JP2010274705A - Control unit of power transmission device for vehicle - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To suppress interference of each feedback control which is performed independently in an electric type differential section and a gear change section, in performing simultaneous gear change in the electric type differential section and the gear change section, in a power transmission device for vehicle equipped with the electric type differential section and the gear change section. <P>SOLUTION: When performing simultaneous gear change in the differential section 11 and an automatic gear change section 20, the feedback control gain in connection with torque control of a first electric motor M1 during gear change is made smaller compared with a case when the gear change does not overlap. Thereby, since fluctuation of first electric motor torque T<SB>M1</SB>is further suppressed in feedback control in the differential section 11 and fluctuation of actual gear change section input torque T<SB>IN</SB>is further suppressed, deviation of a change gradient of an actual gear change section input rotation speed N<SB>AT</SB>from a target gradient is further suppressed in feedback control in the automatic gear change section 20. That is to say, the change of the actual gear change section input rotation speed N<SB>AT</SB>is stabilized, and feedback control of the automatic gear change section 20 is stably carried out. <P>COPYRIGHT: (C)2011,JPO&INPIT

Description

  The present invention relates to a control device for a vehicle power transmission device including an electric differential portion having a differential mechanism capable of performing a differential and a transmission portion constituting a part of a power transmission path, and more particularly to an electric difference. The present invention relates to the control when the shifts of the moving part and the transmission part overlap.

  An electric differential unit in which the differential state of the differential mechanism is controlled by controlling the operation state of the differential motor connected to the differential mechanism, and power transmission from the electric differential unit to the drive wheels 2. Description of the Related Art A vehicular power transmission device is well known that includes a traveling motor coupled to a path so as to be capable of transmitting power and a transmission that forms part of the power transmission path. For example, the power transmission device for vehicles described in patent document 1 is it. In this vehicle power transmission device, an electric power source includes a planetary gear device, a first motor (differential motor) connected to the sun gear of the planetary gear device, and a second motor (traveling motor) connected to the ring gear. And a stepped automatic transmission that is connected to the output side (ring gear) of the electric differential and that performs a shift by clutch-to-clutch control. By controlling, the differential state between the input rotational speed from the engine input from the carrier of the planetary gear device and the output rotational speed of the ring gear as the output member is controlled. In the electric differential section configured as described above, torque control of the differential motor is executed by feedback control so as to maintain the engine rotation speed (engine operating point) at a target value, for example. For example, in the shift control of the electric differential unit, the differential for maintaining the target engine rotation speed that is the engine operating point on the engine optimum fuel consumption line so that the engine output that can obtain the target output requested by the driver is obtained. The motor torque is controlled by feedback control. In addition, because of the configuration including the electric differential unit and the transmission unit, as shown in Patent Document 1, there is a case where a simultaneous shift in which the shift of the electrical differential unit and the transmission of the transmission unit overlap is executed. is there. Even in such a case, the same as the above in the shift of the electric differential unit so that the actual engine rotation speed (hereinafter referred to as the actual engine rotation speed) does not deviate from the target engine rotation speed due to the shift of the transmission unit. The differential motor torque is controlled by feedback control.

  Here, for example, in a transmission unit in which a shift is performed by a known clutch-to-clutch, each engagement device involved in the shift corresponds to the magnitude of the input torque of the transmission unit (hereinafter referred to as the transmission unit input torque). Engagement pressures on the engagement side and release side are set (commanded). For example, the standby pressure in the initial stage of the gear shift and the engagement pressure command value at the time of sweeping (gradual increase or decrease) are set by, for example, feedforward control based on the gear input torque. Further, the inertia phase by feedback control is performed so that the change gradient of the input rotational speed of the transmission unit (hereinafter referred to as the transmission unit input rotational speed) during the inertia phase of the transmission unit becomes, for example, a predetermined target gradient for suppressing the shift shock. The engagement pressure command value on the engagement side or the release side during the sweep in is corrected. Further, in the electric differential unit, the transmission unit input torque is mechanically transmitted to the output side via the electric differential unit, for example, when the differential motor is responsible for the reaction force of the engine torque. Calculated based on the direct torque and the traveling motor torque output by the traveling motor.

