JP4297447B2 - Hybrid drive device - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a hybrid driving device adopting a two-motor split system capable of securing good accelerating performance in the low speed range and assuring high fuel consumption in the high speed range. <P>SOLUTION: The hybrid driving device consists of a planetary gear equipped with an input shaft I coupled with and driven by an engine E, an output shaft O coupled with wheels for driving them, and a first motor MG1 and a second motor MG2 so as to constitute a first, a second, and a third rotating element in the sequence of the rotating speeds, and further a planetary gear PGO for power distribution is provided in which the first motor MG1 is connected with the first rotating element, while the input shaft I is connected with the second rotating element, and an intermediate shaft M2 and the second motor MG2 are connected with the third rotating element. Therein a transmission SC is provided having a plurality of shift ranges for forwarding the power transmitted from the intermediate shaft M2 to the output shaft O, wherein the shift ranges of the transmission SC include at least an accelerating range and a decelerating range. <P>COPYRIGHT: (C)2008,JPO&amp;INPIT

Description

  The present invention includes an input shaft that is drivingly connected to an engine and an output shaft that is drivingly connected to wheels, and includes a first electric motor and a second electric motor, and the first, second, and third rotating elements are arranged in order of rotational speed. A planetary gear that is configured, wherein the first electric motor is connected to the first rotating element, the input shaft is connected to the second rotating element, and an intermediate transmission shaft and the second electric motor are connected to the third rotating element. The present invention relates to a hybrid drive device including a planetary gear for power distribution and a transmission that has a plurality of shift stages and outputs power transmitted from the intermediate transmission shaft to the output shaft.

  As disclosed in Patent Document 1, this type of hybrid drive device includes a first electric motor, a power distribution planetary gear, a first power transmission system that is arranged on the engine side from the input shaft side to the output shaft side. Two electric motors and a transmission are provided in this order.

  In the hybrid drive device having this configuration, the torque that becomes excessive or insufficient on the output side can be compensated for by the second electric motor while maintaining the rotational speed of the engine at the highest fuel efficiency by the action of the first electric motor. .

  In Patent Document 1, a power transmission efficiency is achieved by interposing a transmission having three speeds between a second rotating shaft that rotates integrally with the second electric motor and a drive shaft for outputting power. Has proposed a hybrid drive device that can perform a shift without reducing the speed.

Patent Document 2 discloses a differential that selectively switches between an operating state in which the power distribution mechanism can be operated as an electric uninterrupted transmission (uninterrupted shifting state) and a constant shifting state in which it can be operated as a transmission having a constant gear ratio. It has been proposed to provide a switching clutch C0 and a switching brake B0 as state switching devices. As a result, the hybrid drive device is reduced in size or improved in fuel efficiency.
In this patent document, a transmission capable of realizing four to eight forward speeds and further realizing one or two reverse speeds is employed.

JP 2005-61498 A JP 2005-206136 A

  In the technique disclosed in Patent Document 1, the drive after power distribution using a differential gear mechanism is received as a transmission input and the output after the shift is transmitted to the drive wheels. However, the shift speed is limited to three. At the same time, the highest speed stage can only be realized in a directly connected state where the rotational speed of the output shaft is the same as the rotational speed of the input shaft in the transmission. There is room for improvement.

  With the technique disclosed in Patent Document 2, an output in a decelerated state (underdrive state) can be obtained with respect to the engine rotational speed, and the acceleration performance can be relatively improved in a low speed region requiring traction force. In addition, by locking one gear element of the power distribution mechanism that distributes engine power, the rotational speed of the rotary shaft input to the transmission is increased with respect to the rotational speed of the engine. By transmitting the rotational speed to the transmission output shaft at the same speed, the energy conversion loss is suppressed, and the fuel efficiency in the high speed range can be improved. However, in the technique disclosed in this document, the input rotation speed cannot be increased on the transmission side, and the engine rotation speed cannot be continuously changed in this state, so that the engine is operated in the optimum fuel consumption line. It becomes impossible and fuel consumption is bad. Therefore, one rotating element of the planetary gear is connected to the output shaft of the engine, the other two rotating elements of the planetary gear are connected to the first motor and the transmission shaft connected to the drive wheel, and further to the transmission shaft. There is room for improvement in that a so-called two-motor split system in which two motors are connected, that is, a power distribution mechanism that can operate as an electric uninterrupted transmission can be fully utilized.

  The present invention has been made in view of the above-mentioned problems, and the object thereof is a hybrid drive device that employs a two-motor split system, which can ensure acceleration performance in a low speed region and also has high fuel efficiency in a high speed region. And a hybrid drive device that can be used.

In order to achieve the above object, the first feature configuration of the present application according to the present invention is as follows:
An input shaft that is drivingly connected to the engine and an output shaft that is drivingly connected to the wheels.
A first electric motor and a second electric motor,
Planetary gears constituting first, second, and third rotating elements in order of rotational speed, wherein the first motor is connected to the first rotating element, the input shaft is connected to the second rotating element, An intermediate transmission shaft and the second electric motor are connected to the third rotating element, and a power distribution planetary gear that can transmit the rotation of the input shaft to the intermediate transmission shaft is accelerated.
A transmission having a plurality of shift stages and outputting power transmitted from the intermediate transmission shaft to the output shaft;
The plurality of shift stages of the transmission at least increase the rotational speed of the intermediate transmission shaft and output it to the output shaft; reduce the rotational speed of the intermediate transmission shaft and output the output It has a reduction gear stage that outputs to the shaft,
The transmission includes a first brake, a first clutch and a second clutch that transmit rotation of the intermediate transmission shaft, and first, second, third, and fourth rotating elements in order of rotational speed. Has a planetary gear set,
The first rotating element is selectively fixed to the case by the first brake;
The second rotating element can selectively transmit rotation of the intermediate transmission shaft by the second clutch, and the third rotating element can output output rotation to the output shaft,
The rotation of the intermediate transmission shaft is selectively transmitted by the first clutch, and the first clutch and the second clutch are arranged between the planetary gear set for speed change and the second electric motor;
The output-side transmission member of the first clutch passes through the inner diameter side of the output-side transmission member of the second clutch, and is connected to the fourth rotation element constituting the transmission planetary gear set ,
The first first clutch hub as an output member of the clutch is in Rukoto are located on the inner diameter side of the second clutch hub as an output member of the second clutch.
The “connection” includes a structure capable of directly exchanging driving, and also includes a structure capable of performing indirectly through a plurality of members.
“Transmission” includes a state where rotation is directly transmitted as it is and a state where rotation is indirectly transmitted by increasing or decreasing speed.
“Fixed to the case” includes a state in which the brake is directly fixed from the case and a state in which the brake is indirectly fixed to the case via a member fixed to the case.

  In this hybrid drive device, the transmission is provided with an acceleration gear stage and a deceleration gear stage while adopting a two-motor split system. Therefore, it is possible to accurately cope with acceleration in the low speed region at the deceleration gear, and to cope with improvement in fuel consumption in the high speed region. In addition, because the transmission can handle a wide range of speeds, while using a compact electric motor, it is possible to take full advantage of the two-motor split system while improving driving efficiency at high speeds and large driving force at low speeds. Can do. Also, by using the planetary gear for power distribution, the driving force of the engine is increased, the rotational speed of the engine is increased and input to the transmission, and the rotational speed increased by the transmission is further increased. , Improve energy efficiency and achieve good high speed fuel efficiency.

  In the hybrid drive device having this configuration, the speed increase described above is achieved by operating the second clutch and the first brake to transmit the rotation of the intermediate transmission shaft to the second rotation element and stopping the rotation of the first rotation element. A gear stage is realized, and the rotational speed that is distributed from the third rotating element to the output shaft and input to the transmission can be output in an accelerated state. On the other hand, the first clutch and the first brake are operated to transmit the rotation of the intermediate transmission shaft to the fourth rotation element, and the rotation of the first rotation element is stopped, thereby realizing the reduction gear described above, The rotational speed that is distributed from the rotating element to the output shaft and input to the transmission can be output in a decelerated state.

  Furthermore, in the driving state in which the speed increasing state is realized, the fourth rotating element is forced to rotate at a high speed, but the output side transmission member of the first clutch has a larger inner diameter side than the output side transmission member of the second clutch. Since it is connected to the fourth rotation element through, the influence of the high-speed rotation can be suppressed. As a result, for example, it is possible to prevent the first clutch from being deformed by centrifugal force and the input / output members of the clutch to come into contact with each other to reduce the running performance.

  Further, by storing the first clutch and the second clutch between the planetary gear set for speed change and the second electric motor, the weight parts such as the second electric motor and the planetary gear set are arranged on both sides with this portion interposed therebetween, and the compact and A balanced hybrid drive device can be obtained.

Here, it is preferable that the first electric motor, the power distribution planetary gear, the second electric motor, and the transmission planetary gear set are arranged coaxially in order from the input shaft side.
In the hybrid drive device having the above-described first characteristic configuration of the present application,
The transmission further includes a second brake,
It is preferable that the second rotating element constituting the transmission planetary gear set is selectively fixed to the case by the second brake.

  By adopting this configuration, it is possible to realize a further shift stage at which the first clutch and the second clutch substantially work.

In the hybrid drive apparatus described so far,
The first clutch and the second clutch are arranged on the inner side of the outer surface in the radial direction of the electric motor,
It is preferable that at least a part of the first clutch and the second clutch are arranged so as to overlap in the axial direction.

  The electric motor that can be employed in the hybrid drive device is of a size having a predetermined width. By maintaining the first and second clutches in the above positional relationship, the size of the hybrid drive device in the radial direction is maximized. The outer diameter of the electric motor can be accommodated, and the axial length of the hybrid drive device can be shortened by overlapping the first clutch and the second clutch in the axial direction.

  In this configuration, if the first clutch is arranged on the inner diameter side of the second clutch, the configuration can be made compact in the radial direction while shortening the axial length.

In the configuration including the second brake described above,
The transmission further includes a third clutch,
It is preferable that the rotation of the intermediate transmission shaft is selectively transmitted to the first rotation element constituting the transmission planetary gear set by the third clutch.
In this configuration, the rotation of the intermediate transmission shaft is transmitted to the first rotation element by the third clutch, and the rotation of the second rotation element is stopped by the second brake, whereby the intermediate transmission is transmitted to the third rotation element (output shaft). The driving force and rotation in the opposite direction to the shaft can be output, and the reverse can be realized by mechanical shift. As a result, a driving force can be obtained in the reverse stage, and the reverse can be stably performed for a long time. Further, it is possible to avoid a large amount of heat generation on the motor side as the vehicle moves backward.

