JP2008213686A - Control device for vehicle driving device - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To reduce the shock of shift when an engine is started in the control device of a vehicle driving device equipped with an electric type differential part for controlling the differential state of the rotation speed of an input shaft and the rotation speed of an output shaft and a transmission part configuring a portion of a power transmission path from the electric type differential part to a driving wheel by controlling the operating state of a first electric motor connected to the engine and the rotary element of a differential mechanism so that power can be transmitted. <P>SOLUTION: In the control device of a vehicle driving device, the engagement device of an automatic transmission part 20 is controlled so that the total torque T<SB>TOTAL</SB>of the engagement device in transmission can be set so as to be less than the torque of a transmission member 18 in the complete explosion of an engine 8 until the complete explosion of the engine 8 in gear shift of the automatic transmission part 20 when starting the engine 8. Thus, even when torque fluctuation due to the complete explosion of the engine 8 is generated, the total torque T<SB>TOTAL</SB>or more can be prevented from being transmitted to an output shaft 22 side of the automatic transmission part 20. Therefore, it is possible to suppress the transmission of torque fluctuation to an output shaft 34 side, and to reduce shock due to the start (complete explosion) of the engine 8. <P>COPYRIGHT: (C)2008,JPO&INPIT

Description

  The present invention controls the difference between the rotational speed of the input shaft connected to the engine and the rotational speed of the output shaft by controlling the operating state of the engine and an electric motor connected to the rotating element of the differential mechanism so that power can be transmitted. More particularly, the present invention relates to a control device for a vehicle drive device, comprising: an electric differential unit in which a dynamic state is controlled; and a transmission unit that forms part of a power transmission path from the electric differential unit to a drive wheel Further, the present invention relates to a reduction in shift shock accompanying engine start.

  By controlling the operating state of the engine and the motor connected to the rotating element of the differential mechanism so that power can be transmitted, the differential state between the rotational speed of the input shaft connected to the engine and the rotational speed of the output shaft is controlled. There is known a control device for a vehicle drive device that includes an electric differential portion and a transmission portion that constitutes a part of a power transmission path from the electric differential portion to a drive wheel. For example, this is the control device for a vehicle drive device described in Patent Document 1. In such a vehicle drive device, in general, in an operation region where engine efficiency is poor such as a low vehicle speed region or a low output region, the fuel consumption of the vehicle can be reduced by stopping the engine and running the motor. On the other hand, when an engine efficiency region such as a medium / high vehicle speed region or a medium / high output region is reached, the engine is started and the vehicle travels mainly. Even in the motor travel region, when the battery capacity decreases, the engine is started for power generation. As described above, in the vehicle drive device disclosed in Patent Document 1 or the like, the engine is started and stopped repeatedly to reduce the fuel consumption of the vehicle.

JP 2005-264762 A JP 2006-138426 A JP 2003-74688 A

  By the way, in Patent Document 1 described above, as a starting method of the engine, the first electric motor and the second electric motor are simultaneously rotated in the same rotational direction, so that the rotational speed of the engine is quickly increased to the ignition possible rotational speed to A technique for improving startability is disclosed. Here, for example, if the engine is started during a shift such as a downshift, a torque fluctuation occurs due to a complete explosion of the engine, and this torque fluctuation is transmitted to the output side of the transmission unit, which may cause a shock. It was.

  The present invention has been made against the background of the above circumstances. The object of the present invention is to control the operating state of an engine and a first electric motor connected to a rotating element of a differential mechanism so that power can be transmitted. An electric differential unit that controls a differential state between the rotational speed of the input shaft connected to the engine and the rotational speed of the output shaft, and a power transmission path from the electrical differential unit to the drive wheels. An object of the present invention is to provide a control device for a vehicle drive device that reduces a shift shock that accompanies an engine start.

  To achieve the above object, the gist of the invention according to claim 1 is that: (a) the operating state of the engine and the first motor connected to the rotating element of the differential mechanism so as to be able to transmit power is controlled; An electric differential unit that controls a differential state between the rotational speed of the input shaft connected to the engine and the rotational speed of the output shaft, and a power transmission path from the electrical differential unit to the drive wheels. In the control device for a vehicle drive device comprising a part of the speed change portion, (b) in the speed change of the speed change portion that accompanies the start of the engine, until the engine complete explosion, An underlap control means is provided for controlling the engaging device of the transmission unit so that the sum of the torque capacities of the engaging device is less than the torque of the output shaft at the time of complete explosion of the engine.

  According to a second aspect of the present invention, there is provided a control device for a vehicle drive device according to the first aspect, wherein a second electric motor is connected to an output shaft of the differential mechanism, and the underlap control is performed. During the speed change by the means, there is provided a synchronization control means for synchronously controlling the rotation speed of the output shaft of the differential mechanism by the second electric motor.

  The gist of the invention according to claim 3 is that, in the control device for a vehicle drive device according to claim 1 or 2, the speed change of the speed change portion is performed by releasing and engaging the release side hydraulic friction engagement device. A clutch-to-clutch shift that is performed simultaneously with the engagement of the side hydraulic friction engagement device, and the underlap control means includes a release pressure of the release side hydraulic friction engagement device and an engagement side hydraulic type. The sum of the torque capacities is maintained at a predetermined underlap value by providing a preset standby pressure at least one of the engagement pressures of the friction engagement device.

  According to a fourth aspect of the present invention, there is provided the control device for a vehicle drive device according to the second or third aspect, wherein the electric differential unit is an operating state of the first and / or second electric motor. Is controlled to operate as a continuously variable transmission mechanism.

  According to a fifth aspect of the present invention, there is provided the vehicle drive device control device according to any one of the first to fourth aspects, wherein the transmission unit is a stepped automatic transmission. And

  According to the control device for a vehicle drive device of the first aspect of the present invention, in the shift of the transmission unit that accompanies the start of the engine, the torque capacity of each engaging device related to the shift until the complete explosion of the engine. By controlling the engagement operation of the transmission unit so that the sum is less than the torque of the output shaft at the time of complete explosion of the engine, even if a torque fluctuation occurs due to the complete explosion of the engine, the output side ( Since torque exceeding the torque capacity of the engagement device is not transmitted to the drive wheel side), transmission of torque fluctuation to the drive wheel side is suppressed, and shock due to engine start (complete explosion) is reduced.

  According to the control device for a vehicle drive device of the invention of claim 2, during the speed change by the underlap control means, the rotation speed of the output shaft of the differential mechanism is synchronously controlled by the second electric motor. Thus, the rotational speed of the output shaft of the differential mechanism is quickly synchronized with the target rotational speed at the end of shifting. Thereby, the delay of the rotation speed synchronization by the underlap control means is suppressed, and the delay of the shift is suppressed.

  According to the control device for a vehicle drive device of the invention of claim 3, at least one of the release pressure of the release side hydraulic friction engagement device and the engagement pressure of the engagement side hydraulic friction engagement device. By maintaining a predetermined standby pressure at a predetermined underlap value by providing a pre-set standby pressure on the output side of the transmission unit (driving wheels) Since no torque greater than the underlap value is transmitted to the side), transmission of torque fluctuation to the drive wheel side is suppressed, and shock due to engine start (complete explosion) is reduced. Further, the sum of the torque capacities of the release-side and engagement-side hydraulic friction engagement devices can be easily maintained at a predetermined underlap value by controlling the standby pressure.

  According to the control apparatus for a vehicle drive device of the invention according to claim 4, the continuously variable transmission is configured by the electric differential portion and the transmission portion, and the drive torque can be smoothly changed. The electric differential unit can be operated as an electric continuously variable transmission by continuously changing the gear ratio, and can also be operated as a stepped transmission by changing the gear ratio stepwise. The overall shift of the vehicle drive device can be changed stepwise to obtain drive torque quickly.

  Further, according to the control device for a vehicle drive device of the invention according to claim 5, for example, a continuously variable transmission between an electric differential portion that functions as an electric continuously variable transmission and a stepped automatic transmission. In the state in which the machine is configured and the drive torque can be changed smoothly and the transmission ratio of the electric differential unit is controlled to be constant, the electric differential unit and the stepped automatic transmission are Thus, a state equivalent to the stepped transmission is configured, and the overall shift of the vehicle drive device is changed stepwise, so that the drive torque can be obtained quickly.

  Preferably, the rotating element of the differential mechanism includes a first element connected to the input shaft and the engine, a second element connected to the first electric motor, and a third element connected to the output shaft. A planetary gear set having three rotating elements and the first element is a carrier of the planetary gear set, the second element is a sun gear of the planetary gear set, and the third element is the planetary gear set. It is a ring gear of a gear device. In this way, the axial direction dimension of the differential mechanism is reduced. Further, the differential mechanism is simply constituted by one planetary gear device.