JP 2008-296610 A

  By the way, when simultaneous shifting of the electric differential unit and the transmission unit is executed, feedback control of the differential motor torque in the electric differential unit and the engagement side or release side sweep in the inertia phase in the transmission unit The two feedback controls including the feedback control of the hourly engagement pressure command value operate independently. That is, in the electric differential unit, the differential motor torque is changed by feedback control so as to suppress the deviation of the actual engine rotation speed from the target engine rotation speed due to the change in the transmission section input rotation speed accompanying the shift of the transmission section. There is a possibility that the actual transmission input torque (hereinafter, actual transmission input torque) may be changed. On the other hand, in the transmission unit, the calculated transmission unit input torque is the basis of the engagement pressure command value on the engagement side or the release side at the initial stage of the transmission set by the feedforward control, but during the sweep during the inertia phase. If the actual transmission unit input torque is changed during the shift without being reflected in the engagement pressure command value on the engagement side or the release side by feedback control, the actual transmission unit input rotational speed (hereinafter referred to as the actual transmission unit) is temporarily changed. There is a possibility that a change gradient of the input rotation speed) is greatly deviated from the target gradient, and the rotation change of the actual transmission unit input rotation speed is increased. Therefore, there is a possibility that the change in the differential motor torque is further increased by feedback control in the electric differential section, and the actual transmission section input torque is further changed. Further, this may cause the change gradient of the actual transmission unit input rotational speed to be further deviated from the target gradient in the transmission unit. As described above, when the two feedback controls interfere with each other, the control diverges, and there is a possibility that the desired engine rotation speed and the change gradient of the transmission unit input rotation speed cannot be controlled during the shift control. This can increase the shift shock. In addition, the fact that the change in the rotational speed of the actual transmission unit input rotation speed is increased and the change in the differential motor torque is increased by feedback control in the electrical differential unit means that, for example, supply and demand of power to the differential motor For power control in the power storage device, there is a possibility that the supply and demand of power to the differential motor temporarily exceeds the input / output limit of the power storage device. This may reduce the durability of the power storage device and the equipment related to power supply and demand. The above-described problem is not known.

  The present invention has been made against the background of the above circumstances. An object of the present invention is to provide a vehicle power transmission device including an electric differential unit and a transmission unit. It is an object of the present invention to provide a control device that can suppress interference of feedback control executed independently by an electric differential unit and a transmission unit at the time of simultaneous shifting.

  To achieve the above object, the gist of the present invention is that: (a) a differential mechanism connected to an engine so as to be able to transmit power and a differential motor connected so as to be able to transmit power to the differential mechanism; An electric differential unit in which the differential state of the differential mechanism is controlled by controlling the operation state of the differential motor, and a traveling motor connected to the drive wheels so that power can be transmitted, A power transmission device for a vehicle, comprising: a transmission that is connected in series to an output-side rotating member of the electric differential unit and that forms a part of a power transmission path from the electric differential unit to the drive wheel. (B) The torque control of the differential motor is executed by feedback control so that the engine rotation speed becomes the target engine rotation speed, and separately from the feedback control in the torque control, In the shifting process The shift control of the transmission unit is executed by feedback control so that the change gradient of the input rotation speed of the transmission unit is the target gradient, and (c) the shift of the electric differential unit and the shift of the transmission unit When the gears overlap, the feedback control gain related to the torque control of the differential motor during the gear shift is changed to be smaller than that when the gear shifts do not overlap.

