Vehicular drive system

A vehicular drive system including a first electric motor, a differential mechanism operable to distribute an output of a drive power source to the first electric motor and a power transmitting member, a second electric motor disposed in a power transmitting path between the power transmitting member and a drive wheel, and a power transmitting device disposed between the second electric motor and the drive wheel, wherein the first electric motor, the differential mechanism, the second electric motor and the power transmitting device are arranged in an axial direction of the vehicular drive system, the vehicular drive system being characterized by a support wall disposed between the second electric motor and the power transmitting device and arranged to support the second electric motor, and a first group of oil passages which are formed through the support wall and through which a lubricating oil is supplied to at least one of the first electric motor, the differential mechanism and the second electric motor, and to the power transmitting device.

The present application is based on Japanese Patent Application No. 2004-370038 filed on December 21, the content of which is incorporated herein by reference.

TECHNICAL FIELD

1. Field of the Invention

The present invention relates to a vehicular drive system, and more particularly to techniques for simplifying an arrangement of hydraulic passages for the vehicular drive system.

2. Description of Prior Art

There is known a vehicular drive system including a first electric motor, a second electric motor, and a torque synthesizing and distributing mechanism which is disposed between those two electric motors and which utilizes a differential function of a planetary gear set. Examples of this type of vehicular drive system include a drive system disclosed in JP-2004-116735A. Generally, a planetary gear set functioning as the above-indicated torque synthesizing and distributing mechanism, namely, as a differential mechanism, includes three rotary elements one of which is connected to the first electric motor. One of the other two rotary elements is connected to a drive power source such as an engine, while the other of those two rotary elements is connected to a power transmitting member for mechanically transmitting an output of the planetary gear set to a drive wheel. The second electric motor is disposed on the power transmitting member, or a power transmitting path between the power transmitting member and the drive wheel.

In the drive system disclosed in the above-identified publication JP-2004-116735A, the first electric motor and the second electric motor are isolated from each other by a partition wall, and the planetary gear set is disposed between the first electric motor and the partition wall. The partition wall has hydraulic passages which are commonly used for supplying a lubricating oil for the first electric motor and the planetary gear set which are disposed on the front side of the partition wall, and for the second electric motor disposed on the rear side of the partition wall. JP-7-76229A discloses prior art alternative to that of JP-2004-116735A.

The drive system including the first and second electric motors and the planetary gear set, as described above, may further include a transmission device or other power transmitting device that should also be lubricated. In this drive system, the lubricating oil is supplied from the partition wall between the first and second electric motors, to the first electric motor, second electric motor and differential mechanism, while an lubricating oil is supplied to the above-indicated power transmitting device from hydraulic passages provided in addition to the hydraulic passages formed through the partition wall, as disclosed in the above-identified publication JP-2004-116735A, whereby there is a risk of complexity in the arrangement of the lubricating hydraulic passages.

The present invention was made in view of the background art described above. It is therefore an object of this invention to provide a vehicular drive system which is simple in the arrangement of the lubricating hydraulic passages.

SUMMARY OF THE INVENTION

The object indicated above may be achieved according to the present invention, which provides a vehicular drive system including (a) a first electric motor, (b) a differential mechanism operable to distribute an output of a drive power source to the first electric motor and a power transmitting member, (c) a second electric motor disposed in a power transmitting path between the power transmitting member and a drive wheel, and (d) a power transmitting device disposed between the second electric motor and the drive wheel, the first electric motor, the differential mechanism, the second electric motor and the power transmitting device being arranged in an axial direction of the vehicular drive system, the vehicular drive system being characterized by a support wall disposed between the second electric motor and the power transmitting device and arranged to support the second electric motor, and a first group of oil passages which are formed through the support wall and through which a lubricating oil is supplied to at least one of the first electric motor, the differential mechanism and the second electric motor, and to the power transmitting device.

In the drive system according to the present invention, the lubricating oil is supplied from the support wall between the second electric motor and the power transmitting device, to the devices on the opposite axial sides of the support wall. Accordingly, the arrangement of the lubricating oil passages can be made simpler in the present drive system, than in a drive system wherein two groups of lubricating oil passages are provided for the respective two groups of devices disposed on the respective opposite axial sides of the support wall.

Preferably, an input shaft of said power transmitting device extends through a rotor of the second electric motor in an axial direction of the rotor and is fitted to an input shaft of the differential mechanism, and the input shaft of the power transmitting device has a second group of oil passages to which the lubricating oil is supplied from the first group of oil passages, while the input shaft of the differential mechanism has a third group of oil passages to which the lubricating oil is supplied from the second group of oil passages.

Preferably, the differential mechanism has a fourth groups of oil passages held in communication with the third group of oil passages formed through the input shaft of the differential mechanism.

Preferably, the power transmitting device includes a transmission.

Preferably, the first group of oil passages include an oil passage open in an inner circumferential surface of the support wall. In this case, the radial position at which the first group of oil passages is open toward the second group of oil passages is made close to the axis of the power transmitting device, and the radial position at which the second group of oil passages is open toward the first group of oil passages is made close to the axis of the input shaft of the power transmitting device, so that the diameter of sealing rings disposed adjacent to the second group of oil passages can be reduced, whereby the amount of a dragging power loss due to friction of the input shaft of the power transmitting device with respect to the sealing rings during rotation of the input shaft can be reduced.

Preferably, the second group of oil passages includes a 2-1 oil passage which extends in a radial direction of the input shaft of the power transmitting device and which is open at an axial position of the input shaft of the power transmitting device at which the first group of oil passages is open toward the 2-1 oil passage.

Preferably, the second group of oil passages includes a 2-2 oil passage which extends in an axial direction of the input shaft of the power transmitting device and which is held in communication with the 2-1 oil passage, the 2-2 oil passage being open at one end thereof in an end face of the input shaft of the power transmitting device, which end face is located on the side of the input shaft of the differential mechanism.

Preferably, the second group of oil passages includes a 2-3 oil passage which extends in an axial direction of the input shaft of the power transmitting device and which is held in communication with the 2-1 oil passage, the 2-3 oil passage being open at one end thereof in an end face of the input shaft of the power transmitting device, which end face is remote from the input shaft of the differential mechanism.

Preferably, a rotor support shaft supporting the rotor of the second electric motor, and the input shaft of the power transmitting device are fitted on each other through a spline, and the second group of oil passages includes a 2-4 oil passage which extends in the radial direction of the input shaft of the power transmitting device and which is held in communication with the 2-2 oil passage, the 2-4 oil passage being open in an axial portion of an outer circumferential surface of the input shaft of the power transmitting device in which the spline is formed.

Preferably, a bearing is interposed between a rotor support shaft supporting the rotor of the second electric motor, and the input shaft of the power transmitting device, and the second group of oil passages includes a 2-5 oil passage which extends in the radial direction of the input shaft of the power transmitting device and which is held in communication with the 2-2 oil passage, the 2-5 oil passage being open in an axial portion of an outer circumferential surface of the input shaft of the power transmitting device at which the bearing is located.

Preferably, the second group of oil passages includes a 2-6 oil passage which is located radially inwardly of the differential mechanism and which extends in the radial direction of the input shaft of the power transmitting device, the 2-6 oil passage being held in communication with the 2-2 oil passage.

Preferably, the third group of oil passages includes a 3-1 oil passage which extends in an axial direction of the input shaft of the differential mechanism and which is open at one end thereof toward the 2-2 oil passage.

Preferably, a bearing is interposed between a rotor support shaft supporting the rotor of the second electric motor, and the input shaft of the differential mechanism, and the third group of oil passages includes a 3-2 oil passage which extends in a radial direction of the input shaft of the differential mechanism and which is held in communication with the 3-1 oil passage, the 3-2 oil passage being open in an axial portion of an outer circumferential surface of the input shaft of the differential mechanism at which the bearing is located.

Preferably, an axial end portion of the input shaft of the differential mechanism which is located on the side of the power transmitting member is formed integrally with a portion of the differential mechanism, and the third group of oil passages includes a 3-3 oil passage which is formed through the portion of the differential mechanism, so as to extend in a radial direction of the differential mechanism, the 3-3 oil passage being held in communication with the 2-6 oil passage.

Preferably, the differential mechanism includes a planetary gear set, and the fourth group of oil passages includes a 4-1 oil passage which is formed through a pinion shaft of the planetary gear set, so as to extend in a radial direction of the pinion shaft. the 4-1 oil passage being held in communication with the 3-3 oil passage, the fourth group of oil passages further including a 4-2 oil passage which is formed through the pinion shaft, so as to extend in an axial direction of the pinion shaft, the 4-2 oil passage being held in communication with the 4-1 oil passage.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

Referring to the drawings, there will be described in detail the embodiments of the present invention.

Referring first to the schematic view ofFIG. 1, there is shown a drive system10for a hybrid vehicle, which is constructed according to one embodiment of this invention. The drive system10shown inFIG. 1includes: an input rotary member in the form of a differential mechanism input shaft14disposed on a common axis in a transmission casing12(hereinafter abbreviated as “casing12”) functioning as a stationary member or non-rotary member attached to a body of the vehicle; a power distributing mechanism16connected to this differential mechanism input shaft14either directly, or indirectly via a pulsation absorbing damper (vibration damping device) not shown; a power transmitting device in the form of a step variable automatic transmission20disposed between the power distributing mechanism16and a drive system output shaft22, such that the automatic transmission20is connected in series to the power distributing mechanism16through a power transmitting member18; and an output rotary member in the form of the drive system output shaft22connected to the automatic transmission20.

This drive system10is suitably used for a transverse FR vehicle (front-engine, rear-drive vehicle), and is disposed between a drive power source in the form of an engine8and a pair of drive wheels38, to transmit a vehicle drive force to the pair of drive wheels38through a differential gear device (final speed reduction gear)36and a pair of drive axles, as shown inFIG. 7. It is noted that a lower half of the drive system10, which is constructed symmetrically with respect to its axis, is omitted inFIG. 1.