When adopting a configuration including such a third clutch,
The third clutch is disposed closer to the electric motor than the planetary gear set for shifting, and is disposed on the inner side of the outer radial surface of the electric motor.
The third clutch is preferably arranged so as to be overlapped in the axial direction with respect to one or both of the first clutch and the second clutch.
By adopting this configuration, the radial width and the axial length of the hybrid drive device can be suitably limited, and a compact hybrid drive device can be obtained.
Here, the first clutch is disposed on an inner diameter side from both the second clutch and the third clutch, and at least a part of the first clutch is a shaft with respect to one or both of the second clutch and the third clutch. It is preferable that the third clutch is disposed so as to overlap in the direction, and at least a part of the third clutch is disposed so as to overlap in the radial direction with respect to the second clutch.
Furthermore, it is preferable that at least a part of the first brake is arranged so as to overlap in a radial direction with respect to one or both of the second clutch and the third clutch.
Also in the configuration including the third clutch, the output-side transmission member of the first clutch passes through the inner diameter side of the output-side transmission member of the third clutch, and constitutes the transmission planetary gear set. It is preferably connected to a four-turn element.
Even if it is in the relationship with the output side transmission member of the third clutch, by arranging the output side transmission member of the first clutch on the inner diameter side, this member that may become high speed is affected by the high speed rotation, Problems can be avoided.

  As the transmission hybrid gear set described so far, the following configuration can be adopted when the transmission hybrid gear set is configured as a combination of two planetary gears.

The planetary gear set for speed change, constitutes a sun gear, a combination of the first planetary gear and the second planetary gears of the 3 elements with carrier and a ring gear,
A sun gear of the first planetary gear is selectively fixed to the case by a first brake;
The carrier of the first planetary gear is connected to the ring gear of the second planetary gear, and the rotation of the intermediate transmission shaft is selectively transmitted by the second clutch,
A ring gear of the first planetary gear is coupled to the carrier of the second planetary gear and is drivingly coupled to the output shaft;
The sun gear of the second planetary gear is configured to selectively transmit the rotation of the intermediate transmission shaft by the first clutch.
In this way, a hybrid drive device having an increase gear stage and a reduction gear stage can be realized by a combination of three simple planetary gears .

Now , as the second feature configuration ,
Configuration and outputting pre Symbol first brake, second brake, to the output shaft by setting the engagement and disengagement states of the first clutch and the second clutch rotation speed of the intermediate transmission shaft in four different gear ratio To do.
With this configuration, the operation of the four friction engagement elements is controlled to realize four gear ratios including the speed-up speed stage and the speed-down speed stage, while taking advantage of the two-motor split system, It is possible to realize a configuration including a transmission that performs an automatic shift, and to reduce the size of the electric motor. Furthermore, the entire planetary gear set that constitutes the transmission can be constituted by two planetary gears, and the entire hybrid drive device that combines the size of the transmission itself required for miniaturization and miniaturization of the motor by providing the transmission. Can be made compact. Furthermore, it is possible to satisfactorily cope with a relatively wide speed range and to improve the running fuel consumption.

  As described above, the speed increase described above is achieved by operating the second clutch and the first brake, transmitting the rotation of the intermediate transmission shaft to the second rotating element, and stopping the rotation of the first rotating element. A gear stage is realized, and the rotational speed that is distributed from the third rotating element to the output shaft and input to the transmission can be output in an accelerated state. On the other hand, the first clutch and the first brake are operated to transmit the rotation of the intermediate transmission shaft to the fourth rotation element, and the rotation of the first rotation element is stopped, thereby realizing the reduction gear described above, The rotational speed that is distributed from the rotating element to the output shaft and input to the transmission can be output in a decelerated state. Furthermore, it is possible to realize a further shift stage in which the first clutch and the second brake work.

As the transmission hybrid gear set having the second characteristic configuration, the following configuration can be adopted to configure the transmission hybrid gear set substantially as a combination of two planetary gears .
The pre-Symbol shifting planetary gear set, and configured to include a sun gear, a first planetary gear and the second planetary gears of the 3 elements with carrier and a ring gear,
The sun gear of the first planetary gear is selectively fixed to the case by the first brake,
The carrier of the first planetary gear is connected to the ring gear of the second planetary gear, the rotation of the intermediate transmission shaft is selectively transmitted by the second clutch, and is selectively fixed to the case by the second brake,
A ring gear of the first planetary gear is coupled to the carrier of the second planetary gear and is drivingly coupled to the output shaft;
The sun gear of the second planetary gear is selectively transmitted with the rotation of the intermediate transmission shaft by the first clutch.
By doing so, it is possible to realize a hybrid drive device that is capable of four-speed shifting by including a speed-up speed stage and a speed-down speed stage by combining three simple planetary gears.

In the hybrid drive device having the second characteristic configuration of the present application described so far,
The transmission further includes a third clutch;
It is preferable that the first rotation element constituting the speed-change planetary gear set adopts a configuration in which the rotation of the intermediate transmission shaft is selectively transmitted by the third clutch.
In this way, the rotation of the intermediate transmission shaft is transmitted to the first rotation element by the third clutch, and the rotation of the second rotation element is stopped by the second brake, so that the third rotation element (output shaft) is transmitted. Driving force and rotation in the opposite direction to the intermediate transmission shaft can be output, and the reverse can be realized by mechanical shift. A driving force can be obtained in the reverse stage, and the reverse can be stably performed for a long time. Further, it is possible to avoid a large amount of heat generation on the motor side as the vehicle moves backward.

Regarding the first clutch that has been described so far, the first clutch is engaged, and in the transmission state in which the rotation of the intermediate transmission shaft is transmitted to the fourth rotation element constituting the transmission planetary gear set, in the forward gear stage. It is preferable that the gear stage having the largest gear ratio is formed.
This configuration defines the gear stage in which the first clutch is involved, but the clutch (first clutch) involved in the gear stage having the largest gear ratio (reduction ratio) that requires the largest driving force is relatively It tends to be large.
However, in the hybrid drive device having the first and second characteristic configurations, a balanced arrangement can be realized by disposing the output side transmission member on the inner diameter side.
Furthermore, in the hybrid drive device having the second characteristic configuration, four gear ratios including the speed increasing speed stage and the speed reducing speed stage are realized by controlling the operation of the four friction engagement elements, and the two-motor split system With this advantage, it is possible to sufficiently realize a configuration including a transmission that performs mechanical shifting.

Embodiments of the present invention will be described below with reference to the drawings.
The hybrid drive device HE includes an input shaft I that receives drive force from the engine E, and an output shaft O that outputs drive force to drive wheels (not shown), and is shown in FIG. In the hybrid drive device HE, an input shaft I and an output shaft O are provided along the axial direction of these shafts. Of course, the present invention is not limited to the configuration including these axes along the axial direction, and the input shaft and the output shaft may be arranged in parallel and are arbitrary.

  As shown in FIG. 1, a motor part MP in which electric motors MG1 and MG2 and a power distribution planetary gear PG0 are provided is provided on the input shaft I side, and a transmission part SP in which a transmission SC is provided is provided as an output shaft. It is provided on the O side. A partition wall SW is provided at a boundary portion between the motor unit MP and the transmission unit SP.

The hybrid drive apparatus HE includes a first electric motor MG1, a power distribution planetary gear PG0, and a second electric motor MG2 from the input shaft I toward the output shaft O, and further includes a transmission SC on the output shaft O side. ing.
These devices are generally configured to be housed in a mission case MC that decreases in radial direction width from the input shaft I side toward the output shaft O side.

  Between the input shaft I and the output shaft O, the 1st intermediate shaft M1 and the 2nd intermediate shaft M2 are arrange | positioned coaxially with these shafts. The input shaft I and the first intermediate shaft M1 are connected via a damper D, and fluctuations in engine output are absorbed by the damper D and transmitted to the first intermediate shaft M1. The first intermediate shaft M1 and the second intermediate shaft M2 are connected via a power distribution planetary gear PG0 so as to be capable of driving transmission. The output shaft O is disposed coaxially with these intermediate shafts M1 and M2.

  In the present application, the transmission SC performs the function of shifting the transmission rotational speed transmitted to the second intermediate shaft M2 to the output shaft O or transmitting it at the same speed. Therefore, the input element of the transmission SC is the second intermediate shaft M2, and the output element is the output shaft O. The input element of the transmission SC is called “intermediate transmission shaft”. The transmission SC includes both a deceleration output stage that decelerates and outputs the rotation speed of the input element and a speed-up shift stage that outputs after increasing the speed.

  In the present application, a plurality of embodiments will be described, but the structure on the drive transmission upstream side (on the engine E side and on the left side of the partition wall SW in FIG. 1 and the like) from the transmission SC is common. Therefore, in the following description, first, the structure on the upstream side of the transmission SC will be described, and the structure after the transmission including the transmission SC will be described later.

Drive transmission upstream side structure from the transmission The drive transmission from the first intermediate shaft M1 to the second intermediate shaft M2 proceeds from the engine E side to the transmission SC side, the first electric motor MG1, the power distribution planetary gear PG0, Two electric motors MG2 are provided.

  Both of these electric motors MG1 and MG2 can perform a function as a motor that generates power by receiving power supply and a function as a generator that generates power by receiving power supply. The operation is based on a control command from the electronic control unit ECU. Each of the electric motors MG1 and MG2 is electrically connected to the battery B via the inverter In, and receives the supply of electric power to act as a motor, and stores the electric power generated by acting as a generator in the battery B. Alternatively, it can be sent to another motor to drive it.

  Further, the electronic control unit ECU is configured to control an operating state of engagement / disengagement of a friction engagement element to be described later. For example, the electronic control unit ECU requests the vehicle speed and the vehicle according to a pre-stored shift map. Based on the driving force (for example, the amount of depression of the accelerator), the friction engagement element is controlled so as to realize a suitable shift stage with the transmission SC.

  The basic control mode of the two electric motors MG1 and MG2 by the electronic control unit ECU will be briefly described. For the first electric motor MG1, the engine E is operated with the highest fuel consumption (the rotational speed at which the engine efficiency can be maximized). In a state where the engine is operated in the above state, the rotational speed of the electric motor MG1 is controlled to maintain the engine rotation. For second motor MG2, the output torque is controlled to satisfy the required torque required for the vehicle.

  In the illustrated example, a single planetary gear provided with a sun gear s0, a carrier ca0 that rotatably supports the pinion, and a ring gear r0 is employed as the power distribution planetary gear PG0. The power distribution planetary gear PG0 is connected so that the first intermediate shaft M1 rotates integrally with the carrier ca0, and the rotor of the first electric motor MG1 is connected so as to rotate integrally with the sun gear s0. The ring gear r0 is connected to rotate integrally with the second intermediate shaft M2. The second intermediate shaft M2 is configured to rotate integrally with the rotor of the second electric motor MG2.