  According to this configuration, in the vehicle power transmission device including the electric differential unit and the transmission unit, when the electric differential unit and the transmission unit are simultaneously shifted, torque control of the differential motor during the shift is performed. Since the feedback control gain involved is reduced compared to when the shifts do not overlap, the electric differential unit has a variation in the differential motor torque due to feedback control compared to when the feedback control gain is not reduced. This suppresses fluctuations in the actual transmission unit input torque. Then, by suppressing the fluctuation of the actual transmission unit input torque, the transmission unit is compared with the case where the feedback control gain is not reduced in the shift control in which the change gradient of the transmission unit input rotation speed by the feedback control is the target gradient. Thus, fluctuations in the change gradient of the actual transmission unit input rotational speed are suppressed. Therefore, the fluctuation of the differential motor torque is further suppressed in the feedback control in the electric differential section, and the fluctuation in the actual transmission section input torque is further suppressed. Therefore, in the feedback control in the transmission section, the actual transmission section input rotational speed is reduced. It is further suppressed that the change gradient is deviated from the target gradient. That is, the change in the actual transmission unit input rotational speed is stabilized, and the feedback control of the transmission unit is stably performed. In this way, it is possible to suppress interference of feedback control executed independently by the electric differential unit and the transmission unit. Therefore, the divergence of the control is suppressed, and the shift shock is suppressed by appropriately controlling the desired engine rotation speed and the change gradient of the transmission unit input rotation speed during the shift control. Moreover, since the interference of each feedback control executed independently by the electric differential unit and the transmission unit is suppressed and the fluctuation of the differential motor torque is further suppressed, for example, to the differential motor in the power storage device The power supply and demand is suppressed, and the durability of the power storage device and the equipment related to the power supply and demand is improved.

  Here, preferably, the feedback control gain related to the torque control is changed after the inertia phase is started (for example, during the inertia phase) in the shifting process of the transmission unit. In this way, for example, compared to changing the feedback control gain related to the torque control of the differential motor before the start of the inertia phase in the shifting process, the differential motor torque varies due to the change of the feedback control gain. Thus, it is possible to suppress fluctuations in the driving force that accompany fluctuations in the transmission unit input torque.

  Preferably, the feedback control gain related to the torque control is changed before the start of the inertia phase in the shifting process of the transmission unit. In this case, for example, compared to changing the feedback control gain related to the torque control of the differential motor after the inertia phase is started (for example, during the inertia phase), the differential motor is changed by changing the feedback control gain. It can be suppressed that the feedback controllability of the change gradient of the transmission unit input rotation speed is temporarily lowered in accordance with the variation of the transmission unit input torque due to the variation of the torque. That is, the feedback controllability of the change gradient of the transmission unit input rotational speed can be improved.

  Preferably, the change in the feedback control gain related to the torque control is canceled after the shift of the transmission unit is completed (that is, the original feedback control gain is restored). In this case, for example, when the change of the feedback control gain related to the torque control of the differential motor is canceled during the shift of the transmission unit, the change of the feedback control gain is canceled, so that the differential motor torque is reduced. It is possible to avoid the change gradient of the transmission unit input rotational speed from fluctuating with the variation of the transmission unit input torque due to the variation of the. Therefore, it is possible to suppress a shift shock accompanying a change in the change gradient of the transmission unit input rotation speed. Further, since the change of the feedback control gain is canceled after the shift is completed, not during the shift of the transmission unit, the effect of changing the feedback control gain during the shift of the transmission unit (particularly during the inertia phase) can be sufficiently obtained.

  Preferably, the change of the feedback control gain related to the torque control is canceled at the end of the shift of the transmission unit. In this way, for example, when the change of the feedback control gain related to the torque control of the differential motor is canceled after the shift of the transmission unit is completed, the change of the feedback control gain causes the differential motor torque to be changed. It is possible to prevent fluctuations in the driving force that accompany fluctuations in the transmission unit input torque due to fluctuations in gears after the gear change. In addition, since the change of the feedback control gain is canceled particularly during the shifting process of the transmission unit, the effect of changing the feedback control gain during the shift of the transmission unit (particularly during the inertia phase) can be sufficiently obtained.

  Preferably, the transmission unit is configured to selectively establish a plurality of gear stages by executing clutch-to-clutch shift by releasing the disengagement side engagement device and engaging the engagement side engagement device. At least one of the engagement devices is a stepped automatic transmission, and the feedback control related to the shift control of the transmission unit has a change gradient of the input rotation speed of the transmission unit in the shift process of the transmission unit as a target gradient. This engagement pressure control is executed by feedback control. According to this configuration, when the electric differential unit and the transmission unit are simultaneously shifted, the feedback control gain related to the torque control of the differential motor is reduced in the electric differential unit, thereby reducing the differential motor torque. The feedback control related to the engagement pressure control of the engagement device in the transmission unit as compared with the case where the feedback control gain related to the torque control is not reduced by suppressing the fluctuation and the fluctuation of the actual transmission unit input torque. In this case, fluctuations in the change gradient of the actual transmission unit input rotational speed are suppressed.