  Preferably, the planetary gear device is a single pinion type planetary gear device. In this way, the axial direction dimension of the differential mechanism is reduced. Further, the differential mechanism is simply constituted by one single pinion type planetary gear device.

  Preferably, the overall transmission ratio of the vehicle drive device is formed based on the transmission ratio (gear ratio) of the transmission unit and the transmission ratio of the electric differential unit. In this way, a wide driving force can be obtained by utilizing the gear ratio of the transmission unit.

  Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings.

  FIG. 1 is a skeleton diagram illustrating a speed change mechanism 10 constituting a part of a drive device for a hybrid vehicle to which the present invention is applied. In FIG. 1, a transmission mechanism 10 includes an input shaft 14 as an input rotation member disposed on a common axis in a transmission case 12 (hereinafter referred to as case 12) as a non-rotation member attached to a vehicle body, The differential unit 11 as a continuously variable transmission unit directly connected to the input shaft 14 or indirectly through a pulsation absorbing damper (vibration damping device) (not shown), the differential unit 11 and the drive wheel 34 (FIG. 7). An automatic transmission unit 20 as a power transmission unit connected in series via a transmission member (transmission shaft) 18 in a power transmission path between the output transmission member and the output rotation member connected to the automatic transmission unit 20 As an output shaft 22 in series. The speed change mechanism 10 is preferably used in, for example, an FR (front engine / rear drive) type vehicle vertically installed in a vehicle, and directly to the input shaft 14 or directly via a pulsation absorbing damper (not shown). As a driving power source for traveling, for example, an engine 8 that is an internal combustion engine such as a gasoline engine or a diesel engine is provided between a pair of drive wheels 34 and power from the engine 8 is part of a power transmission path. Is transmitted to the pair of drive wheels 34 through the differential gear device (final reduction gear) 32 (see FIG. 7) and the pair of axles. The transmission mechanism 10 of this embodiment corresponds to the vehicle drive device of the present invention.

  Thus, in the transmission mechanism 10 of the present embodiment, the engine 8 and the differential unit 11 are directly connected. This direct connection means that the connection is made without using a hydraulic power transmission device such as a torque converter or a fluid coupling. For example, the connection via the pulsation absorbing damper is included in this direct connection. Since the speed change mechanism 10 is configured symmetrically with respect to its axis, the lower side is omitted in the skeleton diagram of FIG. The same applies to each of the following embodiments.

  The differential unit 11 is a mechanical mechanism that mechanically distributes the output of the engine 8 input to the first electric motor M1 and the input shaft 14, and distributes the output of the engine 8 to the first electric motor M1 and the transmission member 18. A power distribution mechanism 16 serving as a differential mechanism, and a second electric motor M2 that is operatively connected to rotate integrally with the transmission member 18. The first electric motor M1 and the second electric motor M2 of the present embodiment are so-called motor generators that also have a power generation function, but the first electric motor M1 has at least a generator (power generation) function for generating a reaction force, and the second electric motor. M2 has at least a motor (electric motor) function for outputting driving force as a driving force source for traveling. The differential portion 11 of this embodiment corresponds to the electrical differential portion of the present invention.

  The power distribution mechanism 16 is mainly configured by a single pinion type first planetary gear device 24 having a predetermined gear ratio ρ1 of about “0.418”, for example. The first planetary gear unit 24 includes a first sun gear S1, a first planetary gear P1, a first carrier CA1 that supports the first planetary gear P1 so as to rotate and revolve, and a first sun gear via the first planetary gear P1. A first ring gear R1 meshing with S1 is provided as a rotating element (element). When the number of teeth of the first sun gear S1 is ZS1 and the number of teeth of the first ring gear R1 is ZR1, the gear ratio ρ1 is ZS1 / ZR1. Note that the power distribution mechanism 16 of this embodiment corresponds to the differential mechanism of the present invention.

In the power distribution mechanism 16, the first carrier CA1 is connected to the input shaft 14, that is, the engine 8, the first sun gear S1 is connected to the first electric motor M1, and the first ring gear R1 is connected to the transmission member 18. In the power distribution mechanism 16 configured as described above, the first sun gear S1, the first carrier CA1, and the first ring gear R1, which are the three elements of the first planetary gear device 24, can be rotated relative to each other, so that a differential action is achieved. Therefore, the output of the engine 8 is distributed to the first electric motor M1 and the transmission member 18, and a part of the distributed output of the engine 8 is used. Since the electric energy generated from the first electric motor M1 is stored or the second electric motor M2 is rotationally driven, the differential unit 11 (power distribution mechanism 16) is caused to function as an electrical differential device, for example, a difference. The moving portion 11 is in a so-called continuously variable transmission state (electric CVT state), and the rotation of the transmission member 18 is continuously changed regardless of the predetermined rotation of the engine 8. That is, the differential unit 11 is an electrically stepless variable gear whose ratio γ0 (the rotational speed N _IN of the input shaft 14 / the rotational speed N ₁₈ of the transmission member ₁₈ ) is continuously changed from the minimum value γ0min to the maximum value γ0max. It functions as a transmission. As described above, the operating state of the first electric motor M1, the second electric motor M2, and the engine 8 connected to the power distribution mechanism 16 (differential unit 11) so as to be able to transmit power is controlled, whereby the rotation of the input shaft 14 is controlled. The differential state between the number and the rotational speed of the transmission member 18 is controlled. Note that the transmission member 18 of this embodiment corresponds to the output shaft of the differential mechanism of the present invention.

  The automatic transmission unit 20 constitutes a part of a power transmission path from the differential unit 11 to the drive wheel 34, and includes a single pinion type second planetary gear unit 26, a single pinion type third planetary gear unit 28, and a single The planetary gear type multi-stage transmission includes a pinion type fourth planetary gear device 30 and functions as a stepped automatic transmission. The second planetary gear unit 26 includes a second sun gear S2 via a second sun gear S2, a second planetary gear P2, a second carrier CA2 that supports the second planetary gear P2 so as to rotate and revolve, and a second planetary gear P2. The second ring gear R2 that meshes with the second gear R2 and has a predetermined gear ratio ρ2 of about “0.562”, for example. The third planetary gear device 28 includes a third sun gear S3 via a third sun gear S3, a third planetary gear P3, a third carrier CA3 that supports the third planetary gear P3 so as to rotate and revolve, and a third planetary gear P3. A third ring gear R3 that meshes with the gear, and has a predetermined gear ratio ρ3 of, for example, about “0.425”. The fourth planetary gear unit 30 includes a fourth sun gear S4, a fourth planetary gear P4, a fourth carrier gear CA4 that supports the fourth planetary gear P4 so as to rotate and revolve, and a fourth sun gear S4 via the fourth planetary gear P4. And has a predetermined gear ratio ρ4 of about “0.421”, for example. The number of teeth of the second sun gear S2 is ZS2, the number of teeth of the second ring gear R2 is ZR2, the number of teeth of the third sun gear S3 is ZS3, the number of teeth of the third ring gear R3 is ZR3, the number of teeth of the fourth sun gear S4 is ZS4, When the number of teeth of the fourth ring gear R4 is ZR4, the gear ratio ρ2 is ZS2 / ZR2, the gear ratio ρ3 is ZS3 / ZR3, and the gear ratio ρ4 is ZS4 / ZR4. Note that the automatic transmission unit 20 of this embodiment corresponds to the transmission unit of the present invention.

  In the automatic transmission unit 20, the second sun gear S2 and the third sun gear S3 are integrally connected and selectively connected to the transmission member 18 via the second clutch C2, and the case 12 via the first brake B1. The second carrier CA2 is selectively connected to the case 12 via the second brake B2, the fourth ring gear R4 is selectively connected to the case 12 via the third brake B3, The two ring gear R2, the third carrier CA3, and the fourth carrier CA4 are integrally connected to the output shaft 22, and the third ring gear R3 and the fourth sun gear S4 are integrally connected to connect the first clutch C1. And selectively connected to the transmission member 18.

  In this way, the automatic transmission unit 20 and the differential unit 11 (transmission member 18) are selectively connected via the first clutch C1 or the second clutch C2 used to establish the gear position of the automatic transmission unit 20. It is connected. In other words, the first clutch C1 and the second clutch C2 have a power transmission path between the transmission member 18 and the automatic transmission unit 20, that is, a power transmission path from the differential unit 11 (transmission member 18) to the drive wheels 34. It functions as an engagement device that selectively switches between a power transmission enabling state that enables power transmission on the power transmission path and a power transmission cutoff state that interrupts power transmission on the power transmission path. That is, when at least one of the first clutch C1 and the second clutch C2 is engaged, the power transmission path is brought into a power transmission enabled state, or the first clutch C1 and the second clutch C2 are released. Thus, the power transmission path is brought into a power transmission cutoff state.

