of the inner 13a and the clutch plate of the outer 13b are kept in the non-engaged state, whereby the carrier C1 is not braked and its rotation is allowed. In addition, when the actuator operates, the non-rotatable outer clutch plate of the outer 13b is moved in the axial direction and engaged while pressed against the inner clutch plate of the inner 13a, whereby the carrier C1 is braked and fixed non-rotatably Be done.

  The second brake B2 has an inner 14a, an outer 14b and an actuator (not shown) configured similarly to the first brake B1, the inner 14a is connected to the ring gear R1, and the outer 14b is attached to the case CA. It is axially movable and non-rotatable. In the stopped state of the actuator, the ring gear R1 is not braked, and its rotation is allowed, and when the actuator is actuated, the ring gear R1 is braked and fixed so as not to rotate.

  Further, a one-way clutch OWC is provided in parallel with the first brake B1 between the support shaft 12 of the carrier C1 and the case CA. The one-way clutch OWC is a general mechanical type, and has a ring-shaped outer provided integrally with the case CA and a ring-shaped inner provided integrally with the support shaft 12. In addition, when the carrier C1 rotates in the direction of advancing the drive shafts SL and SR, the one-way clutch OWC cuts off between the case CA and the carrier C1 to allow the carrier C1 to rotate while the carrier C1 is the drive shaft. When attempting to rotate SL and SR in the reverse direction, the case CA and the carrier C1 are connected, and the carrier C1 is set to be fixed in a non-rotatable manner.

  The differential gear 4 has a general configuration, and includes a rotatable differential case 4a, a plurality of pinion gears 4b rotatably supported by the differential case 4a, and left and right side gears 4c and 4d meshing with the pinion gear 4b. The left side gear 4c is connected to the left drive wheel via the left drive shaft SL, and the right side gear 4d is connected to the right drive wheel via the right drive shaft SR.

  The differential gear case 4a is integrally provided with a second sun gear S2 as an input gear. As described above, the second sun gear S2 meshes with the extension portion P2E of the second pinion gear P2 of the first planetary gear mechanism PG1. Further, as shown in FIG. 1, most of the differential gear 4 is disposed in an inner space defined by the extension portion P2E of the second pinion gear P2, the first brake B1 and the one-way clutch OWC, It is housed in Case CA as a whole. The left and right drive shafts SL and SR extend to the outside through the opening of the case CA.

  According to the above configuration, the motive power of the motor 2 is input from the rotor 2b to the sun gear S0 of the second planetary gear mechanism PG0, and the carrier C0 is decelerated by the second planetary gear mechanism PG0 at a predetermined speed reduction ratio. It is input to the sun gear S1 of the first planetary gear mechanism PG1 via the first planetary gear mechanism PG1. The input power is transmitted to the differential gear 4 via the second pinion gear P2 and the second sun gear S2 while being shifted by the first planetary gear mechanism PG1. The transmitted power is distributed to the left and right drive shafts SL and SR by the differential gear 4 and transmitted to the left and right drive wheels. Hereinafter, the shift operation by the transmission 1 will be described in detail.

First, since the second planetary gear mechanism PG0 is a single pinion type, and the ring gear R0 is fixed to the case CA, the sun gear S0 is an input, and the carrier C0 is an output, the second planetary gear mechanism PG0 decelerates The ratio (hereinafter referred to as “S0-S1 reduction ratio”) RS01 is expressed by the following equation (1) using the ring gear R0 and the number of teeth ZR0 and ZS0 of the sun gear S0.
RS01 = (ZR0 / ZS0) +1 (1)
Substituting the values in FIG. 3 into ZR0 and ZS0 on the right side, the S0-S1 reduction ratio RS01 is
RS01 = (147/36) + 1 = 5.083.