  A speed diagram of the power distribution planetary gear PG0 in the case of adopting this connection structure is shown on the left side of FIG. The drawing on the right side of FIG. 5 is a diagram corresponding to the operation of the planetary gear set for shifting PGS1 provided in the first embodiment described later.

  In the velocity diagram, the vertical axis corresponds to the rotational speed of each rotating element. “0” described corresponding to the vertical axis indicates that the rotational speed is 0, with the upper side being positive and the lower side being negative. “1”, “2”, “−1”, and the like indicate the ratio of the planetary gear or the planetary gear set to be described to the input rotation speed. Each vertical line arranged in the horizontal direction corresponds to the rotating element of the planetary gear or planetary gear set described. In the velocity diagram, an inclined thick line or a horizontal thick line extending over a plurality of vertical lines indicates one operation state of the planetary gear (or the planetary gear set).

Returning to the speed diagram shown on the left side of FIG. 5, the vertical line (supplied with “S0”) located at the left end in FIG. 5 corresponds to the first rotation element (rm1) that is the sun gear s0. The rotating element rotates integrally with the rotor of the first electric motor MG1.
A vertical line (supplied as “Ca0”) located at the center corresponds to the second rotating element (rm2) that is the carrier ca0, and this rotating element receives the driving rotation from the engine E, and the first intermediate shaft M1. And rotate together. “Er” written next to “1” on the vertical axis in the figure indicates that this is the engine speed.
The vertical line on the right side (supplied with “R0”) corresponds to the third rotating element (rm3) which is the ring gear r0, and this rotating element is connected to the second intermediate shaft M2 and the second intermediate shaft M2. It rotates integrally with the rotor of the electric motor MG2.
In the case of this power distribution planetary gear PG0, the second rotation element rm2 is an input element, and the third rotation element rm3 is an output element. The second intermediate shaft M2 becomes the “intermediate transmission shaft” described so far.

In the illustrated operating state, the inclined thick line is rising to the right, but the sun gear s0 (first rotation element rm1) is determined by the rotation speed (corresponding to the engine rotation speed Er) of the carrier ca0 that is the second rotation element rm2. The side is low speed (0 in the example shown), and the ring gear r0 (third rotation element rm3) side is high speed. That is, it indicates an overdrive state.
In this operation state, the first electric motor MG1 receives a reaction force against engine driving, and the driving force from the engine E is distributed and transmitted to the first electric motor MG1 and the second intermediate shaft M2. At this time, the first electric motor MG1 functions as a generator. On the other hand, the remaining intermediate power is transmitted to the second intermediate shaft M2 in an overdrive state (a state in which the second intermediate shaft M2 rotates faster than the rotational speed of the engine E). In the second intermediate shaft M2, the remaining engine drive and the driving force added or reduced by the second electric motor MG2 are transmitted to the transmission SC.

In this application, the seventh embodiment is introduced as a configuration of the transmission SC.
The relationship between each embodiment will be described. The first to fourth and seventh embodiments are modes for realizing forward four speeds and reverse one speed, and the fifth embodiment is for realizing six forward speeds and one reverse speed, The sixth embodiment achieves eight forward speeds and two reverse speeds.

First Embodiment FIG. 2 shows a specific configuration after the transmission of this embodiment, and FIG. 3 shows a skeleton diagram.
As can be seen from these drawings, the transmission SC is configured to include a planetary gear set PGS1 including a pair of single planetary gears PG1 and PG2, and a plurality of transmission elements SC corresponding to the rotational elements constituting the planetary gear set PGS1. Friction engagement elements C1, C2, C3, B1, B2, F1 are provided.

  As friction engagement elements, a first clutch C1, a second clutch C2, a third clutch C3, a first brake B1 and a second brake B2 are provided, and a one-way clutch F1 is provided. FIG. 4 shows an operation table of these friction engagement elements.

FIG. 3 is a drive transmission skeleton diagram, FIG. 4 is an operation table of a friction engagement element, FIG. 5 is a speed diagram, and FIG. 6 is a drive force diagram at each shift stage. The following will be described based on FIG.

  When the pair of planetary gears constituting the planetary gear set PGS1 is called the first planetary gear PG1 and the second planetary gear PG2 from the second intermediate shaft M2 that is the input member of the transmission SC toward the output shaft O, it is also found from the skeleton diagram. Thus, these planetary gears PG1 and PG2 are three-element single planetary gears including sun gears s1 and s2, carriers ca1 and ca2, and ring gears r1 and r2, respectively.

  The relationship with the friction engagement element will be described. The sun gear s1 of the first planetary gear PG1 is selectively fixed to the transmission case MC by the first brake B1, and the carrier ca1 of the first planetary gear PG1 is connected to the ring gear r2 of the second planetary gear PG2. A structure is employed in which the rotation of the second intermediate shaft M2 is selectively transmitted by the second clutch C2 and is selectively fixed to the transmission case MC by the second brake B2. The reverse rotation of the carrier ca1 is stopped by the one-way clutch F1.

  Further, the ring gear r1 of the first planetary gear PG1 is connected to the carrier ca2 of the second planetary gear PG2, and is also drivingly connected to the output shaft O. The sun gear s2 of the second planetary gear PG2 is connected to the second intermediate shaft M2 by the first clutch C1. A configuration in which rotation is selectively transmitted is employed.

  Further, the sun gear s1 of the first planetary gear PG1 is configured to selectively transmit the rotation of the second intermediate shaft M2 by the third clutch C3.

  The speed diagram of the transmission SC is shown in the right diagram of FIG. Therefore, from this speed diagram, the rotating member rotates at a speed of about −1 to 2 times the engine speed that may be input to the transmission SC in an overdrive state as the gear shifts. It can be seen that some elements exist.

  What is described on the upper side of the velocity diagram is the correspondence with the vertical lines of the rotating elements constituting each planetary gear. “R1, Ca1, S1” indicate the ring gear r1, carrier ca1, and sun gear s1 of the first planetary gear PG1, respectively. “R2, Ca2, S2” indicate the ring gear r2 and carrier ca2 of the second planetary gear PG2, respectively. The sun gear s2 is shown.

  In the figure, the correspondence between the vertical line and the rotating element will be described. In order from the right to the left, the rotating element (first rotating element rm1) provided with the sun gear s1 of the first planetary gear PG1 and the first planetary gear PG1. A rotation element (second rotation element rm2) configured to rotate integrally with the carrier ca1 and provided with the ring gear r2 of the second planetary gear PG2, and a ring gear r1 of the first planetary gear PG1 are provided and the carrier of the second planetary gear PG2. A rotating element (third rotating element rm3) that rotates integrally with ca2 and a rotating element (fourth rotating element rm4) provided with the sun gear s2 of the second planetary gear PG2 are positioned. Accordingly, in this embodiment, the planetary gear set PGS1 corresponds to a speed change planetary gear set.

  Since the third rotation element rm3 is integrated with the output shaft O, the rotational speed after the speed change by the transmission SC is indicated on the third vertical line from the right attached to the output “「 ”Indicates the rotational speed in that state. The velocity diagrams described below also follow this description format.

  “1ST”, “2ND”, “3RD”, “4TH”, and “REV” are added to “◯” described in the speed diagram to indicate the correspondence with each shift state. Here, “1ST” indicates the first forward speed, “2ND” indicates the second forward speed, “3RD” indicates the third forward speed, and “4TH” indicates the fourth forward speed. Further, “REV” indicates a reverse stage. The speed diagram described below also follows this description format. Incidentally, in FIG. 15 or FIG. 18, “5 to 8 TH” indicates forward 5 to 8 speeds, respectively. Further, “REV1” and “REV2” indicate that there are one reverse speed and two reverse speeds.

  In the planetary gear set PGS1, the rotating elements are ordered according to the direction from the low speed side to the high speed side in the case of forward travel, and from the high speed side to the low speed side in the case of reverse travel. In either case, the order of the rotating elements does not change.

The relationship between each rotation element of the transmission SC and each friction engagement element in this embodiment has the following structure as can be seen from FIG.
The first rotating element rm1 is selectively fixed to the mission case MC by the first brake B1. Further, the rotation of the second intermediate shaft M2 is selectively transmitted to the first rotation element rm1 by the third clutch C3.
The second rotation element rm2 is selectively transmitted to the transmission case MC by the second brake B2 while the rotation of the second intermediate shaft M2 is selectively transmitted by the second clutch C2. The third rotation element rm3 can output the output rotation to the output shaft O. The rotation of the second intermediate shaft M2 is selectively transmitted to the fourth rotation element rm4 by the first clutch C1.

FIG. 4 is an operation table showing the states of the friction engagement elements C1, C2, C3, B1, B2, and F1 at the respective shift speeds. In the operation table, “◯” indicates that the friction engagement element is in an engaged state, and “No mark” indicates that the friction engagement element is in a non-engagement state. Further, “(◯)” indicates that the same state as the engaged state is realized by the frictional engagement element B2 described above by the operation of the one-way clutch F1.
As shown in the table, in the transmission SC according to the present application, any two friction engagement elements are maintained in an engaged state at each shift stage, and the remaining friction engagement elements are disengaged. , Each gear stage is realized.

Shifting The shifting state in this embodiment in which the above configuration is adopted will be described.
FIG. 6 shows the drive transmission with bold lines on the skeleton diagram corresponding to FIG. From the upper side of the figure, drive transmission states at the first forward speed, the second forward speed, the third forward speed, the fourth forward speed and the reverse speed are shown.

1 Forward 1st (1ST)
At this shift stage, as shown in the operation table, only the first clutch C1 and the one-way clutch F1 are engaged. As shown in the uppermost stage, the rotation of the second intermediate shaft M2 is transmitted to the third intermediate shaft M3 by engaging the first clutch C1. In the second planetary gear PG2, when the reverse rotation of the ring gear r2 is stopped by the one-way clutch F1, the input rotation from the sun gear s2 is decelerated and output as the rotation of the carrier ca2.

2 Forward 2nd speed (2ND)
At this gear position, as shown in the operation table, only the first clutch C1 and the first brake B1 are engaged. As shown in the second stage from the top, when the first clutch C1 is engaged, the rotation of the second intermediate shaft M2 is transmitted to the third intermediate shaft M3. On the other hand, with respect to the planetary gear set PGS1 including the first planetary gear PG1 and the second planetary gear PG2, when the rotation of the sun gear s1 of the first planetary gear PG1 is stopped, the input rotation from the sun gear s2 is reduced, and the rotation of the carrier ca2 Is output as The reduction ratio in this state is smaller than the first forward speed.