Further, the automatic transmission unit 20 performs clutch-to-clutch shift by releasing the disengagement side engagement device and engaging the engagement side engagement device, and selectively establishes each gear stage (shift stage). As a result, a gear ratio γ (= rotational speed N ₁₈ of the transmission member ₁₈ / rotational speed N _OUT of the output shaft 22) that changes approximately in a ratio is obtained for each gear. For example, as shown in the engagement operation table of FIG. 2, the first speed gear stage in which the gear ratio γ1 is the maximum value, for example, “3.357” is established by the engagement of the first clutch C1 and the third brake B3. Thus, the engagement of the first clutch C1 and the second brake B2 establishes the second speed gear stage in which the speed ratio γ2 is smaller than the first speed gear stage, for example, about “2.180”. The engagement of the clutch C1 and the first brake B1 establishes the third speed gear stage in which the speed ratio γ3 is smaller than the second speed gear stage, for example, about “1.424”. Engagement of the clutch C2 establishes the fourth speed gear stage in which the speed ratio γ4 is smaller than the third speed gear stage, for example, about “1.000”. In addition, when the second clutch C2 and the third brake B3 are engaged, the reverse gear stage (reverse speed change) in which the speed ratio γR is a value between the first speed gear stage and the second speed gear stage, for example, about “3.209”. Stage) is established. Further, the neutral state “N” is established by releasing the first clutch C1, the second clutch C2, the first brake B1, the second brake B2, and the third brake B3.

  The first clutch C1, the second clutch C2, the first brake B1, the second brake B2, and the third brake B3 (hereinafter referred to as the clutch C and the brake B unless otherwise specified) are conventional automatic transmissions for vehicles. A hydraulic friction engagement device as an engagement element often used in a machine, and a wet multi-plate type in which a plurality of friction plates stacked on each other are pressed by a hydraulic actuator, or an outer peripheral surface of a rotating drum One end of one or two bands wound around is composed of a band brake or the like that is tightened by a hydraulic actuator, and is for selectively connecting the members on both sides of the band brake.

  In the transmission mechanism 10 configured as described above, the differential unit 11 that functions as a continuously variable transmission and the automatic transmission unit 20 constitute a continuously variable transmission as a whole. Further, by controlling the gear ratio of the differential unit 11 to be constant, the differential unit 11 and the automatic transmission unit 20 can configure a state equivalent to a stepped transmission.

Specifically, the differential unit 11 functions as a continuously variable transmission, and the automatic transmission unit 20 in series with the differential unit 11 functions as a stepped transmission, whereby at least one shift of the automatic transmission unit 20 is performed. The rotational speed input to the automatic transmission unit 20 with respect to the stage M (hereinafter referred to as the input rotational speed of the automatic transmission unit 20), that is, the rotational speed of the transmission member 18 (hereinafter referred to as the transmission member rotational speed N ₁₈ ) changes steplessly. As a result, a continuously variable gear ratio width is obtained at the gear stage M. Therefore, the overall speed ratio γT of the transmission mechanism 10 (= the rotational speed N _IN of the input shaft 14 / the rotational speed N _OUT of the output shaft 22) is obtained continuously, and the transmission mechanism 10 constitutes a continuously variable transmission. The overall speed ratio γT of the speed change mechanism 10 is a total speed ratio γT of the speed change mechanism 10 as a whole formed based on the speed ratio γ0 of the differential portion 11 and the speed ratio γ of the automatic speed change portion 20.

For example, first gear or transmission member rotational speed N ₁₈ is continuously variable varying for each gear of the fourth gear and the reverse gear position of the automatic transmission portion 20 indicated in the table of FIG. 2 As a result, each gear stage has a continuously variable transmission ratio width. Therefore, the gear ratio between the gear stages can be continuously changed continuously, and the total gear ratio γT of the transmission mechanism 10 as a whole can be obtained continuously.

  Further, the gear ratio of the differential unit 11 is controlled to be constant, and the clutch C and the brake B are selectively engaged and operated, so that one of the first gear to the fourth gear or the reverse drive By selectively establishing the gear stage (reverse gear stage), a total gear ratio γT of the transmission mechanism 10 that changes approximately in a ratio is obtained for each gear stage. Therefore, a state equivalent to the stepped transmission is configured in the transmission mechanism 10.

  For example, when the gear ratio γ0 of the differential unit 11 is controlled to be fixed to “1”, the first to fourth gear stages of the automatic transmission unit 20 as shown in the engagement operation table of FIG. A total speed ratio γT of the speed change mechanism 10 corresponding to each of the speed gears and the reverse gear is obtained for each gear. Further, if the gear ratio γ0 of the differential unit 11 is controlled to be fixed to a value smaller than “1”, for example, about 0.7 in the fourth speed gear stage of the automatic transmission unit 20, the fourth speed gear stage Is obtained, for example, a total speed ratio γT of about “0.7”.

FIG. 3 is a collinear diagram that can represent, on a straight line, the relative relationship between the rotational speeds of the rotating elements having different connection states for each gear stage in the speed change mechanism 10 including the differential portion 11 and the automatic speed change portion 20. The figure is shown. The collinear diagram of FIG. 3 is a two-dimensional coordinate composed of a horizontal axis indicating the relationship of the gear ratio ρ of each planetary gear unit 24, 26, 28, 30 and a vertical axis indicating the relative rotational speed. X1 represents a rotational speed zero, represents the rotational speed N _E of the engine 8 horizontal line X2 is linked to the rotational speed of "1.0", that is the input shaft 14, horizontal line XG indicates the rotational speed of the power transmitting member 18.

  In addition, three vertical lines Y1, Y2, and Y3 corresponding to the three elements of the power distribution mechanism 16 constituting the differential unit 11 are the first corresponding to the second rotation element (second element) RE2 from the left side. The relative rotation speed of the first ring gear R1 corresponding to the sun gear S1, the first rotation element (first element) RE1 corresponding to the first carrier CA1, and the third rotation element (third element) RE3 is shown. The interval is determined according to the gear ratio ρ1 of the first planetary gear device 24. Further, the five vertical lines Y4, Y5, Y6, Y7, Y8 of the automatic transmission unit 20 correspond to the fourth rotation element (fourth element) RE4 and are connected to each other in order from the left. And the third sun gear S3, the second carrier CA2 corresponding to the fifth rotating element (fifth element) RE5, the fourth ring gear R4 corresponding to the sixth rotating element (sixth element) RE6, and the seventh rotating element ( Seventh element) The second ring gear R2, the third carrier CA3, and the fourth carrier CA4 corresponding to RE7 and connected to each other are connected to the eighth rotation element (eighth element) RE8 and connected to each other. The three-ring gear R3 and the fourth sun gear S4 are respectively represented, and the distance between them is determined according to the gear ratios ρ2, ρ3, and ρ4 of the second, third, and fourth planetary gear devices 26, 28, and 30, respectively. In the relationship between the vertical axes of the nomographic chart, if the distance between the sun gear and the carrier is set to an interval corresponding to “1”, the interval between the carrier and the ring gear is set to an interval corresponding to the gear ratio ρ of the planetary gear device. That is, in the differential unit 11, the interval between the vertical lines Y1 and Y2 is set to an interval corresponding to “1”, and the interval between the vertical lines Y2 and Y3 is set to an interval corresponding to the gear ratio ρ1. Further, in the automatic transmission unit 20, the interval between the sun gear and the carrier is set to an interval corresponding to "1" for each of the second, third, and fourth planetary gear devices 26, 28, and 30, so that the carrier and the ring gear The interval is set to an interval corresponding to ρ.

  If expressed using the collinear diagram of FIG. 3 described above, the speed change mechanism 10 of the present embodiment is configured such that the first rotating element RE1 (the first rotating element RE1) of the first planetary gear device 24 in the power distribution mechanism 16 (the differential unit 11). The carrier CA1) is connected to the input shaft 14, that is, the engine 8, the second rotating element RE2 is connected to the first electric motor M1, and the third rotating element (first ring gear R1) RE3 is connected to the transmission member 18 and the second electric motor M2. Thus, the rotation of the input shaft 14 is transmitted (inputted) to the automatic transmission unit 20 via the transmission member 18. At this time, the relationship between the rotational speed of the first sun gear S1 and the rotational speed of the first ring gear R1 is indicated by an oblique straight line L0 passing through the intersection of Y2 and X2.