  Further, since the first planetary gear mechanism PG1 is a double pinion type, and the second pinion gear P2 of the first planetary gear mechanism PG1 meshes with the second sun gear S2 which is an input gear of the differential gear 4, gear change is achieved. The relationship of the number of rotations between the rotating elements in the device 1 is expressed, for example, as shown in the alignment chart of FIG. That is, four rotating elements are constituted by the sun gear S1 (carrier C0), the ring gear R1, the carrier C1 and the second sun gear S2, and a collinear relationship in which those four rotational speeds are aligned in a straight line is established.

Further, λ1 and λ2 in the same figure respectively indicate the speed ratio between the carrier C1 and the sun gear S1 (hereinafter referred to as “first lever ratio”), with the carrier C1 and the ring gear R1 as a reference (= 1) It is a speed ratio between the second sun gear S2 (hereinafter referred to as "the second lever ratio"), and is expressed by the following equations (2) and (3).
λ1 = (ZR1 / ZS1) (2)
λ2 = (ZR1 / ZS2) (3)

  In the transmission 1 having the above-described configuration, the forward high speed (Hi) mode, the forward low speed (Low) mode, and the reverse (RVS) mode are set as the operation mode, and the separation mode is further set. Hereinafter, the operation of the transmission 1 will be described for each operation mode with reference to the operation table of FIG. 4.

[Hi mode]
As shown in FIG. 4, in the Hi mode, while the motor 2 is driven in a predetermined direction for forward movement, the first brake B1 is stopped and the second brake B2 is operated to brake and fix the ring gear R1. Do. The relationship between the rotational speeds of the rotating elements at this time is expressed as shown in FIG. 5, and while the rotational speed of the ring gear R1 becomes 0 and the power of the sun gear S1 is reversed and decelerated, the second sun gear It is transmitted to S2.

The reduction ratio (hereinafter referred to as "Hi mode reduction ratio") RS12 (Hi) between the sun gear S1 and the second sun gear S2 in the Hi mode is determined from the relationship between the first and second lever ratios λ1 and λ2 shown in FIG. It is expressed by the following equation (4).
RS12 (Hi) = (λ1-1) / (λ2 + 1)
= [(ZR1 / ZS1) -1] / [(ZR1 / ZS2) +1]
... (4)
Substituting the values in FIG. 3 into ZR1, ZS1 and ZS2 on the right side, the Hi mode reduction ratio RS12 (Hi) becomes
RS12 (Hi) = [(147/36) -1] / [(147/93) +1] = 1.195.

  Further, the total reduction ratio between the motor 2-differential gear 4 in the Hi mode is the S0-S1 reduction ratio RS01 (= 5.083) and the Hi mode reduction ratio RS12 (Hi) (= 1. 1) according to the equation (1). 195) to give 5.083 × 1.195 = 6.074.

[Low mode]
In the low mode, the first and second brakes B1 and B2 are stopped while the motor 2 is driven in a predetermined direction for forward movement. In this state, when carrier C1 and ring gear R1 idle, carrier C1 tries to rotate in the same direction as sun gear S1, that is, in the direction to reverse drive shafts SL and SR, so the one-way set as described above The clutch OWC operates. As a result, the case CA and the carrier C1 are connected, and the carrier C1 is non-rotatably fixed. The relationship between the rotational speeds of the rotating elements at this time is expressed as shown in FIG. 6, and while the rotational speed of the carrier C1 becomes 0, the power of the sun gear S1 is reversed and decelerated more than in the Hi mode. In the state, it is transmitted to the second sun gear S2.

The reduction ratio (hereinafter referred to as "Low mode reduction ratio") RS12 (Low) between the sun gear S1 and the second sun gear S2 in the low mode is determined from the relationship between the first and second lever ratios λ1 and λ2 shown in FIG. It is expressed by the following formula (5),
RS12 (Low) = λ1 / λ2
= (ZR1 / ZS1) / (ZR1 / ZS2) (5)
Substituting the values in FIG. 3 into ZR1, ZS1 and ZS2 on the right side, the Low mode reduction ratio RS12 (Low) is
RS12 (Low) = (147/36) / (147/93) = 2.583
Thus, a reduction ratio larger than the Hi mode reduction ratio RS12 (Hi) (= 1.195) is obtained.

