CN110067836A - Motor speed change gear - Google Patents

Abstract

The present invention provides a speed change device for an electric motor, which can shorten the length in the axial direction when a differential gear is used as a differential mechanism, thereby realizing the miniaturization of the entire device. The speed change device (1) for an electric motor of the present invention shifts the power of the electric motor (2) and transmits it to a differential (4). The double pinion type first planetary gear mechanism (PG1) includes: a sun gear (S1) to which power from the electric motor is input; a carrier (C1) that supports a plurality of first and second pinions (P1, P2); and Ring gear (R1), carrier (C1) and ring gear (R1) are braked by 1st and 2nd brakes (B2) respectively. Each second pinion (P2) has a base (P2B) that meshes with the first pinion (P1) and the ring gear, and an extension (P2E) extending from the base to the opposite side of the motor, and the differential (4 ) has a second sun gear (S2) meshed with the extension, and is arranged in the inner space divided by the plurality of extension.

Description

Transmission device for electric motor

technical field

The present invention relates to a speed change device for an electric motor for shifting the power of an electric motor and distributing the power to two output shafts by a differential mechanism.

Background technique

As a conventional transmission device for such an electric motor, the device disclosed in Patent Document 1 is known, for example. The transmission for electric motors (hereinafter simply referred to as “transmission device”) is applied to an electric vehicle, and in the example shown in FIG. The three mechanisms are arranged in order in the axial direction of the output shaft of the electric vehicle, and are housed in the casing.

The speed change mechanism includes first and second clutches and a planetary gear mechanism, and these are arranged in a state of being aligned in the axial direction in this order from the motor side. The planetary gear mechanism includes a double pinion integrally provided with first and second pinions having mutually different numbers of teeth, and a carrier that rotatably supports the double pinion. In addition, the planetary gear mechanism includes: a first sun gear that is connected to the rotor of the motor via a first clutch and that meshes with the first pinion; and a first ring gear that meshes with the first pinion and is fixed to the housing , and also has: a second sun gear, which is connected to the rotor of the motor through a second clutch, meshes with the second pinion, and has a number of teeth greater than that of the first sun gear; and a second ring gear, which is connected to the second small gear. The gears are meshed, and the number of teeth is smaller than the number of teeth of the first ring gear.

The differential mechanism is composed of a double pinion type planetary gear mechanism, and includes: a ring gear connected to the second ring gear of the speed change mechanism; first and second planetary gears meshing with the ring gear in sequence; The gear support is a rotatable carrier, and a sun gear meshed with the second planetary gear. The sun gear of the differential mechanism is connected to one output shaft of the electric vehicle, and the carrier is connected to the other output shaft of the electric vehicle.

In this transmission, when the first clutch is engaged and the second clutch is disengaged, the power of the motor is input to the first sun gear of the transmission mechanism, and after shifting at a predetermined reduction ratio, it is output to the differential through the second ring gear. Ring gear of the drive mechanism. On the other hand, when the first clutch is disengaged and the second clutch is engaged, the power of the motor is input to the second sun gear, and is output to the differential through the second ring gear after shifting at a predetermined reduction ratio smaller than the above-mentioned reduction ratio. Ring gear of the drive mechanism. The power input to the motor of the differential mechanism is distributed to the sun gear and the carrier, and transmitted to the two output shafts.

Patent Document 1: Japanese Patent Laid-Open No. 2002-104001

SUMMARY OF THE INVENTION

As described above, in the conventional transmission, a double-pinion type planetary gear mechanism is used instead of a normal differential as a differential mechanism for distributing the power of the motor after shifting. The reason for this is that the first and second clutches of the transmission mechanism, the two gear sets of the planetary gear mechanism, and the differential mechanism are arranged in line in the axial direction of the output shaft, and due to this relationship, the length in the axial direction of use is smaller than that of the differential. The double pinion type planetary gear mechanism is more advantageous for restraining the length in the axial direction.

However, the double pinion type planetary gear mechanism is more complicated in structure than the differential, which leads to complicated structure and high cost of the entire device. From this point of view, a differential is preferably used as a differential mechanism. However, in this case, if a differential having a large axial length is arranged side by side with the speed change mechanism, the axial length will increase, and the entire device will be affected. Scale up.

The present invention has been made to solve the above-mentioned problems, and an object of the present invention is to provide a speed change device for an electric motor that can reduce the length in the axial direction and reduce the size of the entire device when a differential is used as a differential mechanism.