For example, in the differential section 11, the first rotation element RE1 to the third rotation element RE3 are in a differential state in which they can rotate relative to each other, and are indicated by the intersections of the straight line L0 and the vertical line Y3. when the rotational speed of the first ring gear R1 is substantially constant is constrained to the vehicle speed V, rotational speed of the first carrier CA1 represented by a point of intersection between the straight line L0 and the vertical line Y2 by controlling the engine rotational speed N _E Is increased or decreased, the rotational speed of the first sun gear S1 indicated by the intersection of the straight line L0 and the vertical line Y1, that is, the rotational speed of the first electric motor M1 is increased or decreased.

The rotation of first sun gear S1 are the same speed as the engine speed N _E by controlling the speed of the first electric motor M1 such speed ratio γ0 of the differential unit 11 is fixed to "1" When the straight line L0 is aligned with the horizontal line X2, the rotational speed, i.e., the power transmitting member 18 of the first ring gear R1 is rotated at the same rotation to the engine speed N _E. Alternatively, the rotation of the first sun gear S1 is made zero by controlling the rotation speed of the first electric motor M1 so that the speed ratio γ0 of the differential unit 11 is fixed to a value smaller than “1”, for example, about 0.7. that the transfer member speed N ₁₈ at a rotation speed higher than the engine speed N _E is rotated.

  Further, in the automatic transmission unit 20, the fourth rotation element RE4 is selectively connected to the transmission member 18 via the second clutch C2, and is also selectively connected to the case 12 via the first brake B1, for the fifth rotation. The element RE5 is selectively connected to the case 12 via the second brake B2, the sixth rotating element RE6 is selectively connected to the case 12 via the third brake B3, and the seventh rotating element RE7 is connected to the output shaft 22. The eighth rotary element RE8 is selectively connected to the transmission member 18 via the first clutch C1.

  In the automatic transmission unit 20, when the rotation of the transmission member 18 (third rotation element RE3) that is an output rotation member in the differential unit 11 is input to the eighth rotation element RE8 by engaging the first clutch C1. As shown in FIG. 3, when the first clutch C1 and the third brake B3 are engaged, the intersection of the vertical line Y8 indicating the rotational speed of the eighth rotational element RE8 and the horizontal line XG and the sixth rotational element A first intersection at an oblique line L1 passing through the intersection of the vertical line Y6 indicating the rotation speed of RE6 and the horizontal line X1 and a vertical line Y7 indicating the rotation speed of the seventh rotation element RE7 connected to the output shaft 22 is the first. The rotational speed of the output shaft 22 at high speed (1st) is shown. Similarly, at an intersection of an oblique straight line L2 determined by engaging the first clutch C1 and the second brake B2 and a vertical line Y7 indicating the rotational speed of the seventh rotating element RE7 connected to the output shaft 22. The rotational speed of the output shaft 22 at the second speed (2nd) is shown, and a seventh rotation coupled to the output shaft 22 and the oblique straight line L3 determined by engaging the first clutch C1 and the first brake B1. The rotation speed of the output shaft 22 of the third speed (3rd) is indicated by the intersection with the vertical line Y7 indicating the rotation speed of the element RE7, and is determined by the engagement of the first clutch C1 and the second clutch C2. The rotation speed of the output shaft 22 at the fourth speed (4th) is shown at the intersection of the straight line L4 and the vertical line Y7 indicating the rotation speed of the seventh rotation element RE7 connected to the output shaft 22.

  FIG. 4 illustrates a signal input to the electronic control device 80 for controlling the speed change mechanism 10 of the present embodiment and a signal output from the electronic control device 80. The electronic control unit 80 includes a so-called microcomputer including a CPU, a ROM, a RAM, an input / output interface, and the like, and performs signal processing according to a program stored in the ROM in advance while using a temporary storage function of the RAM. By performing the above, drive control such as hybrid drive control for the engine 8, the first and second electric motors M1, M2 and the shift control of the automatic transmission unit 20 is executed.

The electronic control device 80 receives signals indicating the engine water temperature TEMP _W , the number of operations at the shift position P _{SH of} the shift lever 52 (see FIG. 6), the “M” position, etc. signal representing the signal indicative of engine rotational speed N _E is the rotational speed of the engine 8, a signal representative of the gear ratio sequence set value, a signal for commanding the M mode (manual shift running mode), a signal representing the operation of the air conditioner, the output A signal representing the vehicle speed V corresponding to the rotational speed of the shaft 22 (hereinafter referred to as the output shaft rotational speed) N _OUT , a signal representing the hydraulic oil temperature T _OIL of the automatic transmission unit 20, a signal representing the side brake operation, and a foot brake operation. Signal, catalyst temperature signal, accelerator pedal operation amount corresponding to the driver's required output, accelerator pedal opening Acc signal, cam angle signal, Signal representing no mode setting, signal representing vehicle longitudinal acceleration G, signal representing auto cruise traveling, signal representing vehicle weight (vehicle weight), signal representing wheel speed of each wheel, rotational speed of first motor M1 N _M1 (hereinafter referred to as the first motor rotation speed N _M1 ), a signal indicating the rotation speed N _M2 of the second motor _M2 (hereinafter referred to as the second motor rotation speed N _M2 ), and the power storage device 56 (see FIG. 7). ) And the like representing the charging capacity (charging state) SOC.

Further, a control signal from the electronic control unit 80 to an engine output control unit 58 (see FIG. 7) for controlling the engine output, for example, a throttle valve opening θ of an electronic throttle valve 62 provided in the intake pipe 60 of the engine 8. _Commands a drive signal to the throttle actuator 64 for operating _TH , a fuel supply amount signal for controlling the fuel supply amount to the intake pipe 60 or the cylinder of the engine 8 by the fuel injection device 66, and an ignition timing of the engine 8 by the ignition device 68 Ignition signal for adjusting, supercharging pressure adjusting signal for adjusting supercharging pressure, electric air conditioner driving signal for operating electric air conditioner, command signal for instructing operation of electric motors M1 and M2, shift for operating shift indicator Position (operation position) display signal, gear ratio display signal for displaying gear ratio, and snow mode A snow mode display signal for indicating, an ABS operation signal for operating an ABS actuator that prevents slipping of the wheel during braking, an M mode display signal for indicating that the M mode is selected, A valve command signal for operating an electromagnetic valve (linear solenoid valve) included in a hydraulic control circuit 70 (see FIGS. 5 and 7) to control the hydraulic actuator of the hydraulic friction engagement device of the automatic transmission unit 20, and the hydraulic pressure signal for applying regulates the line pressure P _L by a regulator valve (pressure regulating valve) provided in the control circuit 70 actuates the electric hydraulic pump is a hydraulic pressure source of the original pressure for the line pressure P _L is pressure adjusted Drive command signal, signal for driving the electric heater, signal to the cruise control computer, etc. are output respectively Is done.

  FIG. 5 is a circuit relating to linear solenoid valves SL1 to SL5 for controlling the operation of the hydraulic actuators (hydraulic cylinders) AC1, AC2, AB1, AB2, and AB3 of the clutches C1 and C2 and the brakes B1 to B3 in the hydraulic control circuit 70. FIG.

  In FIG. 5, each hydraulic actuator AC1, AC2, AB1, AB2, AB3 has an engagement pressure PC1, PC2, PB1 corresponding to a command signal from the electronic control unit 80 by the linear solenoid valves SL1 to SL5. , PB2 and PB3 are respectively regulated and supplied directly. The line oil pressure PL is obtained by using, for example, a relief type pressure regulating valve (regulator valve) as an accelerator opening or a throttle opening with a hydraulic pressure generated from an electric oil pump (not shown) or a mechanical oil pump that is rotationally driven by the engine 30. The pressure is adjusted to a value corresponding to the engine load or the like represented.

  The linear solenoid valves SL1 to SL5 are basically the same in configuration and are excited and de-energized independently by the electronic control unit 80, and the hydraulic pressures of the hydraulic actuators AC1, AC2, AB1, AB2, and AB3 are independently regulated. Thus, the engagement pressures PC1, PC2, PB1, PB2, and PB3 of the clutches C1 to C4 and the brakes B1 and B2 are controlled. In the automatic transmission unit 20, for example, as shown in the engagement operation table of FIG. 2, each gear stage is established by engaging a predetermined engagement device. In the shift control of the automatic transmission unit 20, for example, a so-called clutch-to-clutch shift is performed in which release and engagement of the clutch C and the brake B involved in the shift are controlled simultaneously.

FIG. 6 is a diagram illustrating an example of a shift operation device 50 as a switching device that switches a plurality of types of shift positions _PSH by an artificial operation. The shift operation device 50 includes, for example, a shift lever 52 that is disposed beside the driver's seat and is operated to select a plurality of types of shift positions _PSH .