  Preferably, the shift of the transmission unit is a power-on downshift associated with an increase in the required amount of output to the vehicle, and the shift of the electric differential unit sets the engine operating point in advance. It is to be changed along the optimum fuel consumption line. In this way, at the time of the simultaneous shift in which the power-on downshift in the transmission unit and the shift in which the engine operating point is changed along the optimum fuel consumption line overlap in the electric differential unit, the electric differential unit and the transmission unit Interference of each feedback control executed independently can be suppressed.

  Preferably, the transmission unit selectively achieves a plurality of gear stages (shift stages) by selectively connecting the rotating elements of the plurality of sets of planetary gear devices by a friction engagement device. For example, it is composed of various planetary gear type multi-stage transmissions having four forward speeds, five forward speeds, six forward speeds, and more. As a friction engagement device in this planetary gear type multi-stage transmission, a hydraulic friction engagement device such as a multi-plate type, a single plate type clutch or brake engaged by a hydraulic actuator, or a belt type brake is widely used. The oil pump that supplies the hydraulic oil for engaging the hydraulic friction engagement device may be driven by a driving power source (engine) for driving to discharge the hydraulic oil. Alternatively, it may be driven by a dedicated electric motor arranged separately. Further, the clutch or brake may be an electromagnetic engagement device such as an electromagnetic clutch or a magnetic powder clutch in addition to the hydraulic friction engagement device.

  Preferably, the hydraulic control circuit including the hydraulic friction engagement device is responsive to, for example, supplying output hydraulic pressure of a linear solenoid valve directly to a hydraulic actuator (hydraulic cylinder) of the hydraulic friction engagement device. However, it is also possible to control the shift control valve by using the output hydraulic pressure of the linear solenoid valve as a pilot hydraulic pressure, and to supply hydraulic oil from the control valve to the hydraulic actuator.

  Preferably, one linear solenoid valve is provided, for example, corresponding to each of a plurality of hydraulic friction engagement devices. However, the linear solenoid valves are not engaged at the same time or controlled to be engaged or released. When there are a plurality of hydraulic friction engagement devices, various modes are possible, such as providing a common linear solenoid valve for them. In addition, it is not always necessary to control the hydraulic pressure of all the hydraulic friction engagement devices with the linear solenoid valve, and pressure control means other than the linear solenoid valve, such as duty control of the ON-OFF solenoid valve for part or all of the hydraulic control. You can go there. In this specification, “supplying hydraulic pressure” means “applying hydraulic pressure” or “supplying hydraulic oil controlled to the hydraulic pressure”.

  Preferably, an engine that is an internal combustion engine such as a gasoline engine or a diesel engine is widely used as the engine. Further, an electric motor or the like may be used in addition to this engine as an auxiliary driving power source.

  Preferably, the differential mechanism includes a first rotating element coupled to the engine, a second rotating element coupled to the differential electric motor, and a third rotating element coupled to the traveling electric motor. Is a device having three rotating elements. In this way, the differential mechanism is easily configured.

  Preferably, the differential mechanism is a single pinion type planetary gear device, the first rotating element is a carrier of the planetary gear device, and the second rotating element is a sun gear of the planetary gear device. The third rotating element is a ring gear of the 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.