The differential mechanism input shaft14is connected at its one end to the engine8, and the power distributing mechanism16is a mechanism arranged to mechanically synthesize an output of the engine8received from the differential mechanism input shaft14, or to mechanically distribute the output of the engine8. That is, the power distributing mechanism16distributes the output of the engine8to a first electric motor M1and the power transmitting member18, or synthesizes the output of the engine8and the output of the first electric motor M1and transmits a sum of these outputs to the power transmitting member18. In the present embodiment, each of the first electric motor M1and a second electric motor M2is a so-called motor/generator functioning as an electric generator as well as an electric motor. The first electric motor M1should function at least as an electric generator operable to generate an electric energy while generating a reaction force, and the second electric motor M2should function at least as an electric motor operable to generate a vehicle drive force.

The power distributing mechanism16includes a first planetary gear set24of single pinion type functioning as the differential mechanism, a switching clutch C0and a switching brake B0. The first planetary gear set24has rotary elements consisting of a first sun gear S1, a first planetary gear P1; a first carrier CA1supporting the first planetary gear P1such that the first planetary gear P1is rotatable about its axis and about the axis of the first sun gear S1; and a first ring gear R1meshing with the first sun gear S1through the first planetary gear P1. The first planetary gear set24has a gear ratio ρ1of about 0.418, for example. Where the numbers of teeth of the first sun gear S1and the first ring gear R1are represented by ZS1and ZR1, respectively, the above-indicated gear ratio ρ1is represented by ZS1/ZR1.

In the power distributing mechanism16, the first carrier CA1is connected to the differential mechanism input shaft14, that is, to the engine8, and the first sun gear S1is connected to the first electric motor M1, while the first ring gear R1is connected to the power transmitting member18. The switching brake B0is disposed between the first sun gear S1and the casing12, and the switching clutch C0is disposed between the first sun gear S1and the first carrier CA1. When the switching clutch C0and brake B0are both released, the power distributing mechanism16is placed in a differential state in which the first sun gear S1, first carrier CA1and first ring gear R1are rotatable relative to each other, so as to perform a differential function, so that the output of the engine8is distributed to the first electric motor M1and the power transmitting member18, whereby a portion of the output of the engine8which is distributed to the first electric motor M1is used to drive the first electric motor M1to generate an electric energy which is stored or used to drive the second electric motor M2. Accordingly, the power distributing mechanism16is placed in the continuously-variable shifting state in which the rotating speed of the power transmitting member18is continuously variable, irrespective of the rotating speed of the engine8, namely, in the differential state or continuously-variable shifting state in which the power distributing mechanism16functions as an electrically controlled continuously variable transmission whose speed ratio γ0(rotating speed of the differential mechanism input shaft14/rotating speed of the power transmitting member18) is continuously variable from a minimum value γ0min to a maximum value γ0max.

When the switching clutch C0is engaged during running of the vehicle by the output of the engine8while the power distributing mechanism16is placed in the continuously-variable shifting state, the first sun gear S1and the first carrier CA1are connected together, so that the power distributing mechanism16is brought into a locked state or non-differential state in which the three rotary elements of the first planetary gear set24consisting of the first sun gear S1, first carrier CA1and first ring gear R1are rotatable as a unit. In this non-differential state in which the rotating speed of the engine8and the rotating speed of the power transmitting member18are made equal to each other, the power distributing mechanism is placed in a fixed-speed-ratio shifting state in which the power distributing mechanism16functions as a transmission having a fixed speed ratio γ0equal to 1. When the switching brake B0is engaged in place of the switching clutch C0, the power distributing mechanism16is placed in the locked or non-differential state in which the first sun gear S1is not rotatable, so that the rotating speed of the first ring gear R1is made higher than that of the first carrier CA1, whereby the power distributing mechanism16is placed in the fixed-speed-ratio shifting state in which the power distributing mechanism16functions as a speed-increasing transmission having a fixed speed ratio γ0smaller than 1, for example, about 0.7. In the present embodiment described above, the switching clutch C0and brake B0function as a differential-state switching device operable to selectively place the first planetary gear set24in the differential state (continuously-variable shifting state) in which the first planetary gear set24functions as an electrically controlled continuously variable transmission the speed ratio of which is continuously variable, and in the non-differential state, namely, in the locked state in which the first planetary gear set24does not function as the electrically controlled continuously variable transmission having the continuously-variable shifting function, that is, in the fixed-speed-ratio shifting state in which the first planetary gear set24functions as a transmission having a single gear position with one speed ratio or a plurality of gear positions with respective speed ratios.

The automatic transmission20includes a plurality of planetary gear sets, that is, a single-pinion type second planetary gear set26, a single-pinion type third planetary gear set28and a single-pinion type fourth planetary gear set30. The second planetary gear set26has: a second sun gear S2; a second planetary gear P2; a second carrier CA2supporting the second planetary gear P2such that the second planetary gear P2is rotatable about its axis and about the axis of the second sun gear S2; and a second ring gear R2meshing with the second sun gear S2through the second planetary gear P2. For example, the second planetary gear set26has a gear ratio ρ2of about 0.562. The third planetary gear set28has: a third sun gear S3; a third planetary gear P3; a third carrier CA3supporting the third planetary gear P3such that the third planetary gear P3is rotatable about its axis and about the axis of the third sun gear S3; and a third ring gear R3meshing with the third sun gear S3through the third planetary gear P3. For example, the third planetary gear set28has a gear ratio ρ3of about 0.425. The fourth planetary gear set30has: a fourth sun gear S4; a fourth planetary gear P4; a fourth carrier CA4supporting the fourth planetary gear P4such that the fourth planetary gear P4is rotatable about its axis and about the axis of the fourth sun gear S4; and a fourth ring gear R4meshing with the fourth sun gear S4through the fourth planetary gear P4. For example, the fourth planetary gear set30has a gear ratio ρ4of about 0.421. Where the numbers of teeth of the second sun gear S2, second ring gear R2, third sun gear S3, third ring gear R3, fourth sun gear S4and fourth ring gear r4are represented by ZS2, ZR2, ZS3, ZR3, ZS4and ZR4, respectively, the above-indicated gear ratios ρ2, ρ3and ρ4are represented by ZS2/ZR2. ZS3/ZR3, and ZS4/ZR4, respectively.

In the automatic transmission20, the second sun gear S2and the third sun gear S3are integrally fixed to each other as a unit, selectively connected to the power transmitting member18through a second clutch C2, and selectively fixed to the casing12through a first brake B1. The fourth ring gear R4is selectively fixed to the casing12through a third brake B3, and the second ring gear R2, third carrier CA3and fourth carrier CA4are integrally fixed to each other and fixed to the output shaft22. The third ring gear R3and the fourth sun gear S4are integrally fixed to each other and selectively connected to the power transmitting member18through a first clutch C1.

The above-described switching clutch C0, first clutch C1, second clutch C2, switching brake B0, first brake B1, second brake B2and third brake B3are hydraulically operated frictional coupling devices used in a conventional vehicular automatic transmission. Each of these frictional coupling devices is constituted by a wet-type multiple-disc clutch including a plurality of friction plates which are superposed on each other and which are forced against each other by a hydraulic actuator, or a band brake including a rotary drum and one band or two bands which is/are wound on the outer circumferential surface of the rotary drum and tightened at one end by a hydraulic actuator. Each of the clutches C0-C2and brakes B0-B3is selectively engaged for connecting two members between which each clutch or brake is interposed.

In the drive system10constructed as described above, one of a first-gear position (first-speed position) through a fifth-gear position (fifth-speed position), a reverse-gear position (rear-drive position) and a neural position is selectively established by engaging actions of a corresponding combination of the frictional coupling devices selected from the above-described switching clutch C0, first clutch C1, second clutch C2, switching brake B0, first brake B1, second brake B2and third brake B3, as indicated in the table ofFIG. 2. Those gear positions have respective speed ratios γ (input shaft speed NIN/output shaft speed NOUT) which change as geometric series. In particular, it is noted that the power distributing mechanism16is provided with the switching clutch C0and brake B0, so that the power distributing mechanism16can be selectively placed by engagement of the switching clutch C0or switching brake B0, in the fixed-speed-ratio shifting state in which the power distributing mechanism16is operable as a transmission having a single gear position with one speed ratio or a plurality of gear positions with respective speed ratios, as well as in the continuously-variable shifting state in which the power distributing mechanism16is operable as a continuously variable transmission, as described above. In the present drive system10, therefore, a step-variable transmission is constituted by the automatic transmission20, and the power distributing mechanism16which is placed in the fixed-speed-ratio shifting state by engagement of the switching clutch C0or switching brake B0. Further, a continuously variable transmission is constituted by the automatic transmission20, and the power distributing mechanism16which is placed in the continuously-variable shifting state, with none of the switching clutch C0and brake B0being engaged.

Where the drive system10functions as the step-variable transmission, for example, the first-gear position having the highest speed ratio γ1of about 3.357, for example, is established by engaging actions of the switching clutch C0, first clutch C1and third brake B3, and the second-gear position having the speed ratio γ2of about 2.180, for example, which is lower than the speed ratio γ1, is established by engaging actions of the switching clutch C0, first clutch C1and second brake B2, as indicated inFIG. 2. Further, the third-gear position having the speed ratio γ3of about 1.424, for example, which is lower than the speed ratio γ2, is established by engaging actions of the switching clutch C0, first clutch C1and first brake B1, and the fourth-gear position having the speed ratio γ4of about 1.000, for example, which is lower than the speed ratio γ3, is established by engaging actions of the switching clutch C0, first clutch C1and second clutch C2. The fifth-gear position having the speed ratio γ5of about 0705, for example, which is smaller than the speed ratio γ4, is established by engaging actions of the first clutch C1, second clutch C2and switching brake B0. Further, the reverse-gear position having the speed ratio γR of about 3.209, for example, which is intermediate between the speed ratios γ1and γ2, is established by engaging actions of the second clutch C2and the third brake B3. The neutral position N is established by engaging only the switching clutch C0.

Where the drive system10functions as the continuously-variable transmission, on the other hand, the switching clutch C0and the switching brake B0are both released, as indicated inFIG. 2, so that the power distributing mechanism16functions as the continuously variable transmission, while the automatic transmission20connected in series to the power distributing mechanism16functions as the step-variable transmission, whereby the speed of the rotary motion transmitted to the automatic transmission20placed in one of the first-gear, second-gear, third-gear and fourth-gear positions, namely, the rotating speed of the power transmitting member18is continuously changed, so that the speed ratio when the automatic transmission20is placed in one of those gear positions is continuously variable over a predetermined range. Accordingly, the speed ratio of the automatic transmission20is continuously variable across the adjacent gear positions, whereby the overall speed ratio γT of the drive system10is continuously variable.