  The shift lever 52 is in a neutral state, that is, a neutral state in which the power transmission path in the transmission mechanism 10, that is, the automatic transmission unit 20 is interrupted, and a parking position “P (parking) for locking the output shaft 22 of the automatic transmission unit 20. ) ”, Reverse travel position“ R (reverse) ”for reverse travel, neutral position“ N (neutral) ”to establish neutral state where power transmission path in transmission mechanism 10 is cut off, automatic transmission mode established Of the speed change mechanism 10 obtained by the stepless speed change ratio width of the differential unit 11 and each gear stage that is automatically controlled to shift within the range of the first to fourth speed gears of the automatic transmission unit 20. A forward automatic shift travel position “D (drive)” for executing automatic shift control within a change range of the total gear ratio γT that can be shifted, or a manual shift travel mode (manual mode) The by established is provided so as to be manually operated to the forward manual shift drive position for setting a so-called shift range that limits the speed position of the high-speed side of the automatic transmission portion 20 "M (Manual)".

The reverse gear "R" shown in the engagement operation table of FIG 2 in conjunction with the manual operation of the various shift positions P _SH of the shift lever 52, the neutral "N", the shift speed in forward gear "D" etc. For example, the hydraulic control circuit 70 is electrically switched so that is established.

In the shift positions P _SH shown in the “P” to “M” positions, the “P” position and the “N” position are non-traveling positions that are selected when the vehicle is not traveling. As shown in the combined operation table, the first clutch C1 that disables driving of the vehicle in which the power transmission path in the automatic transmission unit 20 in which both the first clutch C1 and the second clutch C2 are released is interrupted. This is a non-driving position for selecting switching to the power transmission cutoff state of the power transmission path by the second clutch C2. The “R” position, the “D” position, and the “M” position are travel positions that are selected when the vehicle travels. For example, as shown in the engagement operation table of FIG. And a power transmission path by the first clutch C1 and / or the second clutch C2 capable of driving a vehicle to which a power transmission path in the automatic transmission 20 is engaged so that at least one of the second clutch C2 is engaged. It is also a drive position for selecting switching to a power transmission enabled state.

  Specifically, when the shift lever 52 is manually operated from the “P” position or the “N” position to the “R” position, the second clutch C2 is engaged and the power transmission path in the automatic transmission unit 20 is changed. When the power transmission is cut off from the power transmission cut-off state and the shift lever 52 is manually operated from the “N” position to the “D” position, at least the first clutch C1 is engaged and the power in the automatic transmission unit 20 is increased. The transmission path is changed from a power transmission cutoff state to a power transmission enabled state. Further, when the shift lever 52 is manually operated from the “R” position to the “P” position or the “N” position, the second clutch C2 is released and the power transmission path in the automatic transmission unit 20 is in a state in which power transmission is possible. From the “D” position to the “N” position, the first clutch C1 and the second clutch C2 are released, and the power transmission in the automatic transmission unit 20 is performed. The path is changed from the power transmission enabled state to the power transmission cut-off state.

FIG. 7 is a functional block diagram for explaining the main part of the control function by the electronic control unit 80. In FIG. 7, the stepped shift control means 82 stores in advance the vehicle speed V and the output torque T _OUT of the automatic transmission unit 20 as variables from the shift diagram shown in FIG. Based on the vehicle state indicated by the actual vehicle speed V and the required output torque T _OUT of the automatic transmission unit 20 from the relationship (shift diagram, shift map) having the upshift line (solid line) and the downshift line (one-dot chain line). Thus, it is determined whether or not the shift of the automatic transmission unit 20 should be executed, that is, the shift stage of the automatic transmission unit 20 to be shifted is determined, and the automatic shift control of the automatic transmission unit 20 is performed so that the determined shift stage is obtained. Execute.

  At this time, the stepped shift control means 82 engages and / or engages the hydraulic friction engagement device involved in the shift of the automatic transmission unit 20 so that the shift stage is achieved, for example, according to the engagement table shown in FIG. A clutch-to-clutch shift is executed by releasing a release command (shift output command, hydraulic pressure command), that is, by releasing the release-side engagement device involved in the shift of the automatic transmission unit 20 and engaging the engagement-side engagement device. Command to output to the hydraulic control circuit 70. In accordance with the command, for example, the hydraulic control circuit 70 releases the disengagement side engagement device and engages the engagement side engagement device so that the shift of the automatic transmission unit 20 is executed. The linear solenoid valve SL is actuated to actuate the hydraulic actuator of the hydraulic friction engagement device involved in the speed change.

The hybrid control means 84 operates the engine 8 in an efficient operating range, while changing the driving force distribution between the engine 8 and the second electric motor M2 and the reaction force generated by the first electric motor M1 to be optimized. Thus, the gear ratio γ0 of the differential unit 11 as an electric continuously variable transmission is controlled. For example, at the traveling vehicle speed V at that time, the target (request) output of the vehicle is calculated from the accelerator opening Acc and the vehicle speed V as the driver's required output amount, and the total required from the target output and the required charging value of the vehicle. Calculate the target output, calculate the target engine output in consideration of transmission loss, auxiliary load, assist torque of the second motor M2, etc. so as to obtain the total target output, and obtain the target engine output. so that the speed N _E and engine torque T _E to control the amount of power generated by the first electric motor M1 controls the engine 8.

For example, the hybrid control means 84 executes the control in consideration of the gear position of the automatic transmission unit 20 for improving power performance and fuel consumption. In such a hybrid control for matching the rotational speed of the power transmitting member 18 determined by the gear position of the engine rotational speed N _E and the vehicle speed V and the automatic transmission portion 20 determined to operate the engine 8 in an operating region at efficient Further, the differential unit 11 is caused to function as an electric continuously variable transmission. That is, the hybrid control means 84, to achieve both the drivability and the fuel consumption when the continuously-variable shifting control in a two-dimensional coordinate composed of the output torque (engine torque) T _E of the engine rotational speed N _E and the engine 8 For example, the target output (total) is set so that the engine 8 is operated along the optimum fuel consumption rate curve (fuel consumption map, relationship) of the engine 8 as shown by the broken line in FIG. target output, required driving force) so that the engine torque T _E and the engine rotational speed N _E for generating an engine output required to satisfy a targeted value of the overall speed ratio γT of the transmission mechanism 10, The gear ratio γ0 of the differential unit 11 is controlled in consideration of the gear position of the automatic transmission unit 20 so that the target value is obtained, and the total gear ratio γT is controlled within the changeable range. I will do it.

  At this time, the hybrid control means 84 supplies the electric energy generated by the first electric motor M1 to the power storage device 56 and the second electric motor M2 through the inverter 54, so that the main part of the power of the engine 8 is mechanically transmitted to the transmission member 18. However, a part of the motive power of the engine 8 is consumed for power generation of the first electric motor M1 and converted into electric energy there, and the electric energy is supplied to the second electric motor M2 through the inverter 54, The second electric motor M2 is driven and transmitted from the second electric motor M2 to the transmission member 18. An electric path from conversion of a part of the power of the engine 8 into electric energy and conversion of the electric energy into mechanical energy by a device related from the generation of the electric energy to consumption by the second electric motor M2 Composed.

Further, the hybrid control means 84 controls the first motor rotation speed N _M1 and / or the second motor rotation speed N _M2 by the electric CVT function of the differential section 11 regardless of whether the vehicle is stopped or traveling. The engine speed _NE can be maintained substantially constant or can be controlled to rotate at an arbitrary speed. In other words, the hybrid control means 84, rotating the first electric motor speed N _M1 and / or the second electric motor rotation speed N _M2 while controlling any rotational speed or to maintain the engine speed N _E substantially constant for any The rotation can be controlled to the speed.

For example, the hybrid control means 84 as can be seen from the diagram of FIG. 3 when raising the engine rotation speed N _E during running of the vehicle, the vehicle speed V the second electric motor rotation speed N which is bound to the (drive wheels 34) The first motor rotation speed N _M1 is increased while maintaining _M2 substantially constant. The hybrid control means 84 when maintaining the engine speed N _E at the nearly fixed level during the shifting of the automatic shifting portion 20, due to the shift of the automatic transmission portion 20 while maintaining the engine speed N _E substantially constant The first motor rotation speed N _M1 is changed in the direction opposite to the change of the second motor rotation speed N _M2 .

  Further, the hybrid control means 84 controls the opening and closing of the electronic throttle valve 62 by the throttle actuator 64 for the throttle control, and controls the fuel injection amount and the injection timing by the fuel injection device 66 for the fuel injection control. For control, a command for controlling the ignition timing by the ignition device 68 such as an igniter is output to the engine output control device 58 alone or in combination, and the output control of the engine 8 is executed so as to generate the necessary engine output. An engine output control means is functionally provided.