In order to achieve the above-mentioned object, a first aspect of the present invention is a speed change device 1 for an electric motor that shifts the power of the motor 2 and distributes the power to two output shafts (the following in the embodiment) by a differential mechanism The same as) drive shafts SL and SR), the transmission device 1 for electric motors is characterized in that the transmission device 1 for electric motors has a double pinion type planetary gear mechanism (first planetary gear mechanism PG1) arranged coaxially with the output shaft, The planetary gear mechanism includes a sun gear S1 to which the power of the motor 2 is input, and a carrier C1 that meshes the plurality of first pinions P1 with the sun gear S1 and the plurality of first pinions P1 respectively. The plurality of second pinions P2 are rotatably supported; and the ring gear R1 meshes with the plurality of second pinions P2, and the transmission device 1 for electric motors further includes: a first brake (a first brake B1, a one-way clutch OWC) for braking the carrier C1; and a second brake B2 for braking the ring gear R1, each of the plurality of second pinions P2 has: the same as the first pinion P1 and the ring gear A base P2B meshed with R1; and an extension P2E extending from the base P2B to the opposite side of the motor 2, and the differential mechanism is constituted by a differential 4 having an extension P2E with the second pinion P2 The meshed second sun gear S2 serves as an input gear, and the differential mechanism is arranged in the inner space partitioned by the extension portions P2E of the plurality of second pinion gears P2.

As described above, the electric motor transmission device of the present invention (hereinafter, referred to as "transmission device" as appropriate) has a double-pinion type planetary gear mechanism, and the differential mechanism is constituted by a differential. The second pinion gear of the planetary gear mechanism has a base portion and an extension portion extending from the base portion to the side opposite to the motor. 2 sun gear meshing. Moreover, the 1st and 2nd brakes which brake the carrier and the ring gear of the planetary gear mechanism, respectively, are provided.

According to this configuration, the power of the electric motor is input to the sun gear of the planetary gear mechanism, and is input to the carrier and the ring gear through the bases of the first pinion gear and the second pinion gear. As will be described later, when the carrier is braked by the first brake and when the ring gear is braked by the second brake, the power of the electric motor corresponds to the number of teeth between the rotating elements of the planetary gear mechanism. The speed is reduced by a predetermined reduction ratio that is different from each other. Therefore, by switching the operation and stop of the first and second brakes, the power of the electric motor can be shifted in a two-stage shift. The power after shifting is input to the differential through the extension of the second pinion gear and the second sun gear, and is distributed to the two output shafts by the differential.

Further, according to the present invention, a space is defined inside the extension portion extending from the base portion of the second pinion to the side opposite to the motor, and the differential is arranged and housed in the inside space. As a result, the length in the axial direction can be shortened as compared with the case where the differentials are arranged side by side with respect to the transmission mechanism in the axial direction, and the overall size of the device can be reduced. In addition, since a differential gear whose structure is simpler than that of the double pinion type planetary gear mechanism is used as the differential mechanism, the structure of the entire device can be simplified and the cost can be reduced.

A second aspect of the present invention is based on the transmission for an electric motor according to the first aspect, wherein each of the plurality of second pinions P2 is constituted by a double pinion, and the double pinion integrally includes a large gear P21 which is Having a base portion P2B and an extension portion P2E; and a pinion P22 having a smaller number of teeth than the large gear P21, the transmission device for electric motors further includes: a second ring gear R2 meshing with the pinion P22; and a third brake B3 , which is used to brake the second ring gear R2.

According to this configuration, when the second ring gear is braked by the third brake, as will be described later, the power of the motor is the same as when the carrier is braked by the first brake as described above, and the second ring gear is braked by the second brake. When the brake brakes the ring gear, it decelerates at a predetermined reduction ratio in the middle. Therefore, by switching the operation and stop of the first to third brakes, the power of the electric motor can be shifted in a three-step shift.

A third aspect of the present invention is based on the transmission for an electric motor according to the second aspect, wherein the pinion P22 of the second pinion P2 is arranged on the opposite side of the motor 2 with respect to the large gear P21 , and the differential 4 is arranged on the opposite side of the motor 2 . In the inner space divided by the extension P2E of the large gear P21 and the pinion P22.

According to this structure, the extension portion of the large gear of the second pinion and the pinion arranged on the opposite side of the motor with respect to the large gear are divided into a longer and larger in the axial direction than that of the first aspect of the present invention. An inner space in which a differential is arranged. As a result, more or all of the differential can be accommodated in the inner space.

A fourth aspect of the present invention is based on the transmission for an electric motor according to the second or third aspect, wherein the motor 2, the planetary gear mechanism, and the differential 4 are accommodated in the case CA, and the first to third brakes B1 -B3 is arrange|positioned at the peripheral wall part of case CA.