The collinear chart ofFIG. 3indicates, by straight lines, a relationship among the rotating speeds of the rotary elements in each of the gear positions of the drive system10, which is constituted by the power distributing mechanism16functioning as the continuously-variable shifting portion or first shifting portion, and the automatic transmission20functioning as the step-variable shifting portion or second shifting portion. The collinear chart ofFIG. 3is a rectangular two-dimensional coordinate system in which the gear ratios ρ of the planetary gear sets24,26,28,30are taken along the horizontal axis, while the relative rotating speeds of the rotary elements are taken along the vertical axis. A lower one of three horizontal lines X1, X2, XG, that is, the horizontal line X1indicates the rotating speed of 0, while an upper one of the three horizontal lines, that is, the horizontal line X2indicates the rotating speed of 1.0, that is, an operating speed NEof the engine8connected to the input shaft14. The horizontal line XG indicates the rotating speed of the power transmitting member18. Three vertical lines Y1, Y2and Y3correspond to three elements of the power distributing mechanism16, and respectively represent the relative rotating speeds of a second rotary element (second element) RE2in the form of the first sun gear S1, a first rotary element (first element) RE1in the form of the first carrier CA1, and a third rotary element (third element) RE3in the form of the first ring gear R1. The distances between the adjacent ones of the vertical lines Y1, Y2and Y3are determined by the gear ratio ρ1of the first planetary gear set24. That is, the distance between the vertical lines Y1and Y2corresponds to “1”, while the distance between the vertical lines Y2and Y3corresponds to the gear ratio ρ1. Further, five vertical lines Y4, Y5, Y6, Y7and Y8corresponding to the automatic transmission20respectively represent the relative rotating speeds of a fourth rotary element (fourth element) RE4in the form of the second and third sun gears S2, S3integrally fixed to each other, a fifth rotary element (fifth element) RE5in the form of the second carrier CA2, a sixth rotary element (sixth element) RE6in the form of the fourth ring gear R4, a seventh rotary element (seventh element) RE7in the form of the second ring gear R2and third and fourth carriers CA3, CA4that are integrally fixed to each other, and an eighth rotary element (eighth element) RE8in the form of the third ring gear R3and fourth sun gear S4integrally fixed to each other. The distances between the adjacent ones of the vertical lines Y4-Y8are determined by the gear ratios ρ2, ρ3and ρ4of the second, third and fourth planetary gear sets26,28,30. Therefore, as shown inFIG. 3, the distance between the vertical lines corresponding to the sun gear and carrier of each of the second, third and fourth planetary gear sets26,28,30corresponds to “1”, while the distance between the vertical lines corresponding to the carrier and ring gear corresponds to the gear ratio ρ.

Referring to the collinear chart ofFIG. 3, the power distributing mechanism16(continuously-variable transmission portion) of the drive system10is arranged such that the first rotary element RE1(first carrier CA1) of the first planetary gear set24, is integrally fixed to the input shaft14, that is, to the engine8, and is selectively connected to the second rotary element RE2(first sun gear S1) through the switching clutch C0, and this rotary element RE2is connected to the first electric motor M1and selectively fixed to the casing12through the switching brake B0, while the third rotary element RE3(first ring gear R1) is fixed to the power transmitting member18and connected to the second electric motor M2, so that a rotary motion of the differential mechanism input shaft14is transmitted to the automatic transmission (step-variable transmission portion)20through the power transmitting member18. A relationship between the rotating speeds of the first sun gear S1and the first ring gear R1is represented by an inclined straight line L0which passes a point of intersection between the lines Y2and X2.

FIGS. 4 and 5correspond to a part of the collinear chart ofFIG. 3which shows the power distributing mechanism16.FIG. 4shows an example of an operating state of the power distributing mechanism16placed in the continuously-variable shifting state with the switching clutch C0and the switching brake B0held in the released state. The rotating speed of the first sun gear S1represented by the point of intersection between the straight line L0and vertical line Y1is raised or lowered by controlling the reaction force generated by an operation of the first electric motor M1to generate an electric energy, so that the rotating speed of the first ring gear R1represented by the point of intersection between the lines L0and Y3is lowered or raised. In the state shown inFIG. 4, the first sun gear S1has a negative rotating speed, that is, the first electric motor M1is operated with an electric power supplied thereto. In this state in which the first sun gear S1has the negative rotating speed, the straight line L0has a relatively large angle of inclination, so that the first ring gear R1and the power transmitting member18connected to the first ring gear R1have relatively high rotating speeds, thereby permitting the vehicle to run at a relatively high speed, but deteriorating the fuel economy of the vehicle by an amount corresponding to the amount of electric power supplied to and consumed by the first electric motor M1. In the present drive system10, however, the automatic transmission20is arranged to raise the input rotating speed received from the power transmitting member18, so that there is a relatively low opportunity in which the first sun gear S1should have a negative rotating speed. Accordingly, the fuel economy can be improved in the present arrangement, than in the case where the automatic transmission20were not able to raise the rotating speed of the power transmitting member18.

FIG. 5shows an operating state of the power distributing mechanism16placed in the step-variable shifting state with the switching clutch C0held in the engaged state. When the first sun gear S1and the first carrier CA1are connected to each other, the three rotary elements indicated above are rotated as a unit, so that the straight line L0is aligned with the horizontal line X2, whereby the power transmitting member18is rotated at a speed equal to the engine speed NE. When the switching brake B0is engaged, on the other hand, the rotation of the first sun gear S1is stopped, so that the straight line L0is inclined in the state indicated inFIG. 3, whereby the rotating speed of the first ring gear R1, that is, the rotation of the power transmitting member18represented by a point of intersection between the lines L0and Y3is made higher than the engine speed NEand transmitted to the automatic transmission20.

In the automatic transmission20, the fourth rotary element RE4is selectively connected to the power transmitting member18through the second clutch C2, and selectively fixed to the casing12through the first brake B1, and the fifth rotary element RE5is selectively fixed to the casing12through the second brake B2, while the sixth rotary element RE6is selectively fixed to the casing12through the third brake B3. The seventh rotary element RE7is integrally fixed to the drive system output shaft22, while the eighth rotary element RE8is selectively connected to the power transmitting member18through the first clutch C1.

When the first clutch C1and the third brake B3are engaged, the automatic transmission20is placed in the first-speed position. The rotating speed of the drive system output shaft22in the first-speed position is represented by a point of intersection between the vertical line Y7indicative of the rotating speed of the seventh rotary element RE7fixed to the drive system output shaft22and an inclined straight line L1which passes a point of intersection between the vertical line Y8indicative of the rotating speed of the eighth rotary element RE8and the horizontal line X2, and a point of intersection between the vertical line Y6indicative of the rotating speed of the sixth rotary element RE6and the horizontal line X1. Similarly, the rotating speed of the drive system output shaft22in the second-speed position established by the engaging actions of the first clutch C1and second brake B2is represented by a point of intersection between an inclined straight line L2determined by those engaging actions and the vertical line Y7indicative of the rotating speed of the seventh rotary element RE7fixed to the drive system output shaft22. The rotating speed of the drive system output shaft22in the third-speed position established by the engaging actions of the first clutch C1and first brake B1is represented by a point of intersection between an inclined straight line L3determined by those engaging actions and the vertical line Y7indicative of the rotating speed of the seventh rotary element RE7fixed to the output shaft22. The rotating speed of the drive system output shaft22in the fourth-speed position established by the engaging actions of the first clutch C1and second clutch C2is represented by a point of intersection between a horizontal line L4determined by those engaging actions and the vertical line Y7indicative of the rotating speed of the seventh rotary element RE7fixed to the drive system output shaft22. In the first-speed through fourth-gear positions in which the switching clutch C0is placed in the engaged state, the eighth rotary element RE8is rotated at the same speed as the engine speed NE, with the drive force received from the power distributing mechanism16, that is, from the power distributing mechanism16. When the switching brake B0is engaged in place of the switching clutch C0, the eighth rotary element RE8is rotated at a speed higher than the engine speed NE, with the drive force received from the power distributing mechanism16. The rotating speed of the output shaft22in the fifth-speed position established by the engaging actions of the first clutch C1, second clutch C2and switching brake B0is represented by a point of intersection between a horizontal line L5determined by those engaging actions and the vertical line Y7indicative of the rotating speed of the seventh rotary element RE7fixed to the output shaft22.

FIG. 6illustrates signals received by an electronic control device40provided to control the drive system10, and signals generated by the electronic control device40. This electronic control device40includes a so-called microcomputer incorporating a CPU, a ROM, a RAM and an input/output interface, and is arranged to process the signals according to programs stored in the ROM while utilizing a temporary data storage function of the ROM, to implement hybrid drive controls of the engine8and electric motors M1and M2, and drive controls such as a shifting control of the automatic transmission20.