  Also, the total reduction ratio between the motor 2-differential gear 4 in the low mode is the product of the S0-S1 reduction ratio RS01 (= 5.083) and the low mode reduction ratio RS12 (Low) (= 2.583). It becomes 5.083x2.583 = 13.129.

[RVS mode]
In this RVS mode, while the motor 2 is driven in the direction opposite to the above-described forward direction, the first brake B1 is operated to brake and fix the carrier C1, and the second brake B2 is stopped, and the ring gear R1 Allow rotation of At this time, as shown in FIG. 7, the relationship between the rotational speeds of the rotating elements is opposite to that in the low mode described above, and the sun gear S1 rotates in the opposite direction to that in the low mode, and the rotational speed of the carrier C1 is Will be zero. Further, the power of the sun gear S1 is reversed and transmitted to the second sun gear S2 in a state decelerated at the same speed reduction ratio as that in the low mode, whereby reverse travel is performed.

[Disengagement mode]
The separation mode is selected when performing a separation operation that cuts off the transmission of power between the motor 2 and the drive wheels. In this separation mode, the first and second brakes B1 and B2 are stopped while the motor 2 is stopped. As shown in FIG. 7, in this state, the power of the drive wheels is transmitted through the differential gear 4, and although the carrier C1 and the ring gear R1 rotate, these only rotate freely, so the motor 2 and the drive wheels There is no transmission of power between.

  The operation of the transmission 1 is summarized as shown in FIG. 7 and is controlled by the operation of the motor 2 and the operation / stop of the first and second brakes B1 and B2 to advance in the low mode and the hi mode. The operation in the two-speed shift, the RVS mode and the disconnection mode is performed.

  In the present embodiment, as described above, the carrier C1 is fixed in the low mode by the one-way clutch OWC, whereas the carrier C1 is fixed in the RVS mode by the first brake B1. This is due to the following reasons. That is, in the low mode, a large braking torque is required to fix the carrier C1, while in the RVS mode, the torque of the motor 2 at that time is suppressed to the required amount on the vehicle's merchandise. , The carrier C1 can be fixed with a small braking torque. From this relationship, the clutch capacity of the first brake B1 is reduced by selectively using the first brake B1 and the one-way clutch OWC as described above.

  As described above, according to the electric motor reduction gear transmission 1 of the present embodiment, the power of the motor 2 is shifted in two steps by switching the operation / stop of the first and second brakes B1 and B2, and the differential It can be distributed to the left and right drive shafts SL and SR by the gear 4. Further, since the power of the motor 2 is decelerated by the second planetary gear mechanism PG0, a larger reduction ratio can be obtained as the entire transmission.

  In addition, a space is drawn inside by the extension portion P2E extending from the base P2B of the second pinion gear P2 of the first planetary gear mechanism PG1 of the double pinion type opposite to the motor 2, the first brake B1 and the one-way clutch OWC. Most of the differential gear 4 is accommodated in the inner space. As a result, the axial length can be shortened as compared with the case where the differential gear is arranged in parallel with the transmission mechanism in the axial direction, and the overall size of the device including the transmission 1 and the differential gear 4 can be reduced. be able to. Further, since the differential gear 4 having a simpler configuration than the double pinion type planetary gear mechanism is used as the differential mechanism, simplification and cost reduction of the entire configuration of the apparatus can be achieved. Furthermore, since the first and second brakes B1 and B2 and the one-way clutch OWC are disposed on the peripheral wall portion of the case CA, compared with the conventional case where they are juxtaposed in the axial direction with respect to the planetary gear mechanism etc. The axial length can be shortened, and the overall size of the device can be further reduced.

  Further, the first brake B1 fixes the carrier C1 in the low mode, which requires a large braking torque, and the first brake B1, which fixes the carrier C1 in the RVS mode, which does not require a large braking torque. The clutch capacity of the brake B1 can be reduced.