3 Forward third speed (3RD)
At this shift stage, as shown in the operation table, only the first clutch C1 and the second clutch C2 are engaged. As shown in the third stage from the top, when the first clutch C1 is engaged, the rotation of the second intermediate shaft M2 is transmitted to the third intermediate shaft M3. On the other hand, since the rotational speeds of the sun gear s2 and the ring gear r2 are the same for the second planetary gear PG2 due to the engagement of the second clutch C2, the planetary gear PG2 is fixed. As a result, the rotation of the second intermediate shaft M2 is transmitted to the output shaft O as it is, so-called a directly connected state.

4 forward 4 speed (4TH)
At this shift stage, as shown in the operation table, only the second clutch C2 and the first brake B1 are engaged. As shown in the second stage from the bottom, when the second clutch C2 is engaged, the rotation of the second intermediate shaft M2 is transmitted to the carrier ca1 of the first planetary gear PG1. In the first planetary gear PG1, since the rotation of the sun gear s1 is stopped by the first brake B1, the input rotation from the carrier ca1 is accelerated and output as the rotation of the ring gear r1.
In this state, the rotation of the second intermediate shaft M2 is accelerated and transmitted to the carrier ca2 of the second planetary gear PG2, but is the output transmission member of the first clutch C1 and the sun gear s2 of the second planetary gear PG2. The third intermediate shaft M3 (in the velocity diagram, it is the fourth rotation element rm4 and corresponds to the fourth vertical line from the left in the drawing) is rotated at a very high speed. However, in this embodiment, since the third intermediate shaft M3 is provided at the axial center position of the transmission SC, it is affected by centrifugal force and the input / output members of the clutch are in contact with each other for traveling. It can suppress that performance falls.

5 Reverse (REV)
At this gear position, as shown in the operation table, only the third clutch C3 and the second brake B2 are engaged. As shown in the lowermost stage, when the third clutch C3 is engaged, the rotation of the second intermediate shaft M2 is transmitted to the sun gear s1 of the first planetary gear PG1. In the first planetary gear PG1, since the rotation of the carrier ca1 is stopped by the second brake B2, the input rotation from the sun gear s1 is reversed and output as the rotation of the ring gear r1.

FIG. 7 shows the shift state in the hybrid drive device according to the present application, which is capable of four-speed shift as described above.
The horizontal axis represents the vehicle speed, and the vertical axis represents the driving force. And the three-dot chain line extended in the upper right direction from the lower side has shown the speed-change boundary accompanying the increase in vehicle speed. In the figure, the shift speed indicated on the left side of the arrow indicates the shift speed before the shift, and the shift speed indicated on the right side of the arrow indicates the shift speed after the shift. The electronic control unit ECU determines a gear position to be changed in the transmission SC based on a map as shown in this figure, and appropriately controls the friction engagement elements. As a result, the load on the motor side can be reduced and a compact device can be used.

  The above is the description of the drive transmission state associated with the speed change in the present embodiment. Hereinafter, with reference to FIG. 2, the planetary gear set PGS1, the arrangement relationship of the friction engagement elements, and the storage relationship in the mission case MC explain.

As can be seen from the figure, the hybrid drive device HE of the present application employs multi-plate type clutches C1, C2, C3 and brakes B1, B2.
The clutches C1, C2, and C3 press a plurality of friction plates provided on the clutch hub ch side against a plurality of friction counterpart plates provided on the clutch drum side by sliding of a piston that is moved in the axial direction by hydraulic pressure. Thus, the drive transmission between the clutch input member (for example, the clutch drum) and the clutch output member (the clutch hub for the clutch drum) is realized in the engaged state, and the drive transmission does not occur in the non-engaged state.

  The brakes B1 and B2 are formed by pressing a plurality of friction plates provided on the brake hub bh side against a plurality of friction counterpart plates provided on the transmission case MC side by pistons moved in the axial direction by hydraulic pressure. The rotation of the rotating member connected to bh is stopped.

  In this example, the one-way clutch F1 includes an outer race ol fixed to the inner peripheral side surface of the brake cylinder bc in which the piston p of the first brake B1 is slidably housed, and a carrier ca1 of the first planetary gear PG1. It is a sprag type one-way clutch provided with a sprag between the inner race il connected to the integrated brake hub bh.

  The arrangement of the friction engagement elements will be described. At the substantially central portion in the axial direction between the second electric motor MG2 and the planetary gear set PGS1, the inner diameter side (the second clutch C2, the third clutch C3, the first brake B1, and the first brake B1) The first clutch C1 is disposed on the inner diameter side of the two brakes B2.

  A second clutch C2, a third clutch C3, a first brake B1, and a second brake B2 are arranged in this order from the second electric motor MG2 side on the outer diameter side of the first clutch C1. The one-way clutch F1 is disposed on the inner diameter side of the brake cylinder bc of the first brake B1 between the first clutch C1 and the planetary gear set PGS1.

As can be seen from FIG. 2, the mission case MC is configured such that its radial width decreases from the input shaft I side to the output shaft O side.
Accordingly, the relationship between the radially outer position of the second electric motor MG2 and the radially outer positions of the second clutch C2, the third clutch C3, the first brake B1, and the second brake B2 described above is such that the latter side is the former side. By setting the inner side more inside, a hybrid drive device is realized which is not as wide as the output shaft O side.

  Further, the second clutch C2, the third clutch C3, the first brake B1, and the second brake B2 are disposed radially outward with respect to the first clutch C1, and the second clutch C2 and the third clutch. The C3 and the first brake B1 are overlapped in the axial direction with respect to the first clutch C1 (there is a region overlapping in position), thereby realizing a compact configuration in the axial direction.

  Now, in the output side transmission member of the first clutch C1 (in this example, the third intermediate shaft M3 connected to the clutch hub ch (this shaft forms the first transmission member t1 in the present application), the output shaft side A sun gear s2 of the second planetary gear PG2 is provided at the end, and is connected to the output transmission member of the second clutch C2 (in this example, the clutch hub ch), which constitutes the fourth rotating element rm4 of the planetary gear set PGS1. The second transmission member t2 has its output shaft side end connected to the carrier ca1 of the first planetary gear PG1 and constitutes the second rotation element rm2 of the planetary gear set PGS1), and the output transmission of the third clutch C3 The member (in this example, the third transmission member t3 connected to the clutch hub ch, the sun gear s1 of the first planetary gear PG1 is at the output shaft side end thereof. 2 and the first rotational element rm1 of the planetary gear set PGS1), as can be seen from FIG. 2, is a member having the same axis as the third intermediate shaft M3 ( 1st transmission member t1) is arrange | positioned, and the 2nd transmission member t2 and the 3rd transmission member t3 are arrange | positioned in order at the radial direction outer side.

  In the velocity diagram shown on the right side of FIG. 5 described above, the speed relationship of these members at the forward fourth speed is higher at the radially inner member, and by adopting this arrangement configuration, In the hybrid drive device in which such an overdrive state occurs, by arranging the member M3 (t1) that can rotate at the highest speed on the shaft center side, the input / output members of the clutch come into contact with each other and the running performance is improved. It can suppress that it falls.

Second Embodiment FIG. 8 shows a specific structure after the transmission SC of this embodiment, and FIG. 9 shows a skeleton diagram thereof. The difference between this embodiment and the first embodiment will be described. In the first embodiment, the first clutch C1 is disposed between the second electric motor MG2 and the planetary gear set PGS1. In this embodiment, the first clutch C1 is disposed on the output shaft O side of the planetary gear set PGS2.

  In this example, the second intermediate shaft M2 and the third intermediate shaft M3 are connected to rotate integrally by a spline. The output from the carrier ca2 of the second planetary gear PG2 is configured to be transmitted to the output shaft O via the output drum OD and the fourth intermediate shaft M4 provided so as to be integrally rotatable therewith. .

  Accordingly, the skeleton diagram is as shown in FIG. 9, and the operation table of the friction engagement elements is the same as that shown in FIG. The velocity diagram is the same as that shown on the right side of FIG.

  Hereinafter, the arrangement relationship between the planetary gear set PGS2 and the friction engagement elements C1, C2, C3, B1, B2, and F1 and the storage relationship in the mission case MC will be described with reference to FIG.

In this embodiment, a planetary gear set PGS2 which is a planetary gear set for transmission is provided, and the friction engagement elements are arranged so as to sandwich the planetary gear set PGS2.
That is, as shown in FIG. 8, between the second electric motor MG2 and the planetary gear set PGS2, the second clutch C2 is located on the inner diameter side (the inner diameter side from the third clutch C3, the first brake B1, and the second brake B2). The third clutch C3, the first brake B1, and the second brake B2 are disposed on the outer diameter side thereof. On the other hand, with respect to the planetary gear set PGS2, the first clutch C1 is disposed on the opposite side (output shaft O side) from the friction engagement element.

  The one-way clutch F1 is disposed on the inner diameter side of the brake cylinder bc of the first brake B1 between the second clutch C2 and the planetary gear set PGS2.

Even in this example, the mission case MC of the hybrid drive apparatus HE is configured such that its axial radial width decreases from the input shaft I side to the output shaft O side.
In this regard, in the relationship between the radially outer position of the second electric motor MG2 and the radially outer positions of the third clutch C3, the first brake B1, and the second brake B2 described above, the latter side has a larger inner diameter than the former side. Thus, a hybrid drive device that does not have a width in the radial direction is realized.

  In addition, the third clutch C3, the first brake B1, and the second brake B2 are disposed radially outside the second clutch C2, and the third clutch C3 and the first brake B1 The clutch C2 is overlapped in the axial direction (there is a region overlapping in position), thereby realizing a compact configuration in the axial direction.

  Hereinafter, an output side transmission member of the first clutch C1 (in the case of this example, the first transmission member t1 connected to the clutch hub ch, the sun gear s2 of the second planetary gear PG2 is provided at the output shaft side end, 4th rotation element rm4 of the planetary gear set PGS2) and the output side transmission member of the second clutch C2 (in this example, the second transmission member t2 connected to the clutch hub ch, and its output shaft side end is Connected to the carrier ca1 of the first planetary gear PG1, rm2 serving as the second rotation element of the planetary gear set PGS2, and the output transmission member of the third clutch C3 (in this example, the third connected to the clutch hub ch) At the transmission member t3, the output side end is connected to the sun gear s1 of the first planetary gear PG1, and the planetary gear set PGS2 requires the first rotation. The relationship to become rm1) to explain.