For example, the hybrid control means 84 basically drives the throttle actuator 60 based on the accelerator opening Acc from a previously stored relationship (not shown), and increases the throttle valve opening θ _TH as the accelerator opening Acc increases. Throttle control is executed so that Further, the engine output control device 58 controls the opening and closing of the electronic throttle valve 62 by the throttle actuator 64 for throttle control according to the command from the hybrid control means 84, and the fuel injection by the fuel injection device 66 for fuel injection control. The engine torque control is executed by controlling the ignition timing by an ignition device 68 such as an igniter for controlling the ignition timing.

Further, the hybrid control means 84 can drive the motor by the electric CVT function (differential action) of the differential portion 11 regardless of whether the engine 8 is stopped or in an idle state. For example, the hybrid control means 84, typically a relatively low output torque T _OUT region or low engine torque T _E region the engine efficiency is poor compared to the high torque region, or a relatively low vehicle speed range of the vehicle speed V That is, the motor travel is executed in the low load region. Further, the hybrid control means 84 controls the first motor rotation speed N _M1 at a negative rotation speed in order to suppress the drag of the stopped engine 8 and improve fuel consumption during the motor running, for example, the first electric motor M1 is rotated in idle and by a no-load state, to maintain the engine speed N _E at zero or substantially zero as needed by the electric CVT function of the differential portion 11 (differential action).

  Further, even in the engine traveling region, the hybrid control means 84 supplies the second motor M2 with the electric energy from the first electric motor M1 and / or the electric energy from the power storage device 56 by the electric path described above. The so-called torque assist for assisting the power of the engine 8 is possible by driving the two-motor M2 and applying torque to the drive wheels 34.

  Further, the hybrid control means 84 makes the first electric motor M1 in a no-load state and freely rotates, that is, idles, so that the differential unit 11 cannot transmit torque, that is, the power transmission path in the differential unit 11 is interrupted. It is possible to make the state equivalent to the state in which the output from the differential unit 11 is not generated. That is, the hybrid control means 84 can place the differential motor 11 in a neutral state (neutral state) in which the power transmission path is electrically cut off by setting the first electric motor M1 to a no-load state.

  Further, the hybrid control means 84 is transmitted from the kinetic energy of the vehicle, that is, from the drive wheels 34 to the engine 8 side in order to improve fuel efficiency, for example, when coasting with the accelerator off (during coasting) or braking with a foot brake. The second electric motor M2 is rotationally driven by the reverse driving force to act as a generator, and the electric energy, that is, the second electric motor generated current is charged to the power storage device 56 via the inverter 54 as a regeneration control means. This regeneration control is controlled so that the regeneration amount is determined based on the braking force distribution of the braking force by the hydraulic brake for obtaining the braking force according to the charging capacity SOC of the power storage device 56 and the brake pedal operation amount. The

  The underlap control means 86 is a main part of the present invention. In the downshift (downshift) accompanied by the start of the engine 8, the torque capacity possessed by each engagement device related to the shift until the engine 8 is started (complete explosion). The engagement device is controlled so that the sum of the values is less than the torque transmitted to the automatic transmission unit 20 when the engine is completely detonated. . Specifically, when the engine 8 is completely detonated during downshifting, torque fluctuation (rise) occurs due to the complete explosion, and the torque fluctuation (rise) is driven by the output shaft 22 of the automatic transmission unit 20 and further driven. By being transmitted to the wheel 34, there is a possibility that a shock due to engine start (complete explosion) may occur. The underlap control means 86 reduces such a shock.

  The underlap control means 86 is executed based on the determination of the downshift / engine start determination means 88. The downshift / engine start determining means 88 is determined based on, for example, a shift diagram shown in FIG. In the downshift determination, for example, it is determined whether or not a downshift line indicated by a one-dot chain line in the shift diagram shown in FIG. Specifically, as indicated by the broken line arrow, for example, when the vehicle state is changed from a to b by depressing the accelerator pedal, a Dow shift diagram from the second gear to the first gear is shown. It is determined that the vehicle is in a downshift running state where a downshift is executed. Further, the engine start determination is made based on whether or not the vehicle state has moved from the motor travel region surrounded by the thick solid line A in the shift diagram of FIG. 8 to the engine travel region. Specifically, for example, when the accelerator pedal is depressed and the vehicle state is changed from the a state, which is the motor travel region, to the b state, which is the engine travel region, as indicated by the broken line, engine start is requested. Determined. When these determinations affirm the downshift determination and the engine start (request) determination, the underlap control means 86 is executed. These downshift determination and engine start (request) determination are determined simultaneously.

  Next, the control operation of the underlap control means 86 will be described in detail using the time chart shown in FIG. In addition, this time chart corresponds to the case where the vehicle state as indicated by the broken-line arrow in FIG. 8 is changed from the a state to the b state as an example, and thereafter the movement from the a state to the b state is performed. An example will be described. It should be noted that the state a is a motor running and second speed gear running state in the vehicle, and the state b is an engine running and first speed gear running state in the vehicle.

At time T1, when the downshift / engine start determination means 88 requests downshift and start of the engine 8, clutch-to-clutch shift control, which is start of the engine 8 and downshift, is performed simultaneously. In the downshift from the a state to the b state, the downshift crosses the downshift line from the second speed gear stage to the first speed gear stage, so the downshift from the second speed gear stage to the first speed gear stage is performed. A certain clutch-to-clutch shift control is executed. Specifically, as shown in FIG. 2, clutch-to-clutch shift control is performed in which the release of the second brake B2 and the engagement of the third brake B3 are performed simultaneously. Here, at the time T1 to T2, the engagement pressure (release pressure) PB2 of the second brake B2 corresponding to the release-side hydraulic friction engagement device is gradually decreased, and a predetermined standby pressure that is set in advance through experiments or the like. Standby (maintain) at PX (hereinafter, the engagement pressure PB2 of the second brake B2 is described as a release pressure PB2 so that it can be easily distinguished). Further, at the time T1 to T2, the engagement pressure PB3 of the third brake B3 corresponding to the hydraulic friction engagement device on the engagement side is gradually increased, and is waited (maintained) at the standby pressure PY set in advance through experiments or the like. . Here, the standby pressure PX and the standby pressure PY are set so as to wait (maintain) at a time set in advance through experiments or the like by timer control or the like, for example, with reference to the time point T1. Specifically, for example, the rotational speed N _E of the engine 8 is set to be waiting to reach the engine ignition of possible ignitable rotational speed being controlled by the first electric motor M1, always engine 8 Control is performed so that the release pressure PB2 of the second brake B2 and the engagement pressure PB3 of the third brake B3 are kept (maintained) at the standby pressures PX and PY before the complete explosion. Note that the second brake B2 of this embodiment corresponds to a disengagement hydraulic friction engagement device that is an engagement device related to the speed change of the present invention, and the third brake B3 is an engagement related to the engagement of the present invention. It corresponds to an engagement side hydraulic friction engagement device which is a device.

Also, starting of the engine 8, first, at time T1 to time T2, the rotational speed N _E of the engine 8 by the first electric motor M1 is raised until it reaches the engine bootable ignitable rotational speed, preferably by the ignition device 68 By igniting at a proper timing, the engine 8 is completely exploded at time T2. Due to the complete explosion of the engine 8, the torque fluctuation (rise) of the transmission member 18 occurs at time T2. Here, the underlap control means 86 enables the third brake B3 by the torque capacity and the standby pressure PY that allow the second brake B2 by the standby pressure PX by the time T2 when the engine 8 is completely exploded. The total T _TOTAL with the torque capacity is controlled so as to be lower than the torque T ₁₈ of the transmission member 18 that rapidly increases due to the complete explosion of the engine 8. That is, the sum _{T TOTAL} torque capacity of the standby pressure PX and the standby pressure PY of the second brake by suitably controlling at least one of standby pressure B2 and the third brake B3, so below the torque _{T 18} of the transmission member 18 Control is performed to maintain a predetermined underlap value T _UNDER . In other words, the standby pressure PX and the standby pressure PY are controlled so that the total torque capacity T _TOTAL that can be transmitted (allowed) by the automatic transmission unit 20 is less than the torque T ₁₈ of the transmission member 18 (underlap value T _UNDER ).

Thus, the complete explosion of the engine 8 at time T2, the torque _{T 18} is rapidly increased in the transmission member 18 (change), the total torque capacity _{T TOTAL} automatic shifting portion 20 can be transmitted (allowable) as described above because below torque T ₁₈ the transmission member 18, the second brake B2 and the third brake B3 is engaged apparatus according to the shift that slippage occurs, the output shaft 22 and the drive wheels 34 of the automatic transmission portion 20 The transmission of torque fluctuation due to the complete explosion of the engine 8 is suppressed.