The electronic control device40is arranged to receive, from various sensors and switches shown inFIG. 6, various signals such as: a signal indicative of a temperature of cooling water of the engine; a signal indicative of a selected operating position of a shift lever; a signal indicative of the operating speed NEof the engine8; a signal indicative of a value indicating a selected group of forward-drive positions of the transmission mechanism; a signal indicative of an M mode (motor-drive mode); a signal indicative of an operated state of an air conditioner; a signal indicative of a vehicle speed corresponding to the rotating speed of the drive system output shaft22; a signal indicative of a temperature of a working oil of the automatic transmission20; a signal indicative of an operated state of a side brake; a signal indicative of an operated state of a foot brake; a signal indicative of a temperature of a catalyst; a signal indicative of an angle of operation of an accelerator pedal; a signal indicative of an angle of a cam; a signal indicative of the selection of a snow drive mode; a signal indicative of a longitudinal acceleration value of the vehicle; a signal indicative of the selection of an auto-cruising drive mode; a signal indicative of a weight of the vehicle; signals indicative of speeds of the drive wheels of the vehicle; a signal indicative of an operating state of a step-variable shifting switch provided to place the power distributing mechanism16in the fixed-speed-ratio shifting state in which the drive system10functions as a step-variable transmission; a signal indicative of a continuously-variable shifting switch provided to place the power distributing mechanism16in the continuously variable-shifting state in which the drive system10functions as the continuously variable transmission; a signal indicative of a rotating speed NM1of the first electric motor M1; and a signal indicative of a rotating speed NM2of the second electric motor M2. The electronic control device40is further arranged to generate various signals such as: a signal to drive an electronic throttle actuator for controlling an angle of opening of a throttle valve; a signal to adjust a pressure of a supercharger; a signal to operate the electric air conditioner; a signal for controlling an ignition timing of the engine8; signals to operate the electric motors M1and M2; a signal to operate a shift-range indicator for indicating the selected operating position of the shift lever; a signal to operate a gear-ratio indicator for indicating the gear ratio; a signal to operate a snow-mode indicator for indicating the selection of the snow drive mode; a signal to operate an ABS actuator for anti-lock braking of the wheels; a signal to operate an M-mode indicator for indicating the selection of the M-mode; signals to operate solenoid-operated valves incorporated in a hydraulic control unit42provided to control the hydraulic actuators of the hydraulically operated frictional coupling devices of the power distributing mechanism16and the automatic transmission20; a signal to operate an electric oil pump used as a hydraulic pressure source for the hydraulic control unit42; a signal to drive an electric heater; and a signal to be applied to a cruise-control computer.

FIG. 7is a functional block diagram illustrating major control functions performed by the electronic control device40. Switching control means50is arranged to determine whether the vehicle condition is in a continuously-variable shifting region in which the drive system10should be placed in the continuously-variable shifting state, or in a step-variable shifting region in which the drive system10should be placed in the step-variable shifting state. This determination is made on the basis of a stored predetermined relationship shown inFIG. 8or9, for example. Where the relationship shown inFIG. 8(switching data map) is used, the determination is made on the basis of the vehicle condition as represented by the actual engine speed NE, and a drive-force-related value relating to the drive force of the hybrid vehicle, for example, an engine output torque TE.

According to the relationship shown inFIG. 8, the step-variable shifting region is set to be a high-torque region (a high-output running region in which the output torque TEof the engine8is not lower than a predetermined value TE1, or a high-speed region in which the engine speed NEis not lower than a predetermined value NE1, namely, a high-vehicle-speed region in which the vehicle speed which is one of the vehicle conditions and which is determined by the engine speed NE and the overall speed ratio γT is not lower than a predetermined value, or a high-output region in which the vehicle output calculated from the output torque TEand speed NEof the engine8is not lower than a predetermined value. Accordingly, the step-variable shifting control is effected when the vehicle is running with a comparatively high output torque or speed of the engine8, or with a comparatively high vehicle output. The step-variable shifting control permits a change of the engine speed NEas a result of a shift-up action of the transmission, that is, a rhythmic change of the speed of the engine8. Namely, the continuously-variable shifting state is switched to the step-variable shifting state (fixed-speed-ratio shifting state) when the vehicle is placed in a high-output running state in which a desire of the vehicle operator to increase the vehicle drive force should be satisfied rather a desired to improve the fuel economy. Accordingly, the vehicle operator can enjoy a comfortable rhythmic change of the engine speed NE. On the other hand, the continuously-variable shifting control is effected when the vehicle is running with a comparatively low output torque or speed of the engine8, or with a comparatively low vehicle output, that is, when the engine8is a normal output state. A boundary line defining the step-variable shifting region and the continuously-variable shifting region inFIG. 8corresponds to a high-vehicle speed determining line defined by a series of high-vehicle-speed upper limit values, or a high-output running determining line defined by a series of high-output upper limit values.

When the relationship shown inFIG. 9is used, the above-indicated determination is made on the basis of the actual vehicle speed V and the drive-force-related value in the form of the output torque TOUT. InFIG. 9, a broken line indicates a threshold vehicle speed V1and a threshold output torque T1which define a predetermined vehicle condition used for switching from the continuously-variable shifting control to the step-variable shifting control, and two-dot chain line indicates a predetermined vehicle condition used for switching from the step-variable shifting control to the continuously-variable shifting control. Thus, there is provided a hysteresis for determination as to whether the shifting state should be switched between the step-variable shifting region and the continuously-variable shifting region. InFIG. 9, a solid line51indicates a boundary line defining a motor drive region in which the vehicle is driven by a drive force generated by the electric motor, with a relatively low vehicle output torque or at a relatively low vehicle speed.FIG. 9also shows a shift boundary data map which uses control parameters in the form of the vehicle speed V and the output torqueTOUT.

When the switching control means50determines that the vehicle condition is in the step-variable shifting region, the switching control means50disables a hybrid control means52to effect a hybrid control or continuously-variable shifting control, and enables a step-variable shifting control means54to effect a predetermined step-variable shifting control. Where the step-variable shifting control means54effects the step-variable shifting control according to the determination made on the basis of the relationship ofFIG. 8, the step-variable shifting control means54effects an automatic shifting control according to a stored predetermined shift boundary data map. Where the determination is made on the basis of the relationship ofFIG. 9, the automatic shifting control is effected according to the shift boundary data map shown inFIG. 9.

FIG. 2indicates the combinations of the operating states of the hydraulically operated frictional coupling devices C0, C1, C2, B0, B1, B2and B3, which are selectively engaged for effecting the step-variable shifting control. In this automatic step-variable shifting control mode, the first-speed through fourth-speed positions are established by an engaging action of the switching clutch C0, and the power distributing mechanism16functions as an auxiliary transmission having a fixed speed ratio of γ0equal to “1”. On the other hand, the fifth-speed position is established by an engaging action of the switching brake B0in place of the switching clutch C0, and the power distributing mechanism16functions as an auxiliary transmission having a fixed speed ratio γ0equal to about 0.7, for example. That is, the drive system10as a whole including the power distributing mechanism16functioning as the auxiliary transmission and the automatic transmission20functions as a so-called “automatic transmission”, in the automatic step-variable shifting control mode.

The drive-force-related value indicated above is a parameter corresponding to the drive force of the vehicle, which may be an output torque TOUTof the automatic transmission20, an engine output torque TE, or an acceleration value of the vehicle, as well as a drive torque or drive force of drive wheels38. The engine output torque TEmay be an actual value calculated on the basis of the operating angle of the accelerator pedal or the opening angle of the throttle valve (or intake air quantity, air/fuel ratio or amount of fuel injection) and the engine speed NE, or an estimated value of the required vehicle drive force which is calculated on the basis of the amount of operation of the accelerator pedal by the vehicle operator or the operating angle of the throttle valve. The vehicle drive torque may be calculated on the basis of not only the output torque TOUT, etc., but also the ratio of a differential gear device and the radius of the drive wheels38, or may be directly detected by a torque sensor or the like.

When the switching control means50determines that the vehicle condition is in the continuously-variable shifting region, on the other hand, the switching control means50commands the hydraulic control unit42to release both of the switching clutch C0and the switching brake B0for placing the power distributing mechanism16in the electrically established continuously-variable shifting state. At the same time, the switching control means50enables the hybrid control means52to effect the hybrid control, and commands the step-variable shifting control means54to select and hold a predetermined one of the gear positions, or to permit an automatic shifting control according to the stored predetermined shift boundary data map. In the latter case, the variable-step shifting control means54effects the automatic shifting control by suitably selecting the combinations of the operating states of the frictional coupling devices indicated in the table ofFIG. 2, except the combinations including the engagement of the switching clutch C0and brake B0. Thus, the power distributing mechanism16placed in the continuously-variable shifting state under the control of the switching control means50functions as the continuously variable transmission while the automatic transmission20connected in series to the power distributing mechanism16functions as the step-variable transmission, so that the drive system provides a sufficient vehicle drive force, such that the speed of the rotary motion transmitted to the automatic transmission20placed in one of the first-speed, second-speed, third-speed and fourth-gear positions, namely, the rotating speed of the power transmitting member18is continuously changed, so that the speed ratio of the drive system when the automatic transmission20is placed in one of those gear positions is continuously variable over a predetermined range. Accordingly, the speed ratio of the automatic transmission20is continuously variable through the adjacent gear positions, whereby the overall speed ratio γT of the drive system10as a whole is continuously variable.

The hybrid control means52controls the engine8to be operated with high efficiency, so as to establish an optimum proportion of the drive forces which are produced by the engine8, and the first electric motor M1and/or the second electric motor M2. For instance, the hybrid control means52calculates the output as required by the vehicle operator at the present running speed V of the vehicle, on the basis of the operating amount of the accelerator pedal and the vehicle running speed, and calculate a required vehicle drive force on the basis of the calculated required output and a required amount of generation of an electric energy to be stored. On the basis of the calculated required vehicle drive force, the hybrid control means52calculates a desired engine speed and a desired total output, and controls the actual output of the engine8and the amount of generation of the electric energy by the first electric motor M1, according to the calculated desired total output and engine speed NE. The hybrid control means52is arranged to control the shifting action of the automatic transmission20, while taking account of the presently selected gear position of the automatic transmission20, so as to improve the fuel economy of the engine8. In the hybrid control, the power distributing mechanism16is controlled to function as the electrically controlled continuously-variable transmission, for optimum coordination of the engine speed NEand vehicle speed V for efficient operation of the engine8, and the rotating speed of the power transmitting member18determined by the selected gear position of the automatic transmission portion20. That is, the hybrid control means52determines a target value of the overall speed ratio γT of the transmission mechanism10so that the engine8is operated according a stored highest-fuel-economy curve that satisfies both of the desired operating efficiency and the highest fuel economy of the engine8. The hybrid control means52controls the speed ratio γ0of the differential portion11, so as to obtain the target value of the overall speed ratio γT, so that the overall speed ratio γT can be controlled within a predetermined range, for example, between 13 and 0.5.