According to this configuration, the first to third brakes are arranged on the peripheral wall portion of the case housing the electric motor, the planetary gear mechanism, and the differential, so that the clutches of the same purpose are arranged in the axial direction with respect to the planetary gear mechanism and the like. Compared with the conventional case, the length in the axial direction can be shortened, and the miniaturization of the entire device can be further achieved.

A fifth aspect of the present invention is based on the electric motor transmission device according to any one of the first to fourth aspects, characterized in that the electric motor transmission device further includes a single-pinion type second planetary gear mechanism PG0, which The second planetary gear mechanism PG0 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 includes a sun gear S0 connected to the motor 2 and a non-rotatable ring gear R0 ; and a carrier C0 that supports rotatably the pinion P0 meshing with the sun gear S0 and the ring gear R0, and is connected to the sun gear S1 of the planetary gear mechanism.

According to this configuration, the power of the electric motor is input to the sun gear of the second planetary gear mechanism, and after being decelerated by a predetermined reduction ratio, it is input to the sun gear of the planetary gear mechanism through the carrier. Therefore, according to this deceleration, A larger reduction ratio can be obtained throughout the transmission.

Description of drawings

FIG. 1 is a diagram schematically showing a transmission for an electric motor according to a first embodiment of the present invention.

FIG. 2 is a diagram showing a first planetary gear mechanism and the like of the electric motor transmission.

FIG. 3 is a table showing an example of setting of the number of teeth of main rotating elements of the electric motor speed change device of FIG. 1 .

FIG. 4 is an operation table of the electric motor speed change device of FIG. 1 .

5 is a velocity diagram showing a relationship between rotational speeds of rotating elements in a forward high speed (Hi) mode of the electric motor transmission.

6 is a velocity diagram showing a relationship between rotational speeds of rotating elements in a forward low speed (Low) mode of the electric motor transmission.

FIG. 7 is a velocity diagram showing the operation of the electric motor speed change device of FIG. 1 in a collective manner.

8 is a diagram schematically showing a transmission for an electric motor according to a second embodiment of the present invention.

FIG. 9 is a table showing an example of setting of the number of teeth of the main rotational elements of the electric motor transmission of FIG. 8 .

FIG. 10 is an operation table of the electric motor transmission of FIG. 8 .

11 is a velocity diagram showing a relationship between rotational speeds of rotating elements in a forward intermediate speed (Mid) mode of the electric motor transmission of FIG. 8 .

FIG. 12 is a velocity diagram showing the operation of the electric motor transmission of FIG. 8 in a collective manner.

Label description

1: Transmission device for electric motor; 2: Motor (electric motor); 4: Differential (differential mechanism); SL: Left drive shaft (output shaft); SR: Right drive shaft (output shaft); PG1: 1st planet Gear mechanism (planetary gear mechanism); S1: the sun gear of the first planetary gear mechanism (sun gear of the planetary gear mechanism); P1: the first pinion of the first planetary gear mechanism (the first pinion of the planetary gear mechanism); P2: The second pinion of the first planetary gear mechanism (the second pinion of the planetary gear mechanism); P2B: The base of the second pinion; P2E: The extension of the second pinion; C1: The first planetary gear mechanism R1: the ring gear of the 1st planetary gear mechanism (the ring gear of the planetary gear mechanism); B1: the 1st brake; OWC: the one-way clutch (the 1st brake); B2: 2nd brake; S2: 2nd sun gear; 51: Transmission for electric motor; P21: Large gear of 2nd pinion; P22: Pinion of 2nd pinion; B3: 3rd brake; CA: Case; PG0 : The second planetary gear mechanism; S0: The sun gear of the second planetary gear mechanism; P0: The pinion of the second planetary gear mechanism; C0: The carrier of the second planetary gear mechanism; R0: The ring gear of the second planetary gear mechanism .

Detailed ways

Hereinafter, preferred embodiments of the present invention will be described in detail with reference to the accompanying drawings. A transmission for an electric motor (hereinafter, referred to as a “transmission device” as appropriate) 1 according to the first embodiment of the present invention shown in FIG. 1 corresponds to, for example, an electric motor (hereinafter, a motor) mounted on a four-wheeled vehicle (not shown) as a power source. The power called "motor") 2 is used for shifting. The power after shifting is distributed to the left and right drive shafts SL and SR coaxial with each other by the differential 4, and is further transmitted to the left and right drive wheels (not shown). The motor 2 , the transmission 1 , and the differential 4 are arranged coaxially with the drive shafts SL and SR so as to be aligned 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 composed of a plurality of iron cores, coils, and the like, and is fixed to the case CA. The rotor 2b is composed of a plurality of magnets or the like, and is arranged to face the stator 2a. By supplying electric power to the stator 2a, the rotor 2b is rotationally driven in the forward rotation direction or the reverse rotation direction.