  Next, a transmission for an electric motor (hereinafter, appropriately referred to as “transmission”) 51 according to a second embodiment of the present invention will be described with reference to FIG. 8 and the like. Compared with the transmission 1 according to the first embodiment described above, the transmission 51 has the second pinion gear P2 of the first planetary gear mechanism PG1 configured with a double pinion gear, and one of the double pinion gears. The main difference is that the meshing second ring gear R2 and the third brake B3 for braking the second ring gear R2 are added. In the following description of the transmission 51, the same or similar components as or to those of the transmission 1 are denoted by the same reference numerals, and the description will be made focusing on parts different from the transmission 1.

  As shown in FIG. 8, the second pinion gear P2 of the first planetary gear mechanism PG1 is constituted by a double pinion gear having a large gear P21 and a small gear P22 which are integral with each other. The large gear P21 has a base P21B and an extension P21E extending from the base P21B to the differential gear 4 side, meshing with the first pinion gear P1 and the ring gear R1 at the base P21B, and the second of the differential gear 4 at the extension P21E. It is engaged with the sun gear S2. The small gear P22 is disposed closer to the differential gear 4 than the extension P21E, meshes with the second ring gear R2, and the third brake B3 is connected to the second ring gear R2.

  FIG. 9 shows an example of setting the number of teeth of the main rotation element in the present embodiment. As is clear from the comparison with FIG. 3, the number of teeth Z of the rotary element such as the sun gear S0 used in the same manner as in the first embodiment is set to the same value as in the first embodiment. Further, the number of teeth ZP22 of the small gear P22 of the second pinion gear P2 added in this embodiment is smaller than the number of teeth ZP21 of the large gear P21, and the number of teeth ZR2 of the second ring gear R2 is correspondingly It is smaller than the number of teeth ZR1 of the ring gear R1.

  Further, as is clear from the comparison between FIG. 3 and FIG. 8, the first brake B1 and the one-way clutch OWC are on the opposite side (differential gear 4 side) to the motor 2 in the first embodiment with respect to the second brake B2. It arrange | positions, It arrange | positions at the motor 2 side in this embodiment. Thus, a larger space, which is longer in the axial direction than in the first embodiment, is defined inside the extension portion P21E of the large gear 21 of the second pinion gear P2 and the small gear P22 disposed on the extension side thereof. In the inner space, the entire differential gear 4 is accommodated.

  Further, the third brake B3 has an inner 15a, an outer 15b, and an actuator (not shown) which are provided on the peripheral wall portion of the case CA and configured similarly to the first and second brakes B1 and B2. The inner 15a is connected to the second ring gear R2, and the outer 15b is provided on the case CA so as to be axially movable and non-rotatable. Therefore, when the actuator is stopped, the second ring gear R2 is not braked and its rotation is allowed, and when the actuator is actuated, the second ring gear R2 is braked and fixed so as not to rotate.

  In the transmission 51 having the above configuration, the speed reduction operation by the second planetary gear mechanism PG0 of the transmission 1 described above and the transmission operation (Hi mode, Low mode, RVS mode, disconnecting mode) by the first planetary gear mechanism PG1 etc. The same is done. Further, since the number of teeth Z of the rotating elements such as the first and second planetary gear mechanisms PG1 and PG0 is the same as in the first embodiment, the S0-S1 reduction ratio RS01 obtained by these operations and the Hi mode reduction ratio RS12 (Hi) and Low mode reduction ratio RS12 (Low) are also the same as in the first embodiment.

  In addition to the above operation, in the transmission 51, the shifting operation in the forward middle speed (Mid) mode is performed as follows. First, the second pinion gear P2 of the double pinion type first planetary gear mechanism PG1 is constituted by a double pinion gear, the second ring gear R2 meshes with its small gear P22, and the second ring gear R2 is connected to the third brake B3. Because of this, the relationship between the rotational speeds of the rotating elements in the transmission 51 is expressed, for example, as shown in the alignment chart of FIG. That is, five rotating elements are constituted by the sun gear S1 (carrier C0), the ring gear R1, the second ring gear R2, the carrier C1 and the second sun gear S2, and the collinear relationship in which the five rotational speeds are aligned in a straight line. Is established.