  As can be seen from FIG. 8, the third intermediate shaft M3, the first transmission member t1, the second transmission member t2, and the third transmission member t3 are members having the same axis, and the third intermediate shaft M3, The shaft M3 is disposed, and the first transmission member t1 and the second transmission member t2 are disposed on the outer side in the radial direction with different positions in the axial direction. The third transmission member t3 is disposed radially outside the second transmission member t2. The second transmission member t2 and the third transmission member t3 can output driving from the left side in the drawing to the right side, and the first transmission member t1 can output driving from the right side in the drawing to the left side.

  In the velocity diagram shown in FIG. 5 described above, the speed of these members at the fourth forward speed is the highest in the first transmission member t1, but by adopting this arrangement configuration, the overdrive as in the present application is performed. In the hybrid drive system where the situation occurs, by disposing the member that can rotate at the highest speed on the axial center side as much as possible, it is possible to prevent the input / output members of the clutch from coming into contact with each other to reduce the running performance can do.

Third Embodiment Based on FIG. 10 which is a skeleton diagram of the third embodiment, FIG. 4 which is an operation table of friction engagement elements, and FIG. 11 which is a velocity diagram, first, drive transmission of this embodiment will be described.

  Also in this embodiment, a pair of planetary gears PG constituting a planetary gear set PGS3 that is a planetary gear set for shifting is moved from the second intermediate shaft M2 that is an input member of the transmission SC toward the output shaft O, and the first planetary gear PG1, When referred to as second planetary gear PG2, these planetary gears are single planetary gears provided with sun gears s1, s2, carriers ca1, ca2, and ring gears r1, r2, respectively.

  In terms of the relationship with the friction engagement element, the ring gear r1 of the first planetary gear PG1 is selectively fixed to the transmission case MC by the first brake B1, and the carrier ca1 of the first planetary gear PG1 is connected to the ring gear r2 of the second planetary gear PG2. A structure is employed in which the rotation of the second intermediate shaft M2 is selectively transmitted by the second clutch C2 and is selectively fixed to the transmission case MC by the second brake B2. The rotation is prevented from being reversed by the one-way clutch F1 as in the previous example.

  Further, the sun gear s1 of the first planetary gear PG1 is connected to the sun gear s2 of the second planetary gear PG2, and the rotation of the intermediate shaft M2 is selectively transmitted by the first clutch C1, so that the carrier ca2 of the second planetary gear PG2 is an output shaft. A structure that is drivingly connected to O is adopted.

  Further, the ring gear r1 of the first planetary gear PG1 is configured so that the rotation of the second intermediate shaft M2 is selectively transmitted by the third clutch C3.

  The speed diagram of the transmission SC is as shown in FIG. From this speed diagram, there is a rotating element that rotates at a speed of about −1 to 2 times the engine speed that may be input to the transmission SC in an overdrive state with a shift. I understand that.

  The correspondence between the three elements constituting each planetary gear PG and the vertical line is shown on the upper side of the velocity diagram. “R1, Ca1, S1” indicates that they correspond to the ring gear r1, carrier ca1, and sun gear s1 of the first planetary gear PG1, respectively. “R2, Ca2, S2” respectively indicates the ring gear of the second planetary gear PG2. It corresponds to r2, carrier ca2, and sun gear s2.

The vertical line shown in the figure will be described. In order from the right side to the left side, the rotating element provided with the ring gear r1 of the first planetary gear PG1 (which becomes the first rotating element rm1 with respect to the planetary gear set PGS3), the first planetary gear. Rotating element configured to rotate integrally with the carrier ca1 of PG1 and provided with the ring gear r2 of the second planetary gear PG2 (becomes the second rotating element rm2 with respect to the planetary gear set PGS3), integrated with the carrier ca2 of the second planetary gear PG2. A rotating element (fourth with respect to the planetary gear set PGS3) provided with a rotating element (which becomes the third rotating element rm3 with respect to the planetary gear set PGS3), the sun gear s1 of the first planetary gear PG1 and the sun gear s2 of the second planetary gear PG2. Rolling the elements rm4) is configured to be positioned.
Since the third rotation element rm3 is integrated with the output shaft O, the rotation speed after the shift by the transmission SC is marked on the second vertical line from the left attached to the output. “○” indicates the rotation speed in that state.

  With respect to “◯” described in the speed diagram, “1ST”, “2ND”, “3RD”, “4TH”, and “REV” are added to indicate the correspondence with the respective shift stages. Identification of these notations is the same as described in the first embodiment.

  In the planetary gear set PGS3, the rotation elements are numbered according to the direction from the low speed side to the high speed side in the forward direction, and in the order from the high speed side to the low speed side in the reverse direction. In any case, the number order of the rotating elements does not change.

  The relationship between each rotation element of the transmission and each friction engagement element in this embodiment has the following structure as can be seen from FIG. The first rotation element rm1 is selectively fixed to the transmission case MC by the first brake B1, the second rotation element rm2 is selectively transmitted to the rotation of the intermediate shaft M2 by the second clutch C2, and is transmitted by the second brake B2. It is selectively fixed to the case MC. The third rotation element rm3 can output the output rotation to the output shaft O, and the rotation of the intermediate shaft M2 is selectively transmitted to the fourth rotation element rm4 by the first clutch C1. Further, the rotation of the intermediate shaft M2 is selectively transmitted to the first rotation element rm1 by the third clutch C3. This relationship is the same as described in the first embodiment.

  Also in this embodiment, the selection of the engagement / disengagement state of each friction engagement element is the same as the combination shown in the operation table shown in FIG. Therefore, in the transmission SC, any two friction engagement elements are maintained in the engaged state at each shift speed, and the remaining friction engagement elements are disengaged to realize each shift speed.

The speed change state in this embodiment in which the above-described configuration is adopted will be described based on an operation table and a speed diagram.
1 Forward 1st (1ST)
At this shift stage, as shown in the operation table, only the first clutch C1 and the one-way clutch F1 are engaged. The rotation of the second intermediate shaft M2 is transmitted to the third intermediate shaft M3, and this rotation is transmitted to the sun gear s2 of the second planetary gear PG2 by engaging the first clutch C1. In the second planetary gear PG2, when the reverse rotation of the ring gear r2 is stopped by the one-way clutch F1, the input rotation from the sun gear s2 is decelerated and output to the output shaft as the rotation of the carrier ca2.

2 Forward 2nd speed (2ND)
At this gear position, as shown in the operation table, only the first clutch C1 and the first brake B1 are engaged. The rotation of the second intermediate shaft M2 is transmitted to the third intermediate shaft M3, and this rotation is transmitted to the sun gear s2 of the second planetary gear PG2 by engaging the first clutch C1. On the other hand, with respect to the planetary gear set PGS3 including the first planetary gear PG1 and the second planetary gear PG2, the rotation of the ring gear r1 of the first planetary gear PG1 is stopped, whereby the input rotation from the sun gears s1 and s2 is decelerated, and the second planetary gear set PGS3. The rotation is output to the output shaft O as the rotation of the carrier ca2 of the planetary gear PG2. The reduction ratio in this state is smaller than the first forward speed.

3 Forward third speed (3RD)
At this shift stage, as shown in the operation table, only the first clutch C1 and the second clutch C2 are engaged. In this state, the rotation of the second intermediate shaft M2 is transmitted to the third intermediate shaft M3, and this rotation is transmitted to the sun gear s2 of the second planetary gear PG2 by engaging the first clutch C1. As for the first planetary gear PG1, the rotational speeds of the sun gear s1 and the carrier ca1 are the same, and for the second planetary gear PG2, the rotational speeds of the sun gear s2 and the ring gear r2 are the same, so that the planetary gear PG is fixed. It becomes. As a result, the rotation of the second intermediate shaft M2 is transmitted to the output shaft O as it is, so-called a directly connected state.

4 forward 4 speed (4TH)
At this shift stage, as shown in the operation table, only the second clutch C2 and the first brake B1 are engaged. By engaging the second clutch C2, the rotation of the second intermediate shaft M2 is transmitted to the carrier ca1 of the first planetary gear PG1 and the ring gear r2 of the second planetary gear PG2. In the first planetary gear PG1, since the rotation of the ring gear r1 is stopped by the first brake B1, the input rotation from the carrier ca1 is accelerated and output as the rotation of the carrier ca2 of the second planetary gear PG2. The

  In this state, the rotation of the second intermediate shaft M2 is accelerated and transmitted to the sun gear s1 of the first planetary gear PG1 and the sun gear s2 of the second planetary gear PG2, but this is the output transmission member of the first clutch C1. One transmission member t1 (in the velocity diagram, which is the fourth rotation element rm4, corresponding to the vertical line at the left end in the drawing) rotates at a very high speed. However, in this embodiment, since the first transmission member t1 is provided on the shaft center side (inner diameter side) of the transmission just outside the third intermediate shaft M3, the centrifugal force It is possible to suppress a decrease in traveling performance due to contact between input / output members of the clutch.

  In this regard, in the third embodiment, as in the second embodiment, as shown in FIG. 10, the second transmission member t2 that is the output side transmission member of the second clutch C2, and the output side of the first clutch C1. Both the first transmission member t1, which is a transmission member, is positioned on the outer diameter side of the third intermediate shaft M3 in the axial radial direction with the axial position between the two different, thereby preventing problems. is doing.

5 Reverse (REV)
At this gear position, as shown in the operation table, only the third clutch C3 and the second brake B2 are engaged. As the third clutch C3 is engaged, the rotation of the second intermediate shaft M2 is transmitted to the ring gear r1 of the first planetary gear PG1. In the first planetary gear PG1, since the rotation of the carrier ca1 is stopped by the second brake B2, the input rotation from the ring gear r1 is reversed, and the rotation of the carrier ca2 of the second planetary gear PG2 is applied to the output shaft. Is output.

3 Fourth Embodiment In the embodiments described so far, a single planetary gear is provided and a planetary gear set for shifting PGS1, PGS2, and PGS3 is configured by appropriately connecting the elements constituting the planetary gear. However, the fourth embodiment is an example in which a so-called Ravigneaux type planetary gear set PGR (PGS4) is used as a transmission planetary gear set.

As shown in FIG. 12, the Ravigneaux type planetary gear set PGR includes a common carrier cc and a ring gear rc for the pair of sun gears s1 and s2.
More specifically, it has a common carrier cc that supports the long pinion lp and the short pinion sp, and this long pinion lp is meshed with the short pinion sp, one sun gear s2 and the common ring gear rc, and the short pinion sp is meshed with the other sun gear s1.
The first planetary gear PG1 configured by the other sun gear s1, the common carrier cc, and the ring gear rc is configured so that the common carrier cc rotatably supports a pair of mutually engaged pinions, and is a double pinion type It is structured to be a planetary gear.
The second planetary gear PG2 configured by one sun gear s2, the common carrier cc, and the ring gear rc is configured such that the common carrier cc rotatably supports a single pinion, and is a single pinion type planetary gear. It is said that.