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 in a state of being aligned in this order from the motor 2 side.

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

The sun gear S0 is integrally provided with 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, which will be described later, of the first planetary gear mechanism PG1. In addition, the ring gear R0 is fixed to the case CA.

On the other hand, the first planetary gear mechanism PG1 has a double-pinion type structure, and includes a sun gear S1, a plurality of (three in this example) first pinion gears P1, a plurality (three in this example) The second pinion P2, the carrier C1, the ring gear R1, and the like are constituted.

As described above, the sun gear S1 is coupled to the carrier C0 of the second planetary gear mechanism PG0. The first and second pinions P1 and P2 are rotatably supported by the support shafts 11 and 12 of the carrier C1, respectively, and the first pinion P1 meshes with the sun gear S1. The second pinion gear P2 has a base portion P2B and an extension portion P2E extending from the base portion P2B toward the differential 4 side, meshes with the first pinion gear P1 and the ring gear R1 at the base portion P2B, and meshes with the first pinion gear P1 and the ring gear R1 at the extension portion P2E as a differential The second sun gear S2 of the input gear of the transmission 4 meshes.

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

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

The first and second brakes B1 and B2 are provided on the peripheral wall portion of the casing CA, and are constituted by, for example, a multi-plate type friction clutch. The first brake B1 includes: an annular inner member 13a having a plurality of clutch plates; an annular outer member 13b having a plurality of clutch plates alternately arranged with the above-described plurality of clutch plates; Driven actuators (not shown), etc. The inner member 13a is connected to the support shaft 12 of the carrier C1, and the outer member 13b is provided on the casing CA so as to be movable in the axial direction and not rotatable. In addition, the operation of the actuator is controlled by a control device (not shown), and the operation and stop of the first brake B1 are switched.

Specifically, in a state where the actuator is stopped, the clutch plate of the inner member 13a and the clutch plate of the outer member 13b are maintained in a non-engaged state, so that the carrier C1 is allowed to rotate without being braked. Further, when the actuator operates, the clutch plate of the non-rotatable outer member 13b moves in the axial direction and is engaged in a state of being pressed against the clutch plate of the inner member 13a, whereby the carrier C1 is braked , is fixed to not rotate.

The second brake B2 includes an outer member 13b, an outer member 14b, and an actuator (not shown) having the same configuration as the first brake B1. The outer member 13b is connected to the ring gear R1, and the outer member 14b is movable in the axial direction. And it is set on the casing CA in a way that it cannot be rotated. In the stopped state of the actuator, the ring gear R1 is allowed to rotate without being braked, and when the actuator operates, the ring gear R1 is braked and fixed so as not to rotate.

In addition, a one-way clutch OWC connected in series with the first brake B1 is provided between the support shaft 12 of the carrier C1 and the case CA. The one-way clutch OWC has a normal mechanical structure, and includes an annular outer member integrally provided with the housing CA, and an annular inner member integrally provided with the support shaft 12 . In addition, the one-way clutch OWC is set so that when the carrier C1 rotates in the direction in which the drive shafts SL and SR advance, the one-way clutch OWC cuts between the case CA and the carrier C1 and allows the rotation of the carrier C1 On the other hand, when the carrier C1 rotates in the direction of retracting the drive shafts SL and SR, the one-way clutch OWC connects the case CA and the carrier C1 to fix the carrier C1 so as not to rotate.

The differential 4 has a normal structure and includes: a rotatable differential case 4a; a plurality of pinions 4b rotatably supported by the differential case 4a; and left and right side surfaces meshing with the pinions 4b Gears 4c, 4d. 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 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 extended 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 4 is arranged in the inner space divided by the extension P2E of the second pinion P2, the first brake B1, and the one-way clutch OWC, and the entirety is accommodated in the inner space. Housing CA. The left and right drive shafts SL and SR extend to the outside through the opening of the casing CA.

According to the above configuration, the power of the motor 2 is input from the rotor 2b to the sun gear S0 of the second planetary gear mechanism PG0, and is reduced by the second planetary gear mechanism PG0 by a predetermined reduction ratio, and then passes through the carrier C0. It is input to the sun gear S1 of the 1st planetary gear mechanism PG1. The input power is transmitted to the differential 4 via the second pinion P2 and the second sun gear S2 in a state of 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 4 and transmitted to the left and right drive wheels. Hereinafter, the shifting operation of the transmission device 1 will be described in detail.