Further, λ3 in the same figure is a speed ratio between the carrier C1 and the second ring gear R2 (hereinafter referred to as “third lever ratio”) based on the carrier C1 and the ring gear R1 and is represented by the following equation (6) Be done.
λ3 = (ZR1 / ZR2) · (ZP22 / ZP21) (6)

  As shown in FIG. 10, in the Mid mode, the first and second brakes B1 and B2 are stopped in a state where the motor 2 is driven in a predetermined direction for forward movement, and rotation of the carrier C1 and the ring gear R1 is permitted. The third brake B3 is operated to brake and fix the second ring gear R2. The relationship between the rotational speeds of the rotating elements at this time is expressed as shown in FIG. 11, and while the rotational speed of the second ring gear R2 becomes 0, the power of the sun gear S1 is reversed and decelerated, 2 transmitted to the sun gear S2.

The reduction ratio (hereinafter referred to as "Mid mode reduction ratio") RS12 (Mid) between the sun gear S1 and the second sun gear S2 in the mid mode is determined from the relationship between the first to third lever ratios λ1 to λ3 shown in FIG. It is expressed by the following equation (7).
RS12 (Mid)
= (Λ1-λ3) / (λ2 + λ3)
= [(ZR1 / ZS1)-(ZR1 / ZR2) (ZP22 / ZP21)]
/ [(ZR1 / ZS2) + (ZR1 / ZR2) · (ZP22 / ZP21)]
... (7)
Substituting the values in FIG. 9 into ZR1, ZS1, ZS2, ZP21 and ZP22 on the right side, the Mid mode reduction ratio RS12 (Mid) becomes
RS12 (Mid)
= [(147/36)-(147/141)-(20/26)]
/[(147/93)+(147/141)·(20/26)]=1.377
Thus, a reduction ratio between the Hi mode reduction ratio RS12 (Hi) (= 1.195) and the Low mode reduction ratio RS12 (Low) (= 2.583) is obtained.

  In addition, the total reduction ratio between the motor 2-differential gear 4 in the Mid mode is S0-S1 reduction ratio RS01 (= 5.083) and Mid mode reduction ratio RS12 (Mid) (= 1. 377) according to equation (1). And the product is 5.083 × 1.377 = 6.999.

  The operation of the transmission 51 described above is summarized as shown in FIG. 12, and by controlling the operation of the motor 2 and the operation / stop of the first to third brakes B1 to B3, the Low mode, Mid mode and Hi Operation is performed according to the forward three-speed shift depending on the mode, the RVS mode and the disconnection mode.

  As described above, according to the motor reduction gear 51 of the present embodiment, the power of the motor 2 is shifted in three steps by switching the operation / stop of the first to third brakes B1 to B3, and the differential The gears 4 can be distributed to the left and right drive shafts SL, SR and the drive wheels.

  The second pinion gear P2 of the first planetary gear mechanism PG1 is a double pinion gear integrally having a large gear P21 and a small gear P22, and an extension portion P2E of the large gear P21 and a small gear P22 extending on the extension side thereof. Thus, a large inner space longer in the axial direction than in the first embodiment is defined. Thus, the entire differential gear 4 can be accommodated in the inner space. Furthermore, as with the first and second brakes B1 and B2, the third brake B3 is also disposed on the peripheral wall portion of the case CA, so that the axial length can be shortened, and the overall size of the device can be reduced. Can.

  In addition, this invention can be implemented in various aspects, without being limited to the described embodiment. For example, the differential gear 4 is not limited to the general one described in the embodiment, and any other type may be employed. In the embodiment, the multi-plate friction clutch is used as the first to third brakes B1 to B3. However, any other appropriate type of brake may be used as long as the target rotating element can be braked. It is also good.