The hybrid control means52controls an inverter58such that the electric energy generated by the first electric motor M1is supplied to an electric-energy storage device60and the second electric motor M2through the inverter58. That is, a major portion of the drive force produced by the engine8is mechanically transmitted to the power transmitting member18, while the remaining portion of the drive force is consumed by the first electric motor M1to convert this portion into the electric energy, which is supplied from the first electric motor M1to the second electric motor M2through the inverter58and consumed by the second electric motor M2, or supplied from the first electric motor M1to the electric-energy storage device60through the inverter58and subsequently consumed by the first electric motor M1. A drive force produced by an operation of the second electric motor M2or first electric motor M1with the electric energy generated by the first electric motor M1is transmitted to the power transmitting member18. Thus, the transmission mechanism10is provided with an electric path through which an electric energy generated by conversion of a portion of a drive force of the engine8is converted into a mechanical energy. This electric path includes components associated with the generation of the electric energy and the consumption of the generated electric energy by the second electric motor M2. The hybrid control means52can establish a motor-drive mode to drive the vehicle by utilizing the electric CVT function of the power distributing mechanism16, irrespective of whether the engine8is in the non-operated state or in the idling state.

In the above-described arrangements of the switching control means50, hybrid control means52and step-variable shifting control means54, the power distributing mechanism16is placed in the continuously-variable shifting state, assuring a high degree of fuel economy of the vehicle, when the vehicle is in a low- or medium-speed running state or in a low- or medium-output running state, with the engine operated in the normal output state. When the vehicle is in a high-speed running state or at a high speed of operation of the engine8, on the other hand, the power distributing mechanism16is placed in the fixed-speed-ratio shifting state in which the output of the engine8is transmitted to the drive wheels38primarily through the mechanical power transmitting path, so that the fuel economy is improved owing to reduction of a loss of conversion of the mechanical energy into the electric energy. When the engine8is in a high-output state, the power distributing mechanism16is placed in the fixed-speed-ratio shifting state. Thus, the power distributing mechanism16is placed in the continuously-variable shifting state, only when the vehicle speed or output is relatively low or medium, so that the required amount of electric energy generated by the first electric motor M1, that is, the maximum amount of electric energy that must be transmitted from the first electric motor M1can be reduced, whereby the required electrical reaction force of the first electric motor M1can be reduced, making it possible to minimize the required sizes of the first and second electric motors M1, M2, and the required size of the drive system10including the electric motors.

FIG. 10shows an example of a manually operable shifting device in the form of a shifting device46. The shifting device46includes a shift lever48, which is disposed laterally adjacent to an operator's seat, for example, and which is manually operated to select one of a plurality of positions consisting of a parking position P for placing the drive system10(namely, automatic transmission20) in a neutral state in which a power transmitting path is disconnected with both of the switching clutch C0and brake B0placed in the released state, and at the same time the drive system output shaft22of the automatic transmission20is in the locked state; a reverse-drive position R for driving the vehicle in the rearward direction; a neutral position N for placing the drive system10in the neutral state; an automatic forward-drive shifting position D; and a manual forward-drive shifting position M. The parking position P and the neutral position N are non-driving positions selected when the vehicle is not driven, while the reverse-drive position R, and the automatic and manual forward-drive shifting positions D, M are driving positions selected when the vehicle is driven. The automatic forward-drive shifting position D provides a highest-speed position, and positions “4” through “L” selectable in the manual forward-drive shifting position M are engine-braking positions in which an engine brake is applied to the vehicle.

The manual forward-drive shifting position M is located at the same position as the automatic forward-drive shifting position D in the longitudinal direction of the vehicle, and is spaced from or adjacent to the automatic forward-drive shifting position D in the lateral direction of the vehicle. The shift lever48is operated to the manual forward-drive shifting position M, for manually selecting one of the positions “D” through “L”. Described in detail, the shift lever48is movable from the manual forward-drive shifting position M to a shift-up position “+” and a shift-down position “−”, which are spaced from each other in the longitudinal direction of the vehicle. Each time the shift lever92is moved to the shift-up position “+” or the shift-down position “−”, the presently selected position is changed by one position. The five positions “D” through “L” have respective different lower limits of a range in which the overall speed ratio γT of the drive system10is aut6omatically variable, that is, respective different lowest values of the overall speed ratio γT which corresponds to the highest output speed of the drive system10. Namely, the five positions “D” through “L” select respective different numbers of the speed positions or gear positions of the automatic transmission20which are automatically selectable, so that the lowest overall speed ratio γT available is determined by the selected number of the selectable gear positions. The shift lever48is biased by biasing means such as a spring so that the shift lever48is automatically returned from the shift-up position “+” and shift-down position “−” back to the manual forward-drive shifting position M. The shifting device46is provided with shift-position sensors operable to detect the presently selected position of the shift lever48, so that signals indicative of the presently selected operating position of the shift lever48and the number of shifting operations of the shift lever48in the manual forward-shifting position M are supplied to the electronic control device40.

When the shift lever46is operated to the automatic forward-drive shifting position D, the switching control means50effects an automatic switching control of the drive system10, and the hybrid control means52effects the continuously-variable shifting control of the power distributing mechanism16, while the step-variable shifting control means54effects an automatic shifting control of the automatic transmission20. When the drive system10is placed in the step-variable shifting state, for example, the shifting action of the drive system10is automatically controlled to select an appropriate one of the first-gear position through the fifth-gear position indicated inFIG. 2. When the drive system10is placed in the continuously-variable shifting state, the speed ratio of the power distributing mechanism16is continuously changed, while the shifting action of the automatic transmission20is automatically controlled to select an appropriate one of the first-gear through fourth-gear positions, so that the overall speed ratio γT of the drive system10is controlled so as to be continuously variable within the predetermined range. The automatic forward-drive position D is a position selected to establish an automatic shifting mode (automatic mode) in which the drive system10is automatically shifted.

When the shift lever48is operated to the manual forward-drive shifting position M, on the other hand, the shifting action of the drive system10is automatically controlled by the switching control means50, hybrid control means52and step-variable shifting control means54, such that the overall speed ratio γT is variable within a predetermined range the lower limit of which is determined by the gear position having the lowest speed ratio, which gear position is determined by the manually selected one of the positions “D” through “L”. When the drive system10is placed in the step-variable shifting state, for example, the shifting action of the drive system10is automatically controlled within the above-indicated predetermined range of the overall speed ratio γT. When the drive system10is placed in the continuously-variable shifting state, the speed ratio of the power distributing mechanism16is continuously changed, while the shifting action of the automatic transmission20is automatically controlled to select an appropriate one of the gear positions the number of which is determined by the manually selected one of the positions “D” through “L”, so that the overall speed ratio γT of the drive system10is controlled so as to be continuously variable within the predetermined range. The manual forward-drive position M is a position selected to establish a manual shifting mode (manual mode) in which the selectable gear positions of the drive system10are manually selected.

FIGS. 11 and 12are fragmentary cross sectional view of the drive system10, respectively. Referring first toFIG. 11, there will be briefly described an arrangement shown therein. As shown inFIG. 11, the casing12of the drive system10consists of a first casing12aaccommodating the first electric motor M1and the power distributing mechanism16, and a second casing12baccommodating the second electric motor M2and the automatic transmission20(not shown inFIG. 11). After the first casing12aand the second casing12bare fixed to each other, the first electric motor M1, the power distributing mechanism16, the second electric motor M2are arranged in this order of description in the right direction away from the engine. The first casing12acooperates with the first electric motor M1and the power distributing mechanism16which are accommodated in the first casing12a, to constitute a first unit140, while the second casing12bcooperates with the second electric motor2and the automatic transmission20which are accommodated in the second casing12b, to constitute a second unit70. The differential mechanism input shaft14, and the transmission input shaft72which is the input shaft of the automatic transmission20are disposed coaxially with the axis of the casing12, such that the differential mechanism input shaft72is located on the left side of the transmission input shaft72. The differential mechanism input shaft14is a member corresponding to the first unit140, while the transmission input shaft72is a member corresponding to the second unit70.

The first casing12ahas an integrally formed first support wall142located between the first electric motor M1and the power distributing mechanism16, while a second support wall76is fixed to the second casing12bsuch that the second support wall76is located between the power distributing mechanism16and the second electric motor M1.

Referring next toFIG. 12, there is disposed a third support wall78on one side (on the right side) of the second electric motor M2which is remote from the first electric motor M1. On this third support wall78, there is disposed the automatic transmission20such that the automatic transmission20is located on one side of the third support wall78which is remote from the second electric motor M2. One end portion of an intermediate shaft80disposed coaxially with the differential mechanism input shaft14and the transmission input shaft72is fitted in one end portion of the transmission input shaft72which is remote from the differential mechanism input shaft14, such that the intermediate shaft80and the transmission input shaft72are rotatable relative to each other. The intermediate shaft80is connected at the other end (not shown) to the drive system output shaft22(not shown inFIG. 12).

The third support wall78consists of an inner cylindrical portion78acoaxial with the transmission input shaft72; a connecting portion78bwhich is fixed at its radially inner end to the axial end of the inner cylindrical portion78aon the side of the second electric motor M2and which extends radially outwardly from the inner cylindrical portion78a; an outer cylindrical portion78cwhich is connected at its one axial end to the radially outer end of the connecting portion78band which axially extends toward the second electric motor M2and has a comparatively large radial wall thickness; and a protruding portion78dwhich extends toward the second electric motor M2from a relatively radially inner portion of the side surface of the connecting portion78bthat faces the second electric motor M2. The third support wall78and the second casing12bconstitute a faucet joint. Namely, the outer circumferential surface of the outer cylindrical portion78cis held in abutting contact with a first abutting surface82of the inner circumferential surface of the second casing12b. Before the third support wall78is fixed to the second casing12bby bolts84, the outer cylindrical portion78cis slidable at its outer circumferential surface on the first abutting surface82, so that the third support wall78can be fitted into the second casing12b, without a press fit. The third support wall78is fitted into the second casing12bafter the intermediate shaft80, components of the automatic transmission20and the transmission input shaft72are assembled into the second casing12b. Subsequently, the third support wall78is fixed to the second casing12bby the bolts84, together with a stator85of the second electric motor M2.