First, since the second planetary gear mechanism PG0 is of a single-pinion type, and its ring gear R0 is fixed to the casing CA, and the sun gear S0 is used as an input and the carrier C0 is used as an output, the ring gear R0 and The number of teeth ZR0 and ZS0 of the sun gear S0 is represented by the following formula (1) to express the reduction ratio (hereinafter referred to as "S0-S1 reduction ratio") RS01 by the second planetary gear mechanism PG0.

RS01=(ZR0/ZS0)+1...(1)

If ZR0 and ZS0 on the right are substituted into the values in Figure 3, the S0-S1 reduction ratio RS01 is:

RS01=(147/36)+1=5.083.

In addition, since the first planetary gear mechanism PG1 is of a double pinion type, and the second pinion P2 of the first planetary gear mechanism PG1 meshes with the second sun gear S2 as the input gear of the differential 4, for example, as shown in FIG. The relationship between the rotational speeds of the rotational elements of the transmission 1 is represented as a collinear diagram of 5 . That is, the sun gear S1 (carrier C0), the ring gear R1, the carrier C1, and the second sun gear S2 constitute four rotating elements, and a collinear relationship in which the rotational speeds of these four rotating elements are aligned on a straight line is established.

In addition, λ1 and λ2 in the figure are respectively the speed ratio between the carrier C1 and the sun gear S1 (hereinafter referred to as "first lever ratio") based on the distance between the carrier C1 and the ring gear R1 (=1). , and the speed ratio between the carrier C1 and the second sun gear S2 (hereinafter referred to as "second lever ratio"), which can be expressed by the following equations (2) and (3).

λ1=(ZR1/ZS1)...(2)

λ2=(ZR1/ZS2)...(3)

In the transmission 1 having the above configuration, a forward high speed (Hi) mode, a forward low speed (Low) mode, and a reverse (RVS) mode are set as operation modes, and a disengagement mode is also 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 this Hi mode, in a state where the motor 2 is driven in a predetermined direction for forward movement, the first brake B1 is stopped, and the second brake B2 is actuated to brake the ring gear R1. and fixed. The relationship between the rotational speeds of the rotating elements at this time is shown in FIG. 5 . The rotational speed of the ring gear R1 is 0, and the power of the sun gear S1 is transmitted to the second sun gear S2 in a reversed and decelerated state.

From the relationship between the first and second lever ratios λ1 and λ2 shown in FIG. 5 , the reduction ratio between the sun gear S1 and the second sun gear S2 in this Hi mode can be expressed by the following formula (4) (hereinafter referred to as the "Hi mode reduction ratio") RS12(Hi).

RS12(Hi)=(λ1-1)/(λ2+1)

=[(ZR1/ZS1)-1]/[(ZR1/ZS2)+1]...(4)

If ZR1, ZS1, ZS2 on the right are substituted into the values in Figure 3, the Hi mode reduction ratio RS12(Hi) is:

RS12(Hi)=[(147/36)-1]/[(147/93)+1]=1.195.

In addition, the total reduction ratio between the motor 2 and the differential 4 in the Hi mode can be obtained through the S0-S1 reduction ratio RS01 (=5.083) of the above formula (1) and the Hi mode reduction ratio RS12 (Hi) (=1.195) The product is expressed as 5.083×1.195=6.074.

[Low Mode]

In this Low 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. In this state, the carrier C1 and the ring gear R1 are idling, and the carrier C1 is going to rotate in the same direction as the sun gear S1, even in the direction in which the drive shafts SL and SR are reversed. Therefore, the one-way clutch OWC is set as described above. working. As a result, the case CA and the carrier C1 are connected, and the carrier C1 is fixed so as not to rotate. The relationship between the rotational speeds of the rotating elements at this time can be expressed as shown in FIG. 6 , the rotational speed of the carrier C1 is 0, and the power of the sun gear S1 is transmitted to the second 2 Sun gear S2.

From the relationship between the first and second lever ratios λ1 and λ2 shown in FIG. 6 , the reduction ratio between the sun gear S1 and the second sun gear S2 in this Low mode (hereinafter referred to as the following equation (5) can be expressed "Low mode reduction ratio") RS12 (Low),

RS12(Low)=λ1/λ2

=(ZR1/ZS1)/(ZR1/ZS2)...(5)

If ZR1, ZS1, ZS2 on the right are substituted into the values in Figure 3, the Low mode reduction ratio RS12 (Low) is:

RS12(Low)=(147/36)/(147/93)=2.583, a larger reduction ratio can be obtained than that of the Hi mode reduction ratio RS12(Hi)(=1.195).