  In the embodiment, in order to reduce the clutch capacity of the first brake B1, the one-way clutch OWC is separately provided and used to fix the carrier C1 in the low mode, but the one-way clutch OWC is omitted. The fixing may be performed by the first brake B1.

  Furthermore, the embodiment is an example in which the transmissions 1 and 51 are applied to the motor 2 as a motive power source of a vehicle, but the present invention is not limited to this and is widely applied to motors of other applications used with a differential mechanism. It can apply. In addition, it is possible to change suitably the composition of details within the limits of the meaning of the present invention.

1 Transmission for motor 2 Motor (motor)
4 Differential gear (differential mechanism)
SL left drive shaft (output shaft)
SR right drive shaft (output shaft)
PG1 1st planetary gear mechanism (planet gear mechanism)
S1 Sun gear of first planetary gear mechanism (sun gear of planetary gear mechanism)
P1 First pinion gear of first planetary gear mechanism (first pinion gear of planetary gear mechanism)
P2 Second pinion gear of the first planetary gear mechanism (second pinion gear of the planetary gear mechanism)
P2B Base of second pinion gear P2E Extension of second pinion gear C1 Carrier of first planetary gear mechanism (carrier of planetary gear mechanism)
R1 Ring gear of first planetary gear mechanism (ring gear of planetary gear mechanism)
B1 1st brake OWC one-way clutch (1st brake)
B2 Second brake S2 Second sun gear 51 Transmission for motor P21 Large gear of second pinion gear P22 Small gear of second pinion gear B3 Third brake CA Case PG0 Second planetary gear mechanism S0 Sun gear of second planetary gear mechanism P0 Second Pinion gear of planetary gear mechanism C0 Carrier of second planetary gear mechanism R0 Ring gear of second planetary gear mechanism

Claims (5)

A transmission for a motor that shifts the power of the motor and distributes the power to two output shafts by a differential mechanism,
It has a double pinion type planetary gear mechanism coaxially arranged with the output shaft,
The planetary gear mechanism is
A sun gear to which the power of the motor is input;
A plurality of first pinion gears meshing with the sun gear, and a carrier rotatably supporting a plurality of second pinion gears meshing respectively with the plurality of first pinion gears;
And a ring gear engaged with the plurality of second pinion gears,
A first brake for braking the carrier;
And a second brake for braking the ring gear.
Each of the plurality of second pinion gears has a base that meshes with the first pinion gear and the ring gear, and an extension that extends from the base to the opposite side of the motor.
The differential mechanism includes a differential gear having a second sun gear meshing with the extension of the second pinion gear as an input gear, and an inner side defined by the extensions of the plurality of second pinion gears. A transmission for an electric motor, which is disposed in the space of Each of the plurality of second pinion gears is constituted by a double pinion gear integrally including a large gear having the base and the extension and a small gear having a smaller number of teeth than the large gear.
A second ring gear meshing with the small gear;
The transmission according to claim 1, further comprising: a third brake for braking the second ring gear.   The small gear of the second pinion gear is disposed opposite to the electric motor with respect to the large gear, and the differential gear is an inner side defined by the extended portion of the large gear and the small gear. A transmission for an electric motor according to claim 2, characterized in that it is disposed in the space of (1).   The motor, the planetary gear mechanism, and the differential gear are housed in a case, and the first to third brakes are disposed on a peripheral wall portion of the case. Gearbox for electric motor according to claim 1. It further comprises a single pinion type second planetary gear mechanism disposed coaxially with the output shaft between the motor and the planetary gear mechanism,
The second planetary gear mechanism is
A sun gear coupled to the motor;
Non-rotatable ring gear,
The motor according to any one of claims 1 to 4, further comprising: a sun gear and a pinion gear meshing with the ring gear rotatably supported, and a carrier connected to the sun gear of the planetary gear mechanism. Transmission.

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