As described above, the underlap control means 86 causes the standby pressure PX of the second brake B2 corresponding to the release-side hydraulic friction engagement device and the third brake B3 corresponding to the engagement-side hydraulic friction engagement device. the standby pressure PY, for a given torque value _{T 18} of the power transmitting member 18 varies with complete explosion of the engine 8 is determined for example experimentally, the sum _{T TOTAL} torque capacity of the second brake B2 and third brake B3 Is set to fall below. Here, if the total torque capacity T _TOTAL of the second brake B2 and the third brake B3 is set too small, torque fluctuations are not transmitted to the output shaft 22 of the automatic transmission unit 20, but the second brake B2 and the third brake B3. There is a possibility that an excessive slip occurs in the brake B3 and an excessive blow-up of the engine 8 due to the slip occurs. Therefore, the total torque capacity T _TOTAL of the second brake B2 and the third brake B3 is set to a suitable underlap value T _UNDER that is less than the torque T ₁₈ of the transmission member 18 and suppresses excessive blowing of the engine 8. Is done. This underlap value T _UNDER is set in advance to an experimentally suitable value, for example. When the underlap value T _UNDER is set, the release side hydraulic friction engagement device (engagement) is set so that the total torque capacity T _{TOTAL of} the second brake B2 and the third brake B3 becomes the _underlap value T _UNDER. The standby pressure PX of the second brake B2 corresponding to the combined device) and the standby pressure PY of the hydraulic friction engagement device (engagement device) on the engagement side are controlled. For example, the standby pressure PX of the second brake B2 and the standby pressure PY of the third brake B3 are controlled to predetermined standby pressures experimentally set in advance. Alternatively, when the underlap value T _UNDER is set to be changed based on the accelerator opening Acc, the second brake is set so that the total T _{TOTAL of} the second brake B2 and the third brake B3 becomes the _underlap value T _UNDER. The standby pressure PX of B2 and the standby pressure PY of the third brake B3 may be set so as to be controlled based on the accelerator opening Acc. It is well known that at the time T2 to T5 after the complete explosion of the engine 8, the engagement pressure of the second brake B2 and the engagement pressure of the third brake B3 are suitably controlled to perform a shift. Normal clutch-to-clutch shift control is executed.

Further, the time T1 to time T3, the rotational speed N ₁₈ of the power transmitting member 18, by being controlled to the second electric motor M2, so at shift completion being determined by the gear ratio and the vehicle speed V after shifting of the automatic shifting portion 20 Synchronously controlled with the rotation speed. This synchronization control is executed by the synchronization control means 90 shown in FIG. 7, and the rotation speed N18 of the transmission member ₁₈ is synchronously controlled by the second electric motor M2, thereby reducing the time required for the rotation speed synchronization, The synchronization delay of the rotational speed generated by the underlap control means 86 is reduced. Note that the synchronous control by the second electric motor M2 is not performed during coasting (in coasting) in which the accelerator opening Acc is zero.

  Then, at the time T4 to T5 before the end of the shift, by switching to running by the engine 8, the torque of the second electric motor M2 is gradually reduced to zero. Note that the second electric motor M2 may output the motor torque according to the assist amount as it is, if motor assist is required even in the engine travel region.

  FIG. 11 is a flowchart for explaining a main part of the control operation of the electronic control unit 80, that is, a control operation of downshift accompanying engine start, which is repeatedly executed with a very short cycle time of, for example, about several milliseconds to several tens of milliseconds. It is.

  First, in step S1 corresponding to the downshift / engine start determining means 88 (hereinafter, step is omitted), whether the vehicle state requires a downshift, that is, whether the vehicle state crosses the downshift line indicated by the broken line in FIG. It is determined whether or not. If S1 is negative, in S6, for example, other shift control such as other shift control and lock-up control of the automatic transmission unit 20 is performed, and this routine is terminated.

If S1 is affirmed, it is determined in S2 corresponding to the downshift / engine start determining means 88 whether start of the engine 8 is requested. If S2 is negative, in S5, for example, the automatic transmission unit 20 should change gears based on the vehicle state indicated by the actual vehicle speed V and the required output torque T _OUT of the automatic transmission unit 20 from the automatic transmission map shown in FIG. The gear stage is determined, and automatic shift control of the automatic transmission unit 20 is executed so that the determined gear stage to be shifted is obtained.

If S2 is affirmed, in S3 corresponding to the underlap control means 86 and the synchronization control means 90, the torque capacities of the second brake B2 and the third brake B3 before the complete explosion of the engine 8 for clutch-to-clutch shift control are obtained. sum _{T TOTAL} is a pressure regulation control standby pressure PY standby pressure PX and the third brake B3 of the second brake B2 to be below the torque _{T 18} of the transmission member 18. Further, in the S3, by controlling the second electric motor M2, so is synchronously controlled rotational speed N ₁₈ of the transmitting member 18 to the rotational speed, which is the target of control during the shift end. Then, in S4 corresponding to the hybrid control means 84, generally well-known clutch-to-clutch shift control is executed, and this routine is terminated.

As described above, according to the present embodiment, in the shift of the automatic transmission unit 20 that accompanies the start of the engine 8, the torques that the second brake B2 and the third brake B3 have for the shift until the complete explosion of the engine 8 respectively. By controlling the engagement operation of the automatic transmission unit 20 so that the total capacity T _TOTAL is lower than the torque of the transmission member 18 at the time of the engine complete explosion, even if torque fluctuation due to the complete explosion of the engine 8 occurs, Since torque _{equal to} or greater than the total torque capacity T _TOTAL of the second brake B2 and the third brake B3 is not transmitted to the output shaft 22 (drive wheel 34) of the automatic transmission unit 20, torque fluctuation is transmitted to the drive wheel 34 side. It is suppressed and the shock due to engine start (complete explosion) is reduced.

Further, according to this embodiment, during the speed change by the underlap control means 86, the rotational speed N ₁₈ be to synchronously controlled by the second electric motor M2, so shift end the rotational speed N ₁₈ of the transmitting member 18 of the transmission member 18 It is quickly synchronized to the target rotational speed. Thereby, the delay of the synchronization of the rotational speed by the underlap control means 86 is suppressed, and the delay of the progress of the shift is suppressed.

Further, according to the present embodiment, by providing the standby pressures PX and PY that are set in advance to the release pressure of the second brake B2 and the engagement pressure of the third brake B3, the sum of the torque capacities is set to a suitable underlap value T. _By maintaining the _UNDER , even if torque fluctuation occurs due to the complete explosion of the engine 8, torque exceeding the _underlap value T _UNDER is not transmitted to the output shaft 22 (drive wheel 34) side of the automatic transmission unit 20. Further, transmission of torque fluctuation to the drive wheel 34 side is suppressed, and shock due to engine start (complete explosion) is reduced. Further, the total torque capacity T _TOTAL of the second brake B2 and the third brake B3 can be easily maintained at a predetermined underlap value T _UNDER by controlling the standby pressures PX and PY.

  Further, according to the present embodiment, the differential unit 11 and the automatic transmission unit 20 constitute a continuously variable transmission, and the drive torque can be changed smoothly. The differential unit 11 can be operated as an electric continuously variable transmission by continuously changing the gear ratio, and can be operated as a stepped transmission by changing the gear ratio stepwise. The overall transmission of the transmission mechanism 10 can be changed in stages, so that drive torque can be obtained quickly.

  Further, according to the present embodiment, for example, the continuously variable transmission is configured by the differential unit 11 that functions as an electrical continuously variable transmission and the stepped automatic transmission unit 20, and the drive torque changes smoothly. In the state where the gear ratio of the differential unit 11 is controlled to be constant, the differential unit 11 and the stepped automatic transmission unit 20 constitute a state equivalent to a stepped transmission. In addition, the overall transmission of the transmission mechanism 10 can be changed in stages, so that driving torque can be obtained quickly.

  As mentioned above, although the Example of this invention was described in detail based on drawing, this invention is applied also in another aspect.

  For example, the underlap control means 86 of the present embodiment is related to the release pressure of the second brake B2 that is the release-side hydraulic friction engagement device and the third brake B3 that is the engagement-side hydraulic friction engagement device. Although the hydraulic pressures for both the combined pressure and the standby pressure PX and PY are controlled, it is not particularly limited to both friction engagement devices, but the release side friction engagement device or the engagement side friction engagement The present invention can also be applied by controlling the hydraulic pressure of any one of the devices as a standby pressure.

  In the transmission mechanism 10 of the present embodiment, the release-side hydraulic friction engagement device is the second brake B2, and the engagement-side hydraulic friction engagement device is the third brake B3. Is applied in the present embodiment, and other hydraulic friction engagement devices may be applied depending on the structure of the automatic transmission unit 20 and the shift map.