In addition, the total reduction ratio between the motor 2 and the differential 4 in the Low mode can be expressed by the product of the S0-S1 reduction ratio RS01 (=5.083) and the Low mode reduction ratio RS12 (Low) (=2.583), which is 5.083 ×2.583=13.129.

[RVS mode]

In this RVS mode, in a state where the motor 2 is driven in a direction opposite to the above-mentioned forward direction, the first brake B1 is actuated, the carrier C1 is braked and fixed, and the second brake B2 is stopped. , allowing the rotation of the ring gear R1. As shown in FIG. 7 , the relationship between the rotational speeds of the rotating elements at this time is positive and negative in the above-described Low mode, the sun gear S1 rotates in the opposite direction to that in the Low mode, and the rotational speed of the carrier C1 is zero. In addition, the motive power of the sun gear S1 is transmitted to the second sun gear S2 in a state in which the power of the sun gear S1 is reversed and decelerated at the same speed reduction ratio as in the Low mode, whereby the reverse travel is performed.

[Cut off mode]

The cut-off mode is selected during the cut-off operation that cuts off the transmission of power between the motor 2 and the drive wheels. In this disengagement 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 driving wheels is transmitted through the differential gear 4, and the carrier C1 and the ring gear R1 rotate, however, they are only idling, and therefore there is no communication between the motor 2 and the driving wheels. of power transmission.

Summarizing the operation of the transmission device 1 described above, as shown in FIG. 7 , by controlling the operation of the motor 2 and the operation and stop of the first and second brakes B1 and B2, the low mode and the high mode are executed. 2-step forward shifting, and actions based on RVS mode and cut-off mode.

In the present embodiment, as described above, the carrier C1 in the Low mode is fixed by the one-way clutch OWC, while the carrier C1 in the RVS mode is fixed by the first brake B1. This is based on the following reasons. That is, in the Low mode, a large braking torque is required to fix the carrier C1, and in the RVS mode, the torque of the motor 2 at this time is suppressed to a level necessary for the commerciality of the vehicle, so that it is possible to The carrier C1 is fixed with a small braking torque. Based on this relationship, by using the first brake B1 and the one-way clutch OWC separately as described above, the clutch capacity of the first brake B1 can be reduced.

As described above, according to the electric motor reduction gear 1 of the present embodiment, by switching the operation and stop of the first and second brakes B1 and B2, the power of the motor 2 can be shifted in two stages, and the power of the motor 2 can be shifted in two steps. The differential 4 distributes the power to the left and right drive shafts SL and SR. In addition, the power of the motor 2 is decelerated by the second planetary gear mechanism PG0, so that the entire transmission can obtain a larger reduction ratio.

In addition, the first brake B1 and the one-way clutch OWC are inside the extension P2E extending from the base P2B of the second pinion P2 which is the double pinion type first planetary gear mechanism PG1 to the opposite side of the motor 2 , and the first brake B1 and the one-way clutch OWC. A space is partitioned, and most of the differential 4 is accommodated in the inner space. As a result, the length in the axial direction can be shortened compared to the case where the differentials are arranged in the axial direction with respect to the transmission mechanism, and the entire device including the transmission 1 and the differential 4 can be downsized. In addition, since the differential gear 4 having a simpler structure than that of the double-pinion type planetary gear mechanism is used as the differential mechanism, simplification and cost reduction of the overall structure of the device can be achieved. Furthermore, since the first and second brakes B1 and B2 and the one-way clutch OWC are arranged on the peripheral wall portion of the casing CA, the axis line can be shortened compared to the conventional case where the planetary gear mechanism and the like are arranged in line in the axis direction. The direction length can further realize the miniaturization of the whole device.

Furthermore, the carrier C1 in the Low mode that requires a large braking torque is fixed by the one-way clutch OWC, and the carrier C1 in the RVS mode that does not require a large braking torque is performed by the first brake B1 Since C1 is fixed, the clutch capacity of the first brake B1 can be reduced.

Hereinafter, referring to FIG. 8 and the like, a description will be given of a transmission for an electric motor (hereinafter, referred to as a “transmission device” as appropriate) 51 according to the second embodiment of the present invention. The transmission device 51 is mainly different from the transmission device 1 of the first embodiment described above in that the second pinion P2 of the first planetary gear mechanism PG1 is constituted by a double pinion; One of the pinions meshes with the second ring gear R2 and the third brake B3 which brakes the second ring gear R2. In the description of the transmission device 51 below, the same reference numerals are given to the same or equivalent components as the transmission device 1 , and the description will center on the parts different from the transmission device 1 .