The outer cylindrical portion78chas an axial end face remote from the second electric motor M2. This axial end face is held in abutting contact with a first radial surface86of the second casing12b, which first radial surface86extends radially inwardly from the axial end of the first abutting surface82which is remote from the second electric motor M2. Accordingly, the third support wall78can be accurately positioned in its axial and radial directions, by simply fitting the third support wall78into the second casing12buntil the outer circumferential surface and the above-indicated axial end face of the outer cylindrical portion78ccome into in abutting contact with the first abutting surface82and the first radial surface86, respectively. The axial end portion of the transmission input shaft72is supported at its axial end portion on the side of the intermediate shaft80, by the inner cylindrical portion78aof the third support wall78via a bearing88fitted in the axial end portion of the inner cylindrical portion78aremote from the second electric motor M2, such that the transmission input shaft72is rotatable relative to the inner cylindrical portion78aof the third support wall78. An inner sleeve89is fitted in the inner cylindrical portion78a.

A rotor support shaft90of the second electric motor M2is supported by the third support wall78at its axial end portion on the side of the third support wall78, via a bearing92fitted in the protruding portion78dof the third support wall78. This rotor support shaft90is fitted on the transmission input shaft72through a spline93, so that the rotor support shaft90is rotated with the transmission input shaft72. A stator85of the second electric motor M2is held in abutting contact with an axial end face of the outer cylindrical portion78cof the third support wall78. That is, the outer cylindrical portion78cis interposed between the stator85and the second casing12b, so that the stator85is positioned in its axial direction. Thus, the outer cylindrical portion78cof the third support wall78functions as a spacer disposed between the stator85and the second casing12b, and the third support wall78is considered to have the integrally formed spacer. The stator85and the third support wall78are fastened together to the second casing12bby the above-described bolts84extending through the stator85and the outer cylindrical portion78cof the third support wall78in the axial direction. Accordingly, the drive system10can be more easily assembled, with the reduced number of the required components, and the radial dimension of the drive system10can be made smaller, than in the case where the stator85and the third support wall78are fixed to the second casing12b, by respective two sets of bolts.

The third support wall78described above has first oil passages in the form of a 1-1 oil passage94, a 1-2 oil passage, a 1-3 oil passage and a 1-4 oil passage to which a lubricating oil is supplied from a secondary regulator valve not shown. The 1-1 oil passage94is formed through the connecting portion78b, so as to extend in the radial direction, and is held at its one end in communication with one end of the 1-2 oil passage96. The 1-2 oil passage96is held at its other end in communication with one end of the 1-3 oil passage98. The 1-3 oil passage98is formed in the inner circumferential surface of the inner cylindrical portion78a, so as to extend in the axial direction of the inner cylindrical portion78a. The 1-4 oil passage100is formed so as to extend in the radial direction, and is held at its one end in communication with the other end of the 1-3 oil passage98which is remote from the 1-2 oil passage96, and is open at the other end in the outer circumferential surface of the inner cylindrical portion78a.

An outer sleeve102is press-fitted on the outer circumferential surface of the inner cylindrical portion78a, and a clutch cylinder104of the second clutch C2is fitted on the outer circumferential surface of the outer sleeve102. A bushing107is interposed between the clutch cylinder104and the outer circumferential surface of the axial end portion of the outer sleeve102which is remote from the connecting portion78b. The clutch cylinder104accommodates a clutch piston108such that the clutch piston108and the clutch cylinder104cooperate to define an oil chamber110therebetween.

The outer sleeve102has an oil hole112formed therethrough in its radial direction such that the oil hole112is held in communication with the above-indicated 1-4 oil passage100. The lubricating oil is fed to the oil hole112through the 1-1 oil passage94, 1-2 oil passage96, 1-3 oil passage98and 1-4 oil passage100, and is further fed through an oil hole formed through the clutch cylinder104, to lubricate friction plates (not shown) of the second clutch C2. The outer sleeve102further has an oil groove114for supplying the oil chamber110with a working oil. This oil groove114is held in communication with working oil passages (not shown) formed through the third support wall78, in addition to the oil passages94,96,98,100. The clutch cylinder104has an oil hole116for communication between the oil groove114and the oil chamber110.

The above-indicated inner sleeve89is fitted on the inner cylindrical portion78aof the third support wall78, so as to partially define the 1-3 oil passage98, and has an oil hole117formed therethrough in its radial direction. the oil hole117is open at its one end in an end portion of the 1-3 oil passage98which is on the side of the second electric motor M2.

The transmission input shaft72has second oil passages in the form of a 2-1 oil passage118, a 2-2 oil passage120, a 2-3 oil passage122, a 2-4 oil passage123, a 2-5 oil passage124(shown inFIG. 11) and a 2-6 oil passage125(shown inFIG. 11), which serve as lubricating oil passages. The 2-1 oil passage118is formed so as to extend in the radial direction, and is held in communication with the 1-3 oil passage98through the oil hole117formed through the inner sleeve89. The 2-2 oil passage120is formed so as to extend in the axial direction, and is held at its one end in communication with the 2-1 oil passage118and is open at its other end in the end face of the transmission input shaft72on the side of the differential mechanism input shaft14. The 2-2 oil passage120is supplied with the lubricating oil through the 1-1 oil passage94, 1-2 oil passage96, 1-3 oil passage98, oil hole117and 3-2 oil passage118. The 2-3 oil passage122is formed so as to extend in the axial direction, and is held at its one end in communication with the 2-1 oil passage118and is open at its other end in the end face of the transmission input shaft72on the side of the intermediate shaft80. The 2-4 oil passage123is formed so as to extend in the radial direction, and is held at its one end in communication with the 2-2 oil passage120and is open at its other end in a portion of the outer circumferential surface of the transmission input shaft72in which the spline93is formed. The lubricating oil fed to the 2-4 oil passage123through the 2-2 oil passage120is further fed through the spline93, to lubricate the bearing92. It is noted that although sealing rings126are fitted on the outer circumferential surface of the transmission input shaft72, at respective axial positions on the respective opposite sides of the 2-1 oil passage118, the amount of a dragging power loss of the transmission input shaft72due to friction with respect to the sealing rings126during rotation of the transmission input shaft72is comparatively small, since the sealing rings126fitted on the transmission input shaft72have a comparatively small diameter.

The transmission input shaft72further have working oil passages such as a first working oil passage127and a second working oil passage128, in addition to the above-described oil passages118,120,122,123,124and125. The first working oil passage127is formed in the axial direction, in parallel with the 2-3 oil passage122, and is open at its one end in the end face of the transmission input shaft72on the side of the intermediate shaft80, like the 2-3 oil passage122. Accordingly, the 2-3 oil passage122and the first working oil passage127can be formed simultaneously. The open end of the first working oil passage127is fluid-tightly closed by a ball129. The second working oil passage128is open at its one end in an oil chamber132formed in the back surface of a clutch piston130of the first clutch C1.

The intermediate shaft80fitted in the transmission input shaft72rotatably relative to the transmission input shaft72has an axial lubricating oil passage134which is held at its one end in communication with the open end of the 2-3 oil passage122. The intermediate shaft80further has a plurality of radial lubricating oil passages136for communication between the axial lubricating oil passage134and the outer circumferential surface of the intermediate shaft80. The lubricating oil is fed to the various components of the automatic transmission20through the above-indicated 1-1 oil passage94, 1-2 oil passage96, 1-3 oil passage98, oil hole117, 2-1 oil passage118, 2-3 oil passage122, axial lubricating oil passage134and radial lubricating oil passages136.

There will next be described in detail the construction shown inFIG. 11. The first casing12ahas a generally cylindrical outer shape. An axial portion of the first casing12awhich accommodates the power distributing mechanism16has a substantially constant outside diameter, while an axial portion of the first casing12aof the first casing12awhich accommodates the first electric motor M1increases in the axial direction toward the engine8(in the left direction as seen inFIG. 11). The first casing12aare open at its opposite axial ends, and has the above-indicated first support wall142integrally formed between the power distributing mechanism16and the first electric motor M1. The first support wall142includes an upright portion142ain the form of a disc extending perpendicularly to the differential mechanism input shaft14, a cylindrical portion142bwhich is fixed at its one axial end to the radially inner end of the upright portion142aand which extends toward the first planetary gear set24, and a protruding portion142cprotruding in the axial direction toward the first electric motor M1, from the side surface of a radially inner portion of the upright portion142a, which side surface faces the first electric motor M1. The cylindrical portion142bhas a radially central through-hole143formed therethrough in its axial direction. The interior of the first casing12ais divided by the first support wall142into a first accommodating space144which is formed on the side of the engine8and which accommodates the first electric motor M1, and a second accommodating space146which accommodates the power distributing mechanism16. The first electric motor M1is installed in the first accommodating space144through the left axial open end of the first casing12a, while the power distributing mechanism16is installed in the second accommodating space146through the right axial open end of the first casing12a, as seen inFIG. 11.

The first casing12aincludes an annular projecting portion148extending in the axial direction coaxially with the differential mechanism input shaft14, such that the projecting portion148partially defines the first accommodating space144having a substantially constant diameter. A lid plate150is fixed to the first casing12a, such that the lid plate150is held at its radially outer end in contact with the end face of the projecting portion148.

The first electric motor M1consists of a stator152, a rotor154, and a rotor support shaft (rotor hub)156formed integrally with the rotor154. The above-described first support wall142functions as a support member, and the rotor support shaft156is supported at its one axial end portion via a bearing158, by the inner circumferential surface of the protruding portion142cof the first support wall142, which is a portion of the first casing12a, such that the rotor support shaft156is rotatable relative to the protruding portion142c. The rotor support shaft156is supported at its other axial end portion by the first casing12a, through a bearing160and the lid plate150fixed to the first casing12a, such that the rotor support shaft156is rotatable relative to the first casing12a.

A sun gear shaft162formed integrally with the first sun gear S1extends at its one axial end portion through the above-described through-hole143, that is, through the cylindrical portion142bof the first support wall142, into the axial end portion of the rotor support shaft156on the side of the first support wall142. The differential mechanism input shaft14is aligned with the axis of the first casing12a, which is radially inward of the rotor support shaft156and the sung ear shaft162, such that the input shaft14is rotatable relative to those rotor support shaft156and the sun gear shaft162. The differential mechanism input shaft14is integrally fixed at its one axial end to the first carrier CA1, so that the output of the engine8(not shown inFIG. 11) is transmitted to the first carrier CA1through the differential mechanism input shaft14.