Now, the connection relationship between each element constituting the Ravigneaux type planetary gear set PGR and the friction engagement element will be described below.
As shown in FIG. 12, which is a skeleton of this embodiment, the rotation of the second intermediate shaft M2 is selectively transmitted to the sun gear s2 of the second planetary gear PG2 by the first clutch C1, and the sun gear s1 of the first planetary gear PG1 is the first gear PG1. One brake B1 is selectively fixed to the mission case MC. Further, the rotation of the second intermediate shaft M2 is selectively transmitted by the third clutch C3. Further, the common carrier cc adopts a structure in which the rotation of the second intermediate shaft M2 is selectively transmitted by the second clutch C2 and is selectively fixed to the transmission case MC by the second brake B2. Also in this example, the reverse rotation is stopped by the one-way clutch F1. Further, the ring gear rc is driven and connected to the output shaft O.

  By adopting this structure, according to the operation table previously shown in FIG. 4, by controlling the state of the first clutch C1, the second clutch C2, the first brake B1, the second brake B2, the first forward speed, 2nd forward speed, 3rd forward speed and 4th forward speed can be realized.

  Further, the transmission SC is also provided with a third clutch C3. The third clutch C3 is intermediately connected to the other sun gear s1, which may be selectively fixed to the transmission case MC by the first brake B1. By making it possible to selectively transmit the rotation of the axis M2, a reverse gear can be realized as in the previous example.

In this example, although a speed diagram is omitted, regarding the planetary gear set PGS4 that is a planetary gear set for speed change, the first rotation element provided with the sun gear s1 of the first planetary gear PG1 is selectively set to the transmission case MC by the first brake B1. The rotation of the intermediate shaft M2 is selectively transmitted to the second rotation element fixed and connected to the common carrier cc by the second clutch C2. The third rotation element provided with the common ring gear rc can output the output rotation to the output shaft O, and the fourth rotation element provided with the sun gear s2 of the second planetary gear PG2 rotates the intermediate shaft M2 by the first clutch C1. Is selectively transmitted.
Further, the rotation of the second intermediate shaft M2 is selectively transmitted to the second rotation element via the third clutch C3 for reverse rotation.

The above is the description of the drive transmission state associated with the speed change in the embodiment. Hereinafter, the arrangement relationship between the planetary gear set and the friction engagement element in this embodiment will be described.
In this example, the friction engagement elements are arranged so as to sandwich the planetary gear set PGS4.
That is, the first clutch C1 is disposed on the inner diameter side between the second electric motor MG2 and the speed change planetary gear set PGS4, and the third clutch C3 and the first brake B1 are disposed on the outer diameter side thereof. ing. On the other hand, the second clutch C2 is disposed on the opposite side (output shaft O side) to the planetary gear set PGS4. The second brake B2 is disposed on the outer diameter side of the second clutch C2, and the one-way clutch F1 is disposed between the first brake B1 and the second brake B2.

  Further, the third clutch C3, the first brake B1, and the second brake B2 are disposed radially outward with respect to the first clutch C1, and the third clutch C3 is disposed with respect to the first clutch C1. By overlapping in the axial direction (there is a region that overlaps in position), a compact configuration is realized also in the axial direction.

In this embodiment, the lower side of the second intermediate shaft M2 extends beyond the position of the first clutch C1 to the input position of the second clutch C2, which is the output shaft O side of the planetary gear set PGS4.
Then, the output side transmission member of the first clutch C1 (the first transmission member t1 is provided with the sun gear s2 at the output shaft side end thereof, which becomes the fourth rotating element of the planetary gear set PGS4), and the second clutch Regarding the relationship with the output transmission member of C2 (the second transmission member t2, the output shaft side end of which is connected to the carrier cc of the Ravigneaux type planetary gear set and becomes the second rotation element of the planetary gear set PGS4) 12, the first transmission member t1 and the second transmission member t2 are members having the same rotational axis, and the third intermediate shaft M3 (which is the second intermediate shaft M2) Rotation is transmitted to the input side of the second clutch C2), a first transmission member t1 is disposed on the radially outer side, and a second transmission member t2 is disposed on the outer diameter side. Yes.

  As described above, the speed of these members at the fourth forward speed is the highest in the first transmission member t1, but by adopting this arrangement configuration, a hybrid in which an overdrive state as in the present application occurs. In the driving device HE, by disposing the member that can rotate at the highest speed on the axial center side as much as possible, for example, it is possible to prevent the input / output members of the clutch from contacting each other to reduce the running performance. Can do.

In the first to fourth embodiments described above, the hybrid drive device has the following functions and effects.
While adopting the 1.2 motor split system, it is possible to realize a speed-up shift stage and a reverse stage that increase the input rotational speed in the transmission by adopting a transmission SC having a relatively simple structure.
2. Although it is possible to realize a state in which the engine rotational speed is increased and input to the transmission, it is possible to prevent the output transmission member of the friction engagement element from over-rotating at all speed stages. In particular, a hybrid drive device that increases the gear ratio is advantageous because the rotational speed of the output transmission member tends to increase.
3. Since it is not necessary to simultaneously control the engagement / disengagement of the clutch and the disengagement / engagement control of other clutches during the shift control, it is difficult to generate a shift shock.

Further multi-stage embodiments The embodiments described so far are examples of a hybrid drive device capable of realizing four forward speeds and one reverse speed with a transmission. The fifth and sixth embodiments described below are the same. The embodiment is an example in which the number of stages is further increased.
In the fifth embodiment, 6 forward speeds and 1 reverse speed are realized, and in the sixth embodiment, 8 forward speeds and 2 reverse speeds are realized.
Hereinafter, these embodiments will be described based on a skeleton diagram, an operation table, and a velocity diagram.
Even in these embodiments, the structure up to the input to the transmission is the same as the structure described above. From the engine E side to the output shaft O side, the first electric motor MG1, the power distribution planetary gear PG0. The second electric motor MG2 is provided, and the rotation of the ring gear r0 of the power distribution planetary gear PG0 is input to the transmission SC. Therefore, a state in which high-speed rotation in an overdrive state is input also occurs.

Fifth Embodiment FIG. 13 shows a skeleton diagram of this embodiment, FIG. 14 shows an operation table, and FIG. 15 shows a velocity diagram.
As shown in the skeleton diagram, planetary gear set PGS5 provided in transmission SC in this embodiment is configured as a combination of first planetary gear PG1 provided on the input side and Ravigneaux type planetary gear set PGR provided on the output side. . In this example, the Ravigneaux type planetary gear set PGR corresponds to a speed change planetary gear set. This is because the first planetary gear PG1 shifts the input rotation to the transmission SC and transmits it to the shifting planetary gear set.

  The first planetary gear PG1 is a single planetary gear provided with a sun gear s1, a carrier ca1, and a ring gear r1. The Ravigneaux type planetary gear set PGR includes a pair of sun gears s2 and s3, a common carrier cc, and a common ring gear rc. The sun gear s1 of the first planetary gear PG1 is always fixed to the mission case MC.

  The Ravigneaux type planetary gear set PGR includes a common carrier cc and ring gear rc for the pair of sun gears s2 and s3. More specifically, it has a common carrier cc that supports the long pinion lp and the short pinion sp, and the long pinion lp meshes with the short pinion sp, one sun gear s2 and the common ring gear rc, and the short pinion sp The second planetary gear PG2 and the sun gear s3 disposed on the output side are common to the planetary gear composed of the sun gear s2 disposed on the input side, the common carrier cc and the ring gear rc. When the planetary gear composed of the carrier cc and the ring gear rc is referred to as a third planetary gear PG3, the second planetary gear PG2 is a single pinion planetary gear, and the third planetary gear PG3 is a double-pinion planetary gear.

  In the speed diagram shown in FIG. 15, the speed diagram relating to the operating state of the first planetary gear PG1 is shown on the left side of the figure, and the speed diagram relating to the operating state of the Ravigneaux type planetary gear set PGR is shown on the right side of the figure.

  In the velocity diagram on the left side of the figure, the vertical line and the correspondence between the elements constituting the first planetary gear PG1 are shown on the upper side of the figure. Here, “S1, Ca1, R1” indicate that the rotating element is provided with the sun gear s1 of the first planetary gear PG1, the rotating element connected to the carrier ca1, and the rotating element provided with the ring gear r1, respectively. Yes.

  In the velocity diagram on the right side of the drawing, the correspondence between the vertical lines and the elements constituting the second and third planetary gears PG2 and PG3 is shown on the upper side of the drawing. Here, “S2, Ca2, R2” indicate that the rotating element is provided with the sun gear s2 of the second planetary gear PG2, the rotating element connected to the carrier cc, and the rotating element provided with the ring gear rc. Yes. “S3, Ca3, R3” indicates that the rotating element is provided with the sun gear s3 of the third planetary gear PG3, the rotating element connected to the carrier cc, and the rotating element provided with the ring gear rc. In the case of the Ravigneaux type planetary gear PGR, the carrier cc and the ring gear rc are common between the second planetary gear PG2 and the third planetary gear PG3.

  Returning to FIG. 13, in this example, one sun gear (the sun gear s3 of the third planetary gear PG3) is selectively transmitted with the reduced rotation reduced by the first planetary gear PG1 by the first clutch C1. The sun gear (the sun gear s2 of the second planetary gear PG2) is selectively fixed to the transmission case MC by the first brake B1. In addition, the carrier cc of the Ravigneaux type planetary gear set PGR is selectively transmitted to the transmission case MC by the second brake B2 while the input rotation of the transmission SC is selectively transmitted by the second clutch C2. Has been. Further, the ring gear rc of the Ravigneaux type planetary gear set PGS is drivingly connected to the output shaft O.

  Further, the carrier cc of the Ravigneaux type planetary gear set PGR is configured such that its reverse rotation is stopped by the one-way clutch F1, and the first planetary gear PG1 is connected to the sun gear s2 of the second planetary gear PG2 in order to realize the reverse gear. The rotation of the carrier cc is selectively transmitted by the third clutch C3.

Shift 1 First forward speed (1ST)
At this shift stage, as shown in the operation table, only the first clutch C1 and the one-way clutch F1 are engaged. By engaging the first clutch C1, the rotation of the carrier ca1 of the first planetary gear PG1 decelerated with respect to the input rotation is transmitted to the first transmission member t1 provided with the sun gear s3 of the third planetary gear PG3. The In the Ravigneaux type planetary gear set PGR, when the reverse rotation of the common carrier cc is stopped by the one-way clutch F1, the input rotation from the sun gear s3 of the third planetary gear PG3 is decelerated and output as the rotation of the common ring gear rc. Output to axis O.