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

FIG. 9 shows an example of setting the number of teeth of the main rotating element in the present embodiment. As can be seen from the comparison with FIG. 3 , the number of teeth Z of the rotating elements such as the sun gear S0 used in the same manner as in the first embodiment is set to the same value as that in the first embodiment. In addition, the number of teeth ZP22 of the pinion 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 correspondingly, the number of teeth ZR2 of the second ring gear R2 is smaller than the number of teeth ZR1 of the ring gear R1.

3 and 8 , the first brake B1 and the one-way clutch OWC are arranged on the opposite side (differential 4 side) of the motor 2 with respect to the second brake B2 in the first embodiment, but are In this embodiment, it is arranged on the motor 2 side. As a result, a space longer and larger in the axial direction than in the first embodiment is demarcated between the extending portion P21E of the large gear 21 of the second pinion P2 and the inner side of the pinion P22 arranged on the extension side thereof. , the entire differential 4 is accommodated in the inner space.

In addition, the third stopper B3 is provided on the peripheral wall portion of the casing CA, and has an inner member 15a, an outer member 15b, and an actuator (not shown) that are configured similarly to the first and second brakes B1 and B2, and the inner member 15a is connected to an actuator (not shown). The second ring gear R2 is connected, and the outer member 15b is provided on the case CA so as to be movable in the axial direction and not rotatable. Therefore, in the state where the actuator is stopped, the second ring gear R2 is allowed to rotate without being braked, and when the actuator is activated, the second ring gear R2 is braked and fixed so as not to rotate.

In the speed change device 51 having the above configuration, the speed reduction operation by the second planetary gear mechanism PG0 and the speed change operation by the first planetary gear mechanism PG1 and the like of the speed change device 1 (Hi mode, Low mode, RVS mode) are performed in the same manner. and cutaway mode). In addition, since the number of teeth Z of the rotating elements of the first and second planetary gear mechanisms PG1, PG0, etc. is the same as that of the first embodiment, the S0-S1 reduction ratio RS01 and the Hi mode reduction ratio RS12 (Hi) are obtained by their operations. and the Low mode reduction ratio RS12 (Low) is also the same as that of the first embodiment.

In addition to the above-mentioned operations, in this transmission device 51, a shifting operation based on a forward intermediate speed (Mid) mode is performed as described below. First, since the second pinion P2 of the double pinion type first planetary gear mechanism PG1 is constituted by a double pinion, the pinion P22 meshes with the second ring gear R2, and the second ring gear R2 is connected to the third brake B3 , the relationship between the rotational speeds of the rotational elements of the transmission 51 can be represented by, for example, the collinear diagram of FIG. 11 . That is, the sun gear S1 (carrier C0), the ring gear R1, the second ring gear R2, the carrier C1, and the second sun gear S2 constitute five rotating elements, and the rotational speeds of these five rotating elements are arranged on a straight line The collinear relationship is established.

In addition, λ3 in the figure is the speed ratio between the carrier C1 and the second ring gear R2 (hereinafter referred to as the "third lever ratio") based on the distance between the carrier C1 and the ring gear R1, and can be obtained by the following formula ( 6) Performance.

λ3=(ZR1/ZR2)·(ZP22/ZP21)...(6)

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

From the relationship between the first to third lever ratios λ1 to λ3 shown in FIG. 11 , the reduction ratio between the sun gear S1 and the second sun gear S2 in this Mid mode can be expressed by the following equation (7) (hereinafter referred to as the "Mid mode reduction ratio") RS12 (Mid).

RS12(Mid)

=(λ1-λ3)/(λ2+λ3)

=[(ZR1/ZS1)-(ZR1/ZR2)·(ZP22/ZP21)]

/[(ZR1/ZS2)+(ZR1/ZR2)·(ZP22/ZP21)]...(7)

If ZR1, ZS1, ZS2, ZP21 and ZP22 on the right are substituted into the values in Figure 9, the Mid mode reduction ratio RS12 (Mid) is:

RS12(Mid)

=[(147/36)-(147/141)·(20/26)]

/[(147/93)+(147/141)·(20/26)]=1.377,

The reduction ratio between the reduction ratio RS12 (Hi) (=1.195) in the Hi mode and the reduction ratio RS12 (Low) (=2.583) in the Low mode can be obtained.

In addition, the total reduction ratio between the motor 2 and the differential 4 in the mid mode can be determined by the S0-S1 reduction ratio RS01 (=5.083) of the formula (1) and the mid mode reduction ratio RS12 (Mid) (=1.377) The product is expressed as 5.083×1.377=6.999.