An annular plate164is fixed to the inner circumferential surface of an axial end portion of the first ring gear R1of the first planetary gear set25, which axial end portion is located on the side of the second unit70, such that the annular plate164is not movable relative to the first ring gear R1in the axial and circumferential direction. This annular plate164is perpendicular to the axis of the differential mechanism input shaft14, and has a central bore. An output shaft166of the first planetary gear set24(that is, the output shaft of the power distributing mechanism16) has a shaft portion166ain the form of a sleeve extending toward the second unit70, and a flange portion166bwhich extends radially from the axial end of the shaft portion166aon the side of the first planetary gear set24. This flange portion166bis welded to the annular plate164, so that the output shaft166and the annular plate164are rotated as a unit. It is noted that the output shaft166functions as the power transmitting member18shown inFIG. 1. It is also noted that the switching clutch C0is disposed between the first support wall142and the first planetary gear set24, while the switching brake B0is disposed radially outwardly of the first planetary gear set24.

The second electric motor M2includes the above-indicated stator84, a rotor168, and the above-indicated rotor support shaft90rotating with the rotor168. The second support wall76disposed on one side of the second electric motor M2on the open end side of the second casing12b(on the side of the first casing12a) has a radially central through-hole172formed therethrough in its axial direction. The second support wall76has a protruding portion76awhich is located radially inwardly of a stator coil85aof the stator85and which axially extends towards the rotor168A bearing174is fitted in the inner circumferential surface of the protruding portion76a.

The second support wall76and the second casing12balso cooperate to constitute a faucet joint. Namely, the outer circumferential surface of the second support wall76is held in abutting contact with a second abutting surface176of the inner circumferential surface of the second casing12b. The second abutting surface176is nearer to the axial open end of the second casing12b, than the above-indicated first abutting surface82, and is radially outwards of the first abutting surface82. Before the second support wall76is fixed to the second casing12bby bolts74, the second support wall76is slidable at its outer circumferential surface on the second abutting surface176. Further, the side surface of the radially outer end portion of the second support wall76, which is on the side of the second electric motor M2, is held in abutting contact with a second radial surface178of the second casing12b, which second radial surface178radially extends from one axial end of the second abutting surface176on the side of the second electric motor M2. Accordingly, the second support wall76can be accurately positioned in its axial and radial directions, by simply fitting the second support wall76into the second casing12buntil the outer circumferential surface and the side surface of the second support wall76come into abutting contact with the second abutting surface176and the second radial surface178of the second casing12b, respectively.

The rotor support shaft90is supported at its one axial end portion by the second support wall76via the above-indicated bearing174. Further, the rotor support shaft80supports, at its axial end portion on the side of the second support wall76, the transmission input shaft72via a bearing180disposed radially inwardly of the above-indicated bearing174. That is, the axial end portion of the transmission input shaft72on the side of the second support wall76is supported by the second support wall76via the bearing180, rotor support shaft90and baring174, and can be accurately positioned in its radial direction since the second support wall76can be accurately positioned in its radial direction, as described above. Further, the other axial end portion of the transmission input shaft72is supported by the third support wall78accurately positioned in its radial direction, so that this other axial end portion can be accurately positioned in its radial direction, whereby the transmission input shaft72can be accurately positioned in its radial direction.

The transmission input shaft72extends through the above-indicated through-hole172toward the first unit140, and is splined to the output shaft166of the first planetary gear set24, at an axial portion corresponding to the through-hole172. The above-indicated 2-5 oil passage124is formed so as to extend in the radial direction, and is held at its one end in communication with the 2-2 oil passage120and open at its other end in the outer circumferential surface of the transmission input shaft72, at an axial position at which the above-indicated bearing180is located. The lubricating oil fed to the 2-5 oil passage124through the 2-2 oil passage120is further fed to the above-indicated bearing180and to the bearing174located radially outwardly of the bearing180.

The axial end portion of the transmission input shaft72on the side of the first unit140extends into the axial end portion of the differential mechanism input shaft14on the side of the second unit70, which is supported by the transmission input shaft72via a bushing181which is located radially inwardly of the first sun gear S1and which is interposed between the differential mechanism input shaft14and the transmission input shaft72. The transmission input shaft72is supported by the second casing12bvia the bearing180, rotor support shaft90, bearing174and second support wall76, so that the axial end portion of the differential input shaft14on the side of the second unit70is supported by the second casing12bvia the above-indicated members181,72,180,90,174,76. Further, the differential mechanism input shaft14is supported by the rotor support shaft156via the bearing182which is interposed between the differential mechanism input shaft14and the inner circumferential surface of the axial end portion of the rotor support shaft156on the side of the lid plate150. The rotor support shaft156is supported by the first casing12avia the bearing160and the lid plate150, so that the differential mechanism input shaft14is supported by the first casing12avia the above-indicated members182,156,160and150. Thus, the differential mechanism input shaft14is supported by the casing12at its two axial portions, which are spaced from each other in the axial direction by a comparatively large distance, so that the differential mechanism input shaft14can be accurately positioned in its radial direction. As described above, the differential mechanism input shaft14supported at its one axial end portion by the transmission input shaft72can be accurately aligned with the transmission input shaft72in the radial direction.

The differential mechanism input shaft14has third oil passages in the form of a radially central 3-1 oil passage183which is open at its one axial end toward the 2-2 oil passage120, and a 3-2 oil passage184which is held at its axial end in communication with the 3-1 oil passage183and open at its other axial end in the outer circumferential surface of an axial portion of the differential mechanism input shaft14on which the bearing182is fitted. The lubricating oil is fed to the 3-1 oil passage183through the 2-2 oil passage120, and is further fed from the 3-1 oil passage183to the 3-2 oil passage184, to lubricate the bearings182,160. It is noted that the above-indicated bushing181functioning as a lubricant sealing member assures a sufficient amount of supply of the lubricating oil from the 2-2 oil passage120to the 3-1 oil passage183, without an additional sealing member.

FIG. 13is a fragmentary view showing in enlargement the power distributing mechanism16shown inFIG. 11. The first sun gear S1is supported by the differential mechanism input shaft14via a bushing186interposed between the first sun gear S1and the differential mechanism input shaft14. The other axial end portion of the sun gear shaft162formed integrally with the first sun gear S1extends into the rotor support shaft156, as described above, and the rotor support shaft156and the sun gear shaft162are fitted on each other through a spline portion188and rotated as a unit. A bushing190is interposed between the differential mechanism input shaft14and the inner circumferential surface of an axial portion of the sun gear shaft162which corresponds to the spline portion188. The sun gear shaft162is supported at its axial end portion on the side of the spline portion188, by the differential mechanism input shaft14via the bushing190. Accordingly, the sun gear shaft162formed integrally with the first sun gear S1is supported by the differential mechanism input shaft14, at its two axial portions via the bushings186,190. Since the differential mechanism input shaft14is positioned accurately in its radial direction, as described above, the first sun gear S1and the sun gear shaft162can also be accurately positioned in their radial direction.

The switching clutch C0includes: a clutch cylinder192fitted on the above-indicated cylindrical portion142bof the first support wall142; a clutch piston194accommodated in the clutch cylinder192; a plurality of pressure plates196and a plurality of friction discs198which are forced against each other by the clutch piston194. The clutch cylinder192includes: a bottom portion192aextending in parallel with the upright portion142aof the first support wall142: an inner cylindrical portion192bfixed to the radially inner end of the bottom portion192aand fitted on the cylindrical portion142bof the first support wall142; and an outer cylindrical portion192cfixed to the radially outer end of the bottom portion192a. The clutch piston194is accommodated in the clutch cylinder192such that the bottom portion192aof the clutch cylinder192cooperates with the clutch piston194to define therebetween an oil chamber200.

The sun gear shaft162has a flange portion162awhich radially extends toward the inner cylindrical portion192bof the clutch cylinder192and which has an outer circumferential surface located radially outwardly of that of the first sun gear S1. The inner circumferential surface of the inner cylindrical portion192bof the clutch cylinder192which is a member of the switching clutch C0, and the outer circumferential surface of the flange portion162aof the sun gear shaft162are welded together at a weld portion202. Thus, the first sun gear S1formed integrally with the sun gear shaft162, and the clutch cylinder192are integrally formed as a unit. Since the first sun gear S1and the sun gear shaft162are accurately positioned in their radial direction, as described above, the clutch cylinder192is also accurately positioned in the radial direction.

A thrust bearing204is disposed between one of the side surfaces of the flange portion162aof the sun gear shaft162, and the axial end face of the cylindrical portion142bof the first support wall142, which end face is opposed to the above-indicated one side surface of the flange portion162a. A thrust force acting on the first sun gear S1in the axial direction toward the first electric motor M1is received by the first support wall142through the thrust bearing204. Since the sun gear shaft162and the rotor support shaft156are fitted on each other through the spline portion188, the thrust force acting on the first sun gear S1in the axial direction toward the first electric motor M1is not transmitted to the rotor support shaft156. A thrust force acting on the first sun gear S1in the axial direction opposite to the axial direction toward the first electric motor M1is received by the differential mechanism input shaft14integral with the first carrier CA1, through a thrust bearing206disposed between the opposite side surfaces of the first sun gear S1and the first carrier CA1.

The above-indicated plurality of pressure plates196are splined to the inner circumferential surface of the outer cylindrical portion192c. A retainer ring208is also splined to the inner circumferential surface of the outer cylindrical portion192c, at an axial position which is nearer to the open end of the clutch cylinder192, than the pressure plate196nearest to the open end. On the other hand, the plurality of friction discs198each interposed between the adjacent pressure plates196are splined to the outer circumferential surface of a clutch hub210which is fixed to the radially outer end portion of the first carrier CA1and which extends toward the clutch piston194in the axial direction of the clutch piston194. A spring retainer plate212is fixed to the outer circumferential surface of the inner cylindrical portion192bof the clutch cylinder192, such that the spring retainer plate212is located radially inwardly of the clutch hub210and in the axial open end portion of the clutch cylinder192, and extends in the radial direction of the clutch cylinder192, and such that the spring retainer plate212is not movable in the axial direction toward the first planetary gear set24. A return spring214is disposed between the spring retainer plate212and the clutch piston194.