1 is a skeleton diagram illustrating a configuration of a vehicle power transmission device to which a control device of the present invention is applied. 2 is an operation chart for explaining a relationship between a shift operation of an automatic transmission unit provided in the vehicle power transmission device of FIG. 1 and an operation combination of a hydraulic friction engagement device used therefor. FIG. 2 is a collinear diagram illustrating a relative rotational speed of each gear stage in the vehicle power transmission device of FIG. 1. It is a figure explaining the input-output signal of the electronic controller provided in the power transmission device for vehicles 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 circuits. It is an example of the shift operation apparatus operated in order to select the multiple types of shift position provided with the shift lever. It is a functional block diagram explaining the principal part of the control function by the electronic controller of FIG. In the vehicle power transmission device of FIG. 1, an example of a pre-stored shift diagram that is a basis for shift determination of the automatic transmission unit, and a pre-stored driving force source switching line for switching between engine travel and motor travel It is a figure which shows an example of a figure, Comprising: It is also a figure which shows each relationship. It is a figure which shows an example of the optimal fuel consumption rate curve of the engine of FIG. Explains the main control operation of the electronic control unit, that is, the control operation for suppressing the interference of each feedback control executed independently by the differential unit and the automatic transmission unit when the differential unit and the automatic transmission unit are simultaneously shifted. It is a flowchart to do. FIG. 11 is a time chart corresponding to the control operation of FIG. 10, and is an example in the case where the simultaneous shift of the 3 → 2 power-on downshift of the automatic transmission unit and the shift of the differential unit is performed. It is another flowchart equivalent to the flowchart of FIG. It is another flowchart equivalent to the flowchart of FIG.

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

  FIG. 1 is a skeleton diagram illustrating a vehicle power transmission device 10 (hereinafter, referred to as a power transmission device 10) to which the control device of the present invention is applied. The power transmission device 10 is suitably used for a hybrid vehicle. . In FIG. 1, a power transmission device 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 ( As shown in FIG. 7), an automatic transmission unit 20 as a transmission unit connected in series via a transmission member 18 in a power transmission path between the transmission unit 18 and an output rotating member connected to the automatic transmission unit 20 An output shaft 22 is provided in series. The power transmission device 10 is preferably used for, 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). For example, a driving power source connected to the engine 8 is provided between an engine 8 which is an internal combustion engine such as a gasoline engine or a diesel engine and a pair of drive wheels 34, and the power from the engine 8 is part of a power transmission path. Is transmitted to the pair of drive wheels 34 sequentially through a differential gear device (final reduction gear) 32 (see FIG. 7) and a pair of axles.

  Thus, in the power transmission device 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 power transmission device 10 is configured symmetrically with respect to its axis, the lower side is omitted in the skeleton diagram of FIG. The same applies to the following embodiments.

  The differential unit 11 is connected to the power distribution mechanism 16 and the power distribution mechanism 16 so as to be capable of transmitting power, and functions as a differential motor for controlling the differential state of the power distribution mechanism 16; It is an electric differential part provided with the 2nd electric motor M2 connected so that power transmission was possible so that it might rotate integrally with the transmission member 18. FIG. The transmission member 18 is an output side rotation member of the differential unit 11, but also corresponds to an input side rotation member of the automatic transmission unit 20.

  The first electric motor M1 and the second electric motor M2 are so-called motor generators having a function as a motor that generates mechanical driving force from electric energy and a function as a generator that generates electric energy from mechanical driving force. is there. In other words, in the power transmission device 10, the electric motor M can function as an alternative to the engine 8 that is the main power source, or as a power source (sub power source) that generates driving force for traveling together with the engine 8. Further, electric energy is generated by regeneration from the driving force generated by another power source and supplied to another electric motor M via an inverter 54 (see FIG. 7), or the electric energy is stored in the power storage device 56 (FIG. 7)).

  The first motor M1 has at least a generator (power generation) function for generating a reaction force, and the second motor M2 functions as a traveling motor that outputs driving force as a second driving force source for traveling. ) Provide at least a function. Preferably, each of the first electric motor M1 and the second electric motor M2 is configured such that the power generation amount as the generator can be continuously changed. The first electric motor M <b> 1 and the second electric motor M <b> 2 are provided in a case 12 that is a casing of the power transmission device 10, and are cooled by hydraulic oil of the automatic transmission unit 20 that is a working fluid of the power transmission device 10.