Summarizing the operation of the above-mentioned transmission 51 , as shown in FIG. 12 , by controlling the operation of the motor 2 and the operation and stop of the first to third brakes B1 to B3 , the Low mode, the Mid mode and the Hi mode are executed. 3-step forward shifting of the mode, and operation based on the RVS mode and the cut-off mode.

As described above, according to the electric motor reduction gear 51 of the present embodiment, by switching the operation and stop of the first to third brakes B1 to B3, the power of the motor 2 can be shifted in three stages, and the power of the motor 2 can be shifted by the difference The transmission 4 distributes the power to the left and right drive shafts SL and SR and the drive wheels.

In addition, the second pinion P2 of the first planetary gear mechanism PG1 is composed of a double pinion integrally including a large gear P21 and a pinion P22, and a pinion extending toward the extension side through an extension P2E of the large gear P21 and the pinion. P22 defines a larger inner space that is longer in the axial direction than that of the first embodiment. As a result, the entire differential 4 can be accommodated in the inner space. Furthermore, like the first and second brakes B1 and B2, since the third brake B3 is also arranged on the peripheral wall portion of the casing CA, the length in the axial direction can be shortened, and the size of the entire device can be reduced.

In addition, the present invention is not limited to the described embodiment, and can be implemented in various forms. For example, the differential gear 4 is not limited to the general structure described in the embodiments, and other arbitrary types of structures may be employed. In addition, in the embodiment, the multi-plate type friction clutches are used as the first to third brakes B1 to B3, but other suitable types of brakes may be used as long as the target rotating element can be braked.

In addition, in the embodiment, in order to reduce the clutch capacity of the first brake B1, the one-way clutch OWC is provided separately, and is used for fixing the carrier C1 in the Low mode. However, the one-way clutch OWC may be omitted and the one-way clutch OWC may be omitted by The first brake B1 fixes the carrier C1.

Furthermore, the embodiment is an example in which the transmissions 1 and 51 are used in the motor 2 as a power source of a vehicle, however, the present invention is not limited to this, and can be widely used in electric motors for other purposes used with a differential mechanism. In addition, the structures of specific parts can be appropriately changed within the scope of the gist of the present invention.

Claims (5)

1. a kind of motor speed change gear carries out speed change to the power of motor, and passes through differential attachment for the power 2 output shafts are distributed to, which is characterized in that,

The motor has the planetary gear mechanism with the double-pinion type of the output shaft arranged coaxial with speed change gear,

The planetary gear mechanism includes

Sun gear, the power of the motor are input into the sun gear；

Tooth rest engages more respectively by multiple 1st pinion gears engaged with the sun gear and with multiple 1st pinion gear A 2nd pinion gear rotatably supports；And

Gear ring is engaged with the multiple 2nd pinion gear,

The motor is also included with speed change gear

1st brake is used to brake the tooth rest；And

2nd brake is used to brake the gear ring,

The multiple 2nd pinion gear respectively includes the base portion engaged with the 1st pinion gear and the gear ring；And from the base The extension portion that portion extends to the opposite side of the motor,

The differential attachment is made of differential mechanism, which has the 2nd engaged with the extension portion of the 2nd pinion gear Sun gear is as input gear, and the differential attachment is configured in by the extension portion of the multiple 2nd pinion gear In the inner space marked off.

2. motor speed change gear according to claim 1, which is characterized in that

The multiple each free duplex pinion gear of 2nd pinion gear is constituted, which is provided integrally with: gear wheel, tool There are the base portion and the extension portion；And pinion gear, the number of teeth of the gear ratio gear wheel is few,

The motor is also included with speed change gear

2nd gear ring is engaged with the pinion gear；And

3rd brake is used to brake the 2nd gear ring.

3. motor speed change gear according to claim 2, which is characterized in that

The pinion gear of 2nd pinion gear configures the opposite side in the motor, the difference relative to the gear wheel Fast device configuration is in the inner space marked off by the extension portion and the pinion gear of the gear wheel.

4. motor speed change gear according to claim 2 or 3, which is characterized in that

The motor, the planetary gear mechanism and the differential mechanism are incorporated in shell, the 1st brake, described 2nd brake and the 3rd brake are configured at the peripheral wall portion of the shell.

5. according to claim 1 to motor speed change gear described in any one in 4, which is characterized in that

The motor speed change gear also has the 2nd planetary gear mechanism of single pinion type, the 2nd planetary gear mechanism with The output shaft is coaxially arranged between the motor and the planetary gear mechanism,

2nd planetary gear mechanism includes

Sun gear links with the motor；

Not revolvable gear ring；And

Tooth rest rotatably supports the pinion gear engaged with the sun gear and the gear ring, and with the row The sun gear of star gear mechanism links.