Oil passages are formed through the first support wall142, for supplying the working oil to the oil chamber200of the switching clutch C0constructed as described above. Described in detail, the upright portion142aof the first support wall142has a first radial oil passage216, and the cylindrical portion142bof the first support wall142has an axial oil passage218which is held at its one axial end in communication with the first radial oil passage216, and a second radial oil hole220which is held at its one axial end in communication with the axial oil passage218and open at its other axial end in the outer circumferential surface of the cylindrical portion142b. Further, the inner cylindrical portion192bof the clutch cylinder192has a third radial oil hole222for communication between the above-indicated second radial oil hole220and the oil chamber200. In the present embodiment, the sun gear shaft162extends through the cylindrical portion142bof the first support wall142and projects from the cylindrical portion142bin the axial direction away from the first sun gear S1, and the sun gear shaft162and the rotor support shaft156are located radially inwardly of the bearing158disposed adjacent to the cylindrical portion142b, and are fitted on each other through the spline188. In this arrangement, the wall thickness (thickness in the radial direction) of the cylindrical portion142bcan be made larger than in an arrangement wherein the rotor support shaft156extends into the cylindrical portion142band fitted on the sun gear shaft162, within the cylindrical portion142b. Accordingly, the above-indicated axial oil passage218and the second radial oil hole220can be formed comparatively easily through the cylindrical portion142b.

A brake hub224consists of an inner cylindrical portion224afitted on the outer circumferential surface of the outer cylindrical portion192cof the clutch cylinder192; a connecting portion224bwhich is connected at its radially inner end portion to one axial end portion of the inner cylindrical portion224aremote from the first support wall142and which extends radially outwardly of the inner cylindrical portion224a; and an outer cylindrical portion224cwhich is connected at its one axial end portion to the radially outer end portion of the connecting portion224band which extends from the connecting portion224bin the axial direction away from the inner cylindrical portion224a. The inner cylindrical portion224ais welded to the outer cylindrical portion192cof the clutch cylinder192, so that the brake hub224is fixedly positioned and is rotated with the clutch cylinder192.

The switching brake B0has: the above-indicated brake hub224; a brake cylinder226fitted in the first casing12a; a brake piston228accommodated in the brake cylinder226; and a plurality of pressure plates230and a plurality of friction discs232which are forced against each other by the brake piston228.

The upright portion142aof the first support wall142includes a radially outer end portion having a relatively large wall thickness and extending toward the switching brake B0. The inner circumferential surface of the first casing12ahas a spline234between the end face of the upright portion142aof the first support wall142which faces the switching brake B0, and the end face of the brake cylinder26which faces the first support wall142. The plurality of pressure plates230are splined to the inner circumferential surface of the first casing12athrough the spline234. A spacer member236in the form of a sleeve is interposed between the first support wall142and one of the plurality of pressure plates230which is nearest to the first support wall142. On the other hand, the plurality of friction discs232are splined to the outer circumferential surface of the outer cylindrical portion224cof the brake hub224.

The above-described brake cylinder226is prevented from moving in one of the opposite axial directions, in abutting contact with the end faces of the teeth of the spline234, and in the other axial direction by a retainer spring238fixed to the first casing12a. A spring retainer plate240is fixed to the axial open end portion of the brake cylinder226, so as to extend in the radial direction, such that the spring retainer plate240is not axially movable toward the first support wall142. The return spring242is interposed between the spring retainer plate140and the brake piston228.

The 2-5 oil passage125formed through the transmission input shaft72is open at its one end in the outer circumferential surface of the transmission input shaft72, at an axial position corresponding to the axial end portion of the differential mechanism input shaft14(at which the input shaft14is connected to the first carrier CA1). The 2-5 oil passage125is held at its other end in communication with the 2-2 oil passage120. Further, a third oil passage in the form of a 3-3 oil passage244is formed so as to extend between the axial end portion of the differential mechanism input shaft14which is connected to the first carrier CA1, and the first carrier CA1. The 3-3 oil passage244is open at its one end in the inner circumferential surface of the differential mechanism input shaft14, such that this one end is opposed to the 2-5 oil passage125. Further, fourth oil passages in the form of a 4-1 oil passage248, a 4-2 oil passage250and a 4-3 oil passage252are formed through a pinion shaft246fitted in the first carrier CA1. The 4-1 oil passage248is formed so as to extend in the radial direction of the pinion shaft246, and is held at its one end in communication with the 3-3 oil passage244. The 4-2 oil passage250is formed in the axial direction of the pinion shaft246, and is held at its one end in communication with the 4-1 oil passage248. The 4-3 oil passage252is held at its one end in communication with the 4-2 oil passage250, and is open at its other end between two needle bearings254,256which are interposed between the pinion shaft246and the first planetary gear P1. The lubricating oil is supplied to the differential mechanism in the form of the first planetary gear set24through the 2-5 oil passage125, 4-1 oil passage248, 4-2 oil passage250and 4-3 oil passage252. The differential mechanism input shaft14further has third oil passages in the form of a 3-4 oil passage258and a 3-5 oil passage260formed therethrough in the axial direction, so that the lubricating oil is supplied to the first planetary gear set24also through those 3-4 oil passage258and 3-5 oil passage260. The lubricating oil which has lubricated the first planetary gear set24is used to lubricate also the pressure plates230and friction discs232of the switching brake B0which is located radially outwardly of the first planetary gear set24.

In the present embodiment described above, the lubricating oil is supplied from the third support wall78between the second electric motor M2and the automatic transmission20, to the first electric motor M1, power distributing mechanism16and second electric motor M2which are disposed on one of opposite axial sides of the third support wall78, and to the automatic transmission20which is disposed on the other axial side of the third support wall78. Accordingly, the arrangement of the lubricating oil passages can be made simpler in the present embodiment, than in the case where two sets of lubricating oil passages are provided for the respective two groups of devices disposed on the respective opposite axial sides of the third support wall78.

While the preferred embodiment of this invention has been described above by reference to the accompanying drawings, for illustrative purpose only, it is to be understood that the present invention may be embodied with various changes and modifications, as described below.

In the illustrated embodiment, the power distributing mechanism16is placed selectively in one of its differential state and non-differential state, so that the drive system10is switchable between the continuously-variable shifting state in which the drive system is operable as an electrically controlled continuously-variable transmission, and the step-variable shifting state in which the drive system is operable as a step-variable transmission. However, the switching between the continuously-variable shifting state and the step-variable shifting state is one form of the switching between the differential state and the non-differential state of the power distributing mechanism16. For instance, the power distributing mechanism16may be operated as a step-variable transmission the speed ratio of which is variable in steps, even while the power distributing mechanism16is placed in the differential state. In other words, the differential state and the non-differential state of the drive system10(power distributing mechanism16) do not necessarily correspond to the continuously-variable shifting state and the step-variable shifting state, respectively, and the drive system10need not be switchable between the continuously-variable shifting state and the step-variable shifting state.

In the power distributing mechanism16in the illustrated embodiment, the first carrier CA1is fixed to the engine8, and the first sun gear S1is fixed to the first electric motor M2while the first ring gear R1is fixed to the power transmitting member18. However, this arrangement is not essential. The engine8, first electric motor M1and power transmitting member18may be fixed to any other elements selected from the three elements CA1, S1and R1of the first planetary gear set24.

While the engine8is directly fixed to the differential mechanism input shaft14in the illustrated embodiment, the engine8may be operatively connected to the input shaft14through any suitable member such as gears and a belt, and need not be disposed coaxially with the input shaft14.

Although the power distributing mechanism16in the illustrated embodiment is provided with the switching clutch C0and the switching brake B0, the power distributing mechanism16need not be provided with both of the switching clutch C0and brake B0. While the switching clutch C0is provided to selectively connect the first sun gear S1and the first carrier CA1to each other, the switching clutch C0may be provided to selectively connect the first sun gear S1and the first ring gear R1to each other, or selectively connect the first carrier CA1and the first ring gear R1. Namely, the switching clutch C0may be arranged to connect any two elements of the three elements of the first planetary gear set24.

While the switching clutch C0is engaged to establish the neutral position N in the drive system10in the illustrated embodiment, the switching clutch C0need not be engaged to establish the neutral position.

The frictional coupling devices used as the switching clutch C0, switching brake B0, etc. in the illustrated embodiment may be replaced by a coupling device of a magnetic-power type, an electromagnetic type or a mechanical type, such as a powder clutch (magnetic powder clutch), an electromagnetic clutch and a meshing type dog clutch.

The drive system10according to the illustrated embodiment is a drive system for a hybrid vehicle in which the drive wheels38can be driven by not only the engine8but also the first electric motor or the second electric motor M2. However, the principle of the present invention is applicable to a vehicular drive system in which the power distributing mechanism16is not operable in a hybrid control mode, and functions only as a continuously variable transmission so-called an “electric CVT”.

In the illustrated embodiment, the first support wall142is formed integrally with the casing12, while the second support wall76and the third support wall78are formed separately from the casing12and fixed to the casing12through the bolts74,84. However, it is possible that the first support wall142is formed separately from the casing12and fixed to the casing12through bolts or other fastening means, while the second support wall76and the third support wall78are formed integrally with the casing12.

While the power distributing mechanism16is constituted by one planetary gear set in the illustrated embodiment, the power distributing mechanism16may be constituted by two or more planetary gear sets. In this case, the power distributing mechanism16functions as a transmission having three or more gear positions in the fixed-speed-ratio shifting state. Further, the planetary gear set may be replaced by a differential gear device including a pinion rotated by the engine, and a pair of bevel gears meshing with the pinion and operatively connected to the first electric motor M1and the power transmitting member18.

The power transmitting device in the form of the automatic transmission20in the illustrated embodiment includes the three planetary gear sets26,28and30. However, the construction of the automatic transmission is not limited to the details of the illustrated embodiment, in the number of the planetary gear sets, the number of the gear positions, and the selective connections of the clutches C and brakes B to the elements of the planetary gear sets. Further, the automatic transmission20may be replaced by a speed reducing mechanism including one planetary gear set, as disclosed in JP-2004-116735A, or by a power transmitting device not arranged to change its speed ratio.

It is to be understood that the embodiment described above is given for illustrating the present invention and that the invention may be embodied with various other changes and modifications which may occur to those skilled in the art.