  The power distribution mechanism 16 is a differential mechanism that is coupled to the engine 8 so as to be able to transmit power, and includes a single pinion type differential unit planetary gear device 24 having a predetermined gear ratio ρ0 of about “0.416”, for example. The mechanical mechanism is configured as a main body and mechanically distributes the output of the engine 8 input to the input shaft 14. The differential unit planetary gear unit 24 includes a differential unit sun gear S0, a differential unit planetary gear P0, a differential unit carrier CA0 that supports the differential unit planetary gear P0 so as to rotate and revolve, and a differential unit planetary gear P0. The differential part ring gear R0 meshing with the differential part sun gear S0 is provided as a rotating element (element). The gear ratio ρ0 is ZS0 / ZR0 where ZS0 is the number of teeth of the differential sun gear S0 and ZR0 is the number of teeth of the differential ring gear R0.

In the power distribution mechanism 16, the differential carrier CA0 is connected to the input shaft 14, that is, the engine 8, the differential sun gear S0 is connected to the first electric motor M1, and the differential ring gear R0 is connected to the transmission member 18. ing. In the power distribution mechanism 16 configured in this way, the differential unit sun gear S0, the differential unit carrier CA0, and the differential unit ring gear R0, which are the three elements of the differential unit planetary gear unit 24, can be rotated relative to each other. Thus, the differential action is operable, that is, the differential action is enabled (differential state), so that the output of the engine 8 is distributed to the first electric motor M1 and the transmission member 18 and distributed. Since a part of the output of the engine 8 is stored with electric energy generated from the first electric motor M1 or the second electric motor M2 is rotationally driven, the differential unit 11 (power distribution mechanism 16) is electrically For example, the differential unit 11 is set to a so-called continuously variable transmission state (electric CVT state) by functioning as a differential device, and the rotation of the transmission member 18 is continuously changed regardless of the predetermined rotation of the engine 8. That is, when the power distribution mechanism 16 is in a differential state, the differential unit 11 is also in a differential state, and the differential unit 11 has a gear ratio γ0 (the rotational speed N _IN of the input shaft 14 / the rotational speed of the transmission member 18). N ₁₈ ) is in a continuously variable transmission state that functions as an electrical continuously variable transmission in which N ₁₈ ) is continuously changed from the minimum value γ0min to the maximum value γ0max. When the power distribution mechanism 16 is set to the differential state in this way, one or both of the operating states of the first electric motor M1 and the second electric motor M2 connected to the power distribution mechanism 16 (differential unit 11) so as to be able to transmit power are provided. By controlling (operating point), the differential state of the power distribution mechanism 16, that is, the differential state of the rotational speed of the input shaft 14 and the rotational speed of the transmission member 18 is controlled.

  The automatic transmission unit 20 (transmission unit) constitutes a part of a power transmission path from the engine 8 to the drive wheel 34, and includes a single pinion type first planetary gear unit 26 and a single pinion type second planetary gear unit. 28 is a planetary gear type multi-stage transmission that functions as a stepped automatic transmission in which a plurality of gear ratios are mechanically set in stages. The first planetary gear unit 26 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 S1 via the first planetary gear P1. The first ring gear R1 meshing with the first gear R1 has a predetermined gear ratio ρ1 of about “0.488”, for example. The second planetary gear device 28 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 has a predetermined gear ratio ρ2 of about “0.455”, for example. When the number of teeth of the first sun gear S1 is ZS1, the number of teeth of the first ring gear R1 is ZR1, the number of teeth of the second sun gear S2 is ZS2, and the number of teeth of the second ring gear R2 is ZR2, the gear ratio ρ1 is ZS1 / ZR1. The gear ratio ρ2 is ZS2 / ZR2.

  In the automatic transmission unit 20, the first sun gear S1 is connected to the transmission member 18 via the third clutch C3 and selectively connected to the case 12 via the first brake B1, and the first carrier CA1 and the second ring gear are connected. R2 is integrally connected to the transmission member 18 via the second clutch C2, and is selectively connected to the case 12 via the second brake B2, and the first ring gear R1 and the second carrier CA2 Are integrally connected to the output shaft 22, and the second sun gear S2 is selectively connected to the transmission member 18 via the first clutch C1. Further, the first carrier CA1 and the second ring gear R2 are connected to a case 12 which is a non-rotating member via a one-way clutch F1, and is allowed to rotate in the same direction as the engine 8 and is prohibited from rotating in the reverse direction. As a result, the first carrier CA1 and the second ring gear R2 function as rotating members that cannot rotate in reverse.