2 Forward 2nd speed (2ND)
At this gear position, as shown in the operation table, only the first clutch C1 and the first brake B1 are engaged. By engaging the first clutch C1, the rotation of the carrier ca1 of the first planetary gear PG1 decelerated with respect to the input rotation is transmitted to the first transmission member t1 provided with the sun gear s3 of the third planetary gear PG3. The In the Ravigneaux type planetary gear set PGR, the rotation of the sun gear s2 of the second planetary gear PG2 is stopped by the first brake B1, whereby the input rotation of the third planetary gear PG3 from the sun gear s3 is decelerated, and the common ring gear rc Output to the output shaft as rotation. The reduction ratio in this state is smaller than the first forward speed.

3 Forward third speed (3RD)
At this gear position, as shown in the operation table, only the first clutch C1 and the third clutch C3 are engaged. The input rotation is transmitted to the common carrier cc while the rotation of the carrier ca1 of the first planetary gear PG1 is transmitted to the sun gear s3 of the third planetary gear PG3. As a result, the rotation of the carrier ca1 of the first planetary gear PG1 is transmitted to the output shaft O as it is.

4 forward 4 speed (4TH)
At this gear stage, as shown in the operation table, only the first clutch C1 and the second clutch C2 are engaged. When the first clutch C1 is engaged, the rotation of the carrier ca1 of the first planetary gear PG1 whose input rotation is decelerated is transmitted to the sun gear s3 of the third planetary gear PG3, and the input rotation of the transmission SC is transmitted to the common carrier cc. . As a result, the rotation in the acceleration state with respect to the rotation of the sun gear s3 and the rotation in the deceleration state with respect to the common carrier cc is output as the rotation of the ring gear r3 of the third planetary gear PG3.

5 forward 5 speed (5TH)
At this shift stage, as shown in the operation table, only the second clutch C2 and the third clutch C3 are engaged. As the second clutch C2 is engaged, the input rotation of the transmission SC is transmitted to the common carrier cc, and the rotation of the carrier ca1 of the first planetary gear PG1 is transmitted as the rotation of the sun gear s2 of the second planetary gear PG2. Is done.

  In this state, with respect to the third planetary gear PG3, the first transmission member t1 that is the output side transmission member of the first clutch C1 rotates at a very high speed. However, in this embodiment, the first transmission member t1 is provided on the shaft center side (inner diameter side) of the transmission SC just outside the intermediate shaft M3. It can suppress that the input / output member of a clutch contacts and the driving | running | working performance falls.

6 Forward 6th (6TH)
At this shift stage, as shown in the operation table, only the second clutch C2 and the first brake B1 are engaged. As the second clutch C2 is engaged, the input rotation of the transmission SC is transmitted to the common carrier cc. Then, since the rotation of the sun gear s2 of the second planetary gear PG2 is stopped by the first brake B1, the rotation in the accelerated state with respect to the rotation of the common carrier cc is output as the ring gear rc of the third planetary gear PG3. The
This speed increase is larger than the fifth forward speed.

  In this state, with respect to the third planetary gear PG3, the first transmission member t1 that is the output side transmission member of the first clutch C1 rotates at a higher speed than the state in the fifth forward speed. However, since the first transmission member t1 is provided on the shaft center side (inner diameter side) of the transmission SV immediately outside the intermediate shaft M3, the first transmission member t1 is affected by centrifugal force and the like, and the clutch is engaged. It can suppress that output members contact and driving performance falls.

  In this embodiment, the relationship between the input rotation and the output rotation of the transmission SC is a rotation in which the input rotation is transmitted from the second intermediate shaft M2 to the ring gear r1 of the first planetary gear PG1, and the output rotation is a Ravigneaux type planetary gear. The common ring gear rc of the set PGR is rotated. Therefore, in the transmission SC, an overdrive state (becomes a speed increasing gear) is realized with five forward speeds and six forward speeds.

7 Reverse (REV)
At this gear position, as shown in the operation table, only the third clutch C3 and the second brake B2 are engaged. By engaging the third clutch C3, the rotation of the carrier ca1 of the first planetary gear PG1, which is a reduced input rotation, is transmitted to the sun gear s2 of the second planetary gear PG2. In the second planetary gear PG2, since the rotation of the carrier cc is stopped by the second brake B2, the input rotation from the sun gear s2 is reversed and output to the output shaft as the rotation of the common ring gear rc.

  As previously indicated, in this embodiment, the Ravigneaux type planetary gear set PGR corresponds to the speed change planetary gear set referred to in the present application, but below the vertical line described in the speed diagram shown in FIG. The first to fourth rotating elements rm1, rm2, rm3, and rm4 are shown in FIG.

Sixth Embodiment FIG. 16 shows a skeleton diagram of this embodiment, FIG. 17 shows an operation table, and FIG. 18 shows a velocity diagram.
As shown in the skeleton diagram, in this embodiment as well, as in the fifth embodiment, the first planetary gear PG1 provided on the input side and the Ravigneaux type planetary gear set PGR provided on the output side are combined.
In this example, the first planetary gear PG1 is a double-pinion planetary gear that includes a sun gear s1, a carrier ca1, and a ring gear r1, and rotatably supports a pair of pinions that the carrier ca1 meshes with each other.

This Ravigneaux type planetary gear set PGR is also provided with a common carrier cc and ring gear rc for the pair of sun gears s2, s3. This example also has a common carrier cc that supports the long pinion lp and the short pinion sp, and the long pinion lp meshes with the short pinion sp, one sun gear s2 and the common ring gear rc, and the short pinion sp Although the structure which meshes | engages with the sun gear s3 is adopted, the planetary which consists of the sun gear s2 arrange | positioned at the input side, the common carrier cc, and the ring gear rc is set to the 2nd planetary gear PG2, and the sun gear s3 arranged at the output side A planetary composed of the common carrier cc and the ring gear rc is referred to as a third planetary gear PG3.
Further, a fourth clutch C4 is newly added to the fifth embodiment, and this fourth clutch C4 is used to input rotation of the transmission SC and to rotate the carrier ca1 of the first planetary gear PG1. A configuration is employed in which transmission is selectively transmitted to the sun gear s2 of the second planetary gear PG2.

  In the speed diagram shown in FIG. 18, a speed diagram relating to the operating state of the first planetary gear PG1 is shown on the left side of the figure, and a speed diagram relating to the operating state of the Ravigneaux type planetary gear set PGR is shown on the right side of the figure.

  The description of the velocity diagram shown in FIG. 18 in the present embodiment conforms to the description of FIG. 15 in the fifth embodiment.

  Even in this example, one sun gear (the sun gear s3 of the third planetary gear PG3) is selectively transmitted by the first clutch C1 to the reduced rotation reduced by the first planetary gear PG1, and the other sun gear (the second planetary gear). The sun gear s2) of PG2 is selectively fixed to the transmission case MC by the first brake B1. Further, the carrier cc of the Ravigneaux type planetary gear set PGR is selectively transmitted to the transmission case MC by the second brake B2 while the input rotation of the transmission SC is selectively transmitted by the second clutch C2. ing. Further, the ring gear rc of the Ravigneaux type planetary gear set PGR is driven and connected to the output shaft O.

  Further, the carrier cc of the Ravigneaux type planetary gear set PGR is configured such that its reverse rotation is stopped by the one-way clutch F1, and the first planetary gear PG1 is connected to the sun gear s2 of the second planetary gear PG2 in order to realize the reverse gear. The configuration in which the rotation of the ring gear r1 is selectively transmitted by the third clutch C3 is employed.

  Further, as described above, the fourth clutch C4 is used to selectively rotate the carrier ca1 of the first planetary gear PG1 to the sun gear s2 of the second planetary gear PG2 as well as the input rotation of the transmission SC. The structure which transmits to is adopted.

Shifting The shifting operation from the first forward speed (1ST) to the third forward speed (3RD) is substantially the same as that of the corresponding shift stage of the fifth embodiment, and thus description thereof is omitted.
4 forward 4th speed At this speed, only the first clutch C1 and the fourth clutch C4 are engaged, as shown in the operation table. By engaging the first clutch C1, the rotation of the ring gear r1 of the first planetary gear PG1 is transmitted to the first transmission member t1. By engaging the fourth clutch C4, in the Ravigneaux type planetary gear set PGR, the rotation of the carrier ca1 of the first planetary gear PG1 is output to the output shaft O in a decelerated state.

5 Forward 5th Speed At this speed, only the first clutch C1 and the second clutch C2 are engaged as shown in the operation table. This state is substantially the same as the forward four-stage described in the fifth embodiment, and the rotation of the common carrier cc is decelerated and output to the output shaft as the rotation of the common ring gear rc.

6 Forward 6th (6TH)
At this gear stage, as shown in the operation table, only the second clutch C2 and the fourth clutch C4 are engaged. In this operating state, the input rotation of the transmission SC is directly output to the output shaft O as the rotation of the common ring gear rc.

7 Forward 7th (7TH)
At this shift stage, as shown in the operation table, only the second clutch C2 and the third clutch C3 are engaged. When the second clutch C2 is engaged, the rotation of the second intermediate shaft M2 is transmitted to the common carrier cc, and when the third clutch C3 is engaged, the rotation of the ring gear r1 of the first planetary gear PG1. Is transmitted as the rotation of the sun gear s2 of the second planetary gear PG2. As a result, the rotation in the accelerated state with respect to the rotation of the common carrier cc is output as the ring gear rc of the third planetary gear PG3.

  In this state, with respect to the third planetary gear PG3, the first transmission member t1 that is the output side transmission member of the first clutch C1 rotates at a very high speed. However, in this embodiment, the first transmission member t1 is provided on the shaft center side (inner diameter side) of the transmission SC just outside the third intermediate shaft M3. It can be suppressed that the input / output members of the clutch are brought into contact with each other under the influence of force or the like and the running performance is deteriorated.

8 Forward 8th (8TH)
At this shift stage, as shown in the operation table, only the second clutch C2 and the first brake B1 are engaged. As the second clutch C2 is engaged, the rotation of the second intermediate shaft M2 is transmitted to the common carrier cc. Since the rotation of the sun gear s2 of the second planetary gear PG2 is stopped by the first brake B1, the rotation in the accelerated state is output as the ring gear rc with respect to the rotation of the common carrier cc.
In this state, with respect to the third planetary gear PG3, the first transmission member t1 that is the output side transmission member of the first clutch C1 rotates at a higher speed than the state in the seventh forward speed. However, since the first transmission member t1 is provided on the shaft center side (inner diameter side) of the transmission just outside the second intermediate shaft M2, the first transmission member t1 is affected by centrifugal force and the like, It can suppress that input-output members contact and driving performance falls.