  In addition, the second electric motor M2 of the present embodiment is directly connected to the transmission member 18, but the connection position of the second electric motor M2 is not limited thereto, and the power transmission path between the differential portion 11 and the drive wheels 34. May be connected directly or indirectly via a transmission or the like.

  In the above-described embodiment, the differential unit 11 functions as an electric continuously variable transmission whose gear ratio γ0 is continuously changed from the minimum value γ0min to the maximum value γ0max. The present invention can be applied even if the gear ratio γ0 of the moving portion 11 is not changed continuously but is changed stepwise using a differential action.

  In the above-described embodiment, the differential unit 11 includes a differential limiting device that is provided in the power distribution mechanism 16 and is operated as at least a two-stage forward transmission by limiting the differential action. It may be.

  In the power distribution mechanism 16 of the above-described embodiment, the first carrier CA1 is connected to the engine 8, the first sun gear S1 is connected to the first electric motor M1, and the first ring gear R1 is connected to the transmission member 18. However, the connection relationship is not necessarily limited thereto, and the engine 8, the first electric motor M1, and the transmission member 18 are connected to any of the three elements CA1, S1, and R1 of the first planetary gear device 24. It can be done.

  In the above-described embodiment, the engine 8 is directly connected to the input shaft 14. However, the engine 8 only needs to be operatively connected via, for example, a gear, a belt, or the like, and needs to be disposed on a common shaft center. Absent.

  In the above-described embodiment, the first motor M1 and the second motor M2 are arranged concentrically with the input shaft 14, the first motor M1 is connected to the first sun gear S1, and the second motor M2 is connected to the transmission member 18. However, the first motor M1 is operatively connected to the first sun gear S1 through, for example, a gear, a belt, a speed reducer, etc., and the second motor M2 is a transmission member. 18 may be connected.

  In the above-described embodiment, the automatic transmission unit 20 is connected in series with the differential unit 11 via the transmission member 18, but a counter shaft is provided in parallel with the input shaft 14 and is concentric on the counter shaft. In addition, the automatic transmission unit 20 may be arranged. In this case, the differential unit 11 and the automatic transmission unit 20 are coupled so as to be able to transmit power, for example, as a transmission member 18 via a pair of transmission members including a counter gear pair, a sprocket and a chain.

  Further, the power distribution mechanism 16 as the differential mechanism of the above-described embodiment includes, for example, a pinion that is rotationally driven by an engine and a pair of bevel gears that mesh with the pinion, the first electric motor M1 and the transmission member 18 (second electric motor M2). ) May be a differential gear device that is operatively coupled to.

  Further, the power distribution mechanism 16 of the above-described embodiment is composed of one set of planetary gear devices, but is composed of two or more planetary gear devices, and has three or more stages in the non-differential state (constant speed change state). It may function as a transmission. The planetary gear device is not limited to a single pinion type, and may be a double pinion type planetary gear device. Further, even when the planetary gear device is constituted by two or more planetary gear devices, the engine 8, the first and second electric motors M1 and M2, the transmission member 18, and the output depending on the configuration are provided to each rotating element of these planetary gear devices. The shaft 22 may be connected so as to be able to transmit power, and the stepped speed change and the stepless speed change may be switched by the control of the clutch C and the brake B connected to the rotating elements of the planetary gear device.

  In the above-described embodiment, the engine 8 and the differential unit 11 are directly connected. However, the engine 8 and the differential unit 11 are not necessarily connected directly, and are connected via a clutch between the engine 8 and the differential unit 11. May be.

  In the above-described embodiment, the differential unit 11 and the automatic transmission unit 20 are connected in series. However, the present invention is not limited to such a configuration, and the transmission mechanism 10 as a whole has an electrical difference. The present invention is applicable and mechanically independent as long as the structure includes a function for performing a movement and a function for performing a shift on a principle different from that based on an electric differential as a whole of the transmission mechanism 10. There is no need. Further, the arrangement position and arrangement order of these are not particularly limited.

In addition, the shift operating device 50 of the above-described embodiment includes the shift lever 52 operated to select a plurality of types of shift positions P _SH. Instead of the shift lever 52, for example, a push button type Switches that can select multiple types of shift positions P _SH , such as switches and slide switches, or devices and foot operations that can switch between multiple types of shift positions P _SH in response to the driver's voice regardless of manual operation it may be a plurality of shift positions P _SH is switched device such as a. In addition, the shift range is set by operating the shift lever 52 to the “M” position, but the gear stage is set, that is, the highest speed gear stage of each shift range is set as the gear stage. May be. In this case, the automatic transmission unit 20 performs gear shifting by switching the gear. For example, when the shift lever 52 is manually operated to the upshift position “+” or the downshift position “−” in the “M” position, the automatic transmission unit 20 selects any one of the first speed gear to the fourth speed gear. Is set according to the operation of the shift lever 52.

  The above description is only an embodiment, and the present invention can be implemented in variously modified and improved forms based on the knowledge of those skilled in the art.

BRIEF DESCRIPTION OF THE DRAWINGS FIG. 1 is a skeleton diagram illustrating a configuration of a hybrid vehicle drive device according to an embodiment of the present invention. FIG. 2 is an operation chart for explaining a combination of operations of a hydraulic friction engagement device used for a speed change operation of the drive device of FIG. 1. FIG. FIG. 2 is a collinear diagram illustrating a relative rotational speed of each gear stage in the drive device of FIG. 1. It is a figure explaining the input-output signal of the electronic controller provided in the drive device of FIG. It is a circuit diagram regarding the linear solenoid valve which controls the action | operation of each hydraulic actuator of the clutch C and the brake B among hydraulic control apparatuses. It is an example of the shift operation apparatus operated in order to select multiple types of shift positions provided with the shift lever. It is a functional block diagram explaining the principal part of the control function of the electronic control apparatus of FIG. It is a figure which shows an example of the shift map used in the shift control of a drive device, and an example of the drive force source map used in the drive force source switching control which switches engine driving | running | working and motor driving | running | working, Comprising: is there. A broken line is an optimal fuel consumption rate curve of the engine and is an example of a fuel consumption map. It is a time chart explaining the control action of the downshift accompanying engine starting. FIG. 5 is a flowchart for explaining a control operation of the electronic control unit of FIG. 4, that is, a control operation of downshift accompanying engine start.

Explanation of symbols

  8: Engine 10: Transmission mechanism (vehicle drive device) 11: Differential unit (electrical differential unit) 14: Input shaft 16: Power distribution mechanism (differential mechanism) 18: Transmission member (output shaft of differential mechanism) 20: Automatic transmission unit (transmission unit) 34: Drive wheel 86: Underlap control unit 90: Synchronous control unit B2: Second brake (engagement device, release side hydraulic friction engagement device) B3: Third brake ( Engagement device, engagement side hydraulic friction engagement device) M1: first electric motor M2: second electric motor

Claims (5)

The differential state between the rotational speed of the input shaft connected to the engine and the rotational speed of the output shaft is controlled by controlling the operating state of the engine and the first electric motor connected to the rotating element of the differential mechanism so as to transmit power. A control device for a vehicle drive device, comprising: an electric differential portion that is controlled; and a speed change portion that constitutes a part of a power transmission path from the electric differential portion to a drive wheel,
In the shifting of the transmission unit that accompanies the start of the engine, until the engine completes explosion, the sum of the torque capacities of the respective engagement devices related to the shifting is less than the torque of the output shaft at the time of engine explosion. The vehicle drive device control device further comprises an underlap control means for controlling the engagement device of the transmission unit. A second electric motor is connected to the output shaft of the differential mechanism,
2. The vehicle drive device control device according to claim 1, further comprising synchronization control means for synchronously controlling the rotation speed of the output shaft of the differential mechanism by the second electric motor during shifting by the underlap control means. . The shift of the transmission unit is a clutch-to-clutch shift in which the release of the release-side hydraulic friction engagement device and the engagement of the engagement-side hydraulic friction engagement device are executed at the same time.
The underlap control means provides the torque by providing a preset standby pressure as at least one of a release pressure of the release side hydraulic friction engagement device and an engagement pressure of the engagement side hydraulic friction engagement device. 3. The control device for a vehicle drive device according to claim 1, wherein the sum of the capacities is maintained at a predetermined underlap value.   4. The vehicle drive device according to claim 2, wherein the electric differential unit operates as a continuously variable transmission mechanism by controlling an operating state of the first and / or second electric motors. 5. Control device.   5. The control device for a vehicle drive device according to claim 1, wherein the transmission unit is a stepped automatic transmission.

Images