  In this embodiment, the relationship between the input rotation and the output rotation of the transmission SC is a rotation in which the input rotation is transmitted from the second intermediate shaft M2 to the carrier ca1 of the first planetary gear PG1, and the output rotation is the Ravigneaux type planetary gear. The common ring gear rc of the set PGR is rotated. Therefore, in the transmission SC, the seventh forward speed and the eighth forward speed are in the overdrive state (the speed-up speed stage).

9 Reverse 1 (REV1)
At this gear position, as shown in the operation table, only the third clutch C3 and the second brake B2 are engaged. By engaging the third clutch C3, the rotation of the ring gear r1 of the first planetary gear PG1 is transmitted to the sun gear s2 of the second planetary gear PG2. In the second planetary gear PG2, since the rotation of the common carrier cc is stopped by the second brake B2, the input rotation from the sun gear s2 is reversed and output to the output shaft O as the rotation of the common ring gear rc. The

10 Reverse 2 (REV2)
At this gear position, as shown in the operation table, only the fourth clutch C4 and the second brake B2 are engaged. As the fourth clutch C4 is engaged, the rotation of the carrier ca1 of the first planetary gear PG1 is transmitted to the sun gear s2 of the second planetary gear PG2. In the second planetary gear PG2, since the rotation of the common carrier cc is stopped by the second brake B2, the input rotation of the sun gear s2 is reversed and output to the output shaft as the rotation of the common ring gear rc. This rotation is faster than the first reverse speed.

Also in this embodiment, this Ravigneaux type planetary gear set PGR corresponds to a planetary gear set for shifting. The correspondence of the first to fourth rotating elements rm1, rm2, rm3, and rm4 is shown below the vertical line shown in the velocity diagram shown in FIG.

Another Embodiment of First Embodiment Seventh Embodiment FIG. 19 shows a skeleton diagram of this embodiment. Even in this example, the first clutch C1, the second clutch C2, and the third clutch C3 are provided between the second electric motor MG2 and the planetary gear set for shifting. Further, the planetary gear set PGS7 is configured as a combination of a pair of single planetary gears PG1 and PG2.
When described in comparison with the first embodiment, the relationship among the rotating elements in the second planetary gear PG2 and the first clutch C1, the output shaft O, and the second brake B2 is the same, but in this example, the first planetary gear PG1. The sun gear s1 is connected to the output side transmission member t1 of the first clutch C1, and the rotation of the second intermediate shaft M2 is selectively transmitted. The carrier ca1 employs a structure in which the rotation of the second intermediate shaft M2 is selectively transmitted from the second clutch C2, and is configured to rotate integrally with the ring gear r2 of the second planetary gear PG2. As a result, the operation control by the second brake B2 and the one-way clutch F1 is also received.
Further, the rotation of the ring gear r1 can be stopped by the first brake B1, and the rotation of the second intermediate shaft M2 is selectively transmitted by the third clutch C3.

  Even in this embodiment, four forward stages and one reverse stage can be realized by using the friction engagement elements C1, C2, C3, B1, B2, and F1 according to the operation table shown in FIG.

  In the hybrid drive device that employs the two-motor split method, it has been possible to provide a hybrid drive device that can ensure acceleration performance in the low speed region and also have high fuel efficiency in the high speed region.

The figure which shows the whole schematic structure of a hybrid drive unit The figure which shows the detail of the structure of the transmission vicinity of 1st embodiment. Skeleton diagram of the first embodiment Operation table of the first embodiment Velocity diagram of the first embodiment Explanatory drawing of the drive state of the first embodiment Driving force diagram of the first embodiment The figure which shows the detail of the structure of the transmission vicinity of 2nd embodiment. Skeleton diagram of the second embodiment Skeleton diagram of the third embodiment Velocity diagram of the third embodiment Skeleton diagram of the fourth embodiment Skeleton diagram of the fifth embodiment Operation table of the fifth embodiment Velocity diagram of the fifth embodiment Skeleton diagram of the sixth embodiment Operation table of the sixth embodiment Velocity diagram of the sixth embodiment Skeleton diagram of another embodiment of the first embodiment

Explanation of symbols

HE Hybrid drive device MC Mission case E Engine I Input shaft D Damper MG1 First motor PG0 Power distribution planetary gear MG2 Second motor In Inverter B Battery MP Motor part ECU Electronic control device M1 First intermediate shaft M2 Second intermediate shaft (intermediate) Transmission shaft)
SC transmission O output shaft C1 first clutch (friction engagement element)
C2 Second clutch (friction engagement element)
C3 3rd clutch (friction engagement element)
B1 First brake (friction engagement element)
B2 Second brake (friction engagement element)
F1 one-way clutch (friction engagement element)
PG1 First planetary gear PG2 Second planetary gear PGS Planetary gear set PGR Ravigneaux type planetary gear set s Sun gear ca Carrier r Ring gear t Transmission member rm Rotating element

Claims (15)

An input shaft that is drivingly connected to the engine and an output shaft that is drivingly connected to the wheels.
A first electric motor and a second electric motor,
Planetary gears constituting first, second, and third rotating elements in order of rotational speed, wherein the first motor is connected to the first rotating element, the input shaft is connected to the second rotating element, An intermediate transmission shaft and the second electric motor are connected to the third rotating element, and a power distribution planetary gear that can transmit the rotation of the input shaft to the intermediate transmission shaft is accelerated.
A transmission having a plurality of shift stages and outputting power transmitted from the intermediate transmission shaft to the output shaft;
The plurality of shift stages of the transmission at least increase the rotational speed of the intermediate transmission shaft and output it to the output shaft; reduce the rotational speed of the intermediate transmission shaft and output the output It has a reduction gear stage that outputs to the shaft,
The transmission includes a first brake, a first clutch and a second clutch that transmit rotation of the intermediate transmission shaft, and first, second, third, and fourth rotating elements in order of rotational speed. Has a planetary gear set,
The first rotating element is selectively fixed to the case by the first brake;
The second rotating element can selectively transmit rotation of the intermediate transmission shaft by the second clutch, and the third rotating element can output output rotation to the output shaft,
The rotation of the intermediate transmission shaft is selectively transmitted by the first clutch, and the first clutch and the second clutch are arranged between the planetary gear set for speed change and the second electric motor;
The output-side transmission member of the first clutch passes through the inner diameter side of the output-side transmission member of the second clutch, and is connected to the fourth rotation element constituting the transmission planetary gear set ,
The hybrid drive system first clutch hub as an output member of the first clutch that is disposed on the inner diameter side of the second clutch hub as an output member of the second clutch.   2. The hybrid drive device according to claim 1, wherein the first electric motor, the power distribution planetary gear, the second electric motor, and the transmission planetary gear set are arranged coaxially in order from the input shaft side. The transmission further includes a second brake,
3. The hybrid drive device according to claim 1, wherein the second rotating element constituting the speed-changing planetary gear set is selectively fixed to the case by the second brake. The first clutch and the second clutch are arranged on the inner side of the outer surface in the radial direction of the electric motor,
4. The hybrid drive device according to claim 1, wherein at least a part of the first clutch and the second clutch are arranged to overlap in the axial direction. 5.   The hybrid drive apparatus according to claim 4, wherein the first clutch is disposed on an inner diameter side with respect to the second clutch. The transmission further includes a third clutch,
The hybrid drive device according to claim 3, wherein the rotation of the intermediate transmission shaft is selectively transmitted to the first rotation element constituting the transmission planetary gear set by the third clutch. The third clutch is disposed closer to the electric motor than the planetary gear set for shifting, and is disposed on the inner side of the outer radial surface of the electric motor.
The hybrid drive apparatus according to claim 6, wherein the third clutch is disposed so as to overlap in the axial direction with respect to one or both of the first clutch and the second clutch. The first clutch is disposed on the inner diameter side with respect to both the second clutch and the third clutch, and at least a part of the first clutch exceeds the one or both of the second clutch and the third clutch in the axial direction. Wrapped and placed
The hybrid drive device according to claim 7, wherein at least a part of the third clutch is disposed so as to overlap in a radial direction with respect to the second clutch.   The hybrid drive apparatus according to claim 8, wherein the first brake is disposed so as to be at least partially overlapped in a radial direction with respect to one or both of the second clutch and the third clutch.   The output side transmission member of the first clutch passes through the inner diameter side of the output side transmission member of the third clutch and is coupled to the fourth rotation element constituting the transmission planetary gear set. The hybrid drive device as described in any one of Claims. The planetary gear set for speed change is
A three-element first planetary gear and a second planetary gear comprising a sun gear, a carrier and a ring gear;
A sun gear of the first planetary gear is selectively fixed to the case by a first brake;
The carrier of the first planetary gear is connected to the ring gear of the second planetary gear, and the rotation of the intermediate transmission shaft is selectively transmitted by the second clutch,
A ring gear of the first planetary gear is coupled to the carrier of the second planetary gear and is drivingly coupled to the output shaft;
The hybrid drive apparatus according to claim 1 or 2, wherein the rotation of the intermediate transmission shaft is selectively transmitted to the sun gear of the second planetary gear by the first clutch.   4. The rotational speed of the intermediate transmission shaft is output to the output shaft at four different gear ratios according to the engagement / non-engagement state of the first brake, the second brake, the first clutch, and the second clutch. The hybrid drive device described. The planetary gear set for speed change is
A three-element first planetary gear and a second planetary gear comprising a sun gear, a carrier and a ring gear;
The sun gear of the first planetary gear is selectively fixed to the case by the first brake,
The carrier of the first planetary gear is connected to the ring gear of the second planetary gear, the rotation of the intermediate transmission shaft is selectively transmitted by the second clutch, and is selectively fixed to the case by the second brake,
A ring gear of the first planetary gear is coupled to the carrier of the second planetary gear and is drivingly coupled to the output shaft;
The hybrid drive device according to claim 12, wherein the rotation of the intermediate transmission shaft is selectively transmitted to the sun gear of the second planetary gear by the first clutch. The transmission further includes a third clutch;
The hybrid drive device according to claim 12 or 13, wherein the rotation of the intermediate transmission shaft is selectively transmitted to the first rotation element constituting the transmission planetary gear set by the third clutch.   A gear stage having the largest gear ratio in the forward speed is formed in the transmission state in which the rotation of the intermediate transmission shaft is transmitted to the fourth rotation element constituting the speed change planetary gear set by engaging the first clutch. The hybrid drive apparatus as described in any one of Claims 1-14.

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