In the automatic transmission unit 20 configured as described above, for example, clutch-to-clutch shift is executed by releasing the disengagement side engagement device and engagement of the engagement side engagement device, and a plurality of gear stages (shift stages) are generated. By being established selectively, a transmission 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 stage. For example, as shown in the engagement operation table of FIG. 2, the first speed gear stage having a gear ratio of about “3.20” is established by the engagement of the first clutch C1 and the one-way clutch F. The first gear C1 and the first brake B1 are engaged to establish a second speed gear stage with a gear ratio of about “1.72”, and the first clutch C1 and the second clutch C2 are engaged to change the gear ratio. The third speed gear stage that is about “1.00” is established, and the fourth speed gear stage that is about “0.67” is established by engagement of the second clutch C2 and the first brake B1. Then, the reverse gear stage in which the gear ratio becomes about “2.04” is established by the engagement of the third clutch C3 and the second brake B2. Further, the neutral "N" state is established by releasing the first clutch C1, the second clutch C2, the third clutch C3, the first brake B1, and the second brake B2. In addition, the second brake B2 is engaged during the engine braking of the first gear.

  Thus, the power transmission path in the automatic transmission unit 20 is a combination of the operation of engagement and release of the first clutch C1, the second clutch C2, the third clutch C3, the first brake B1, and the second brake B2. Thus, the state is switched between a power transmission enabling state that enables power transmission through the power transmission path and a power transmission cutoff state that interrupts power transmission. That is, any one of the first to fourth gear stages and the reverse gear stage is established, so that the power transmission path is in a state capable of transmitting power, and none of the gear stages is established. When the neutral “N” state is established, the power transmission path is brought into a power transmission cutoff state.

  The first clutch C1, the second clutch C2, the third clutch C3, the first brake B1, and the second brake B2 (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 power transmission device 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. 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, the transmission unit input rotational speed), that is, the rotational speed of the transmission member 18 (hereinafter referred to as transmission member rotational speed N ₁₈ ). Is continuously changed, and a continuously variable transmission ratio width is obtained at the gear M. Therefore, the overall transmission gear ratio γT (= the rotational speed N _IN of the input shaft 14 / the rotational speed N _OUT of the output shaft 22) of the power transmission device 10 is obtained continuously, and a continuously variable transmission is configured in the power transmission device 10. The Overall speed ratio γT of the power transmission device 10 is the total speed ratio γT of the entire power transmission device 10 which is formed on the basis of the gear ratio gamma _AT gear ratio γ0 and the automatic transmission portion 20 of the differential portion 11 . 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 power transmission device 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, and either the first gear to the fourth gear or the reverse drive When the gear stage (reverse gear stage) is selectively established, a total gear ratio γT of the power transmission device 10 that changes in a substantially equal ratio is obtained for each gear stage. Therefore, a state equivalent to the stepped transmission is configured in the power transmission device 10.

FIG. 3 illustrates a gear stage in a power transmission device 10 including a differential unit 11 that functions as a continuously variable transmission unit or a first transmission unit and an automatic transmission unit 20 that functions as a stepped transmission unit or a second transmission unit. The collinear diagram which can represent on a straight line the relative relationship of the rotational speed of each rotation element from which a connection state differs for every 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, and 28 and a vertical axis indicating the relative rotational speed. horizontal line X1 of the lower of the horizontal line indicates the rotation speed zero, represents the rotational speed N _E of the engine 8 upper horizontal line X2 is linked to the rotational speed of "1.0", that is the input shaft 14, horizontal line XG (X3 ) indicates the rotational speed of the third rotating element RE3, which will be described later is input from the rotational speed N ₁₈ the differential portion 11 of the transmission member 18 to the automatic shifting portion 20.

  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 indicate the differential corresponding to the second rotation element (second element) RE2 in order from the left side. This shows the relative rotational speed of the differential part ring gear R0 corresponding to the part sun gear S0, the differential part carrier CA0 corresponding to the first rotational element (first element) RE1, and the third rotational element (third element) RE3. These intervals are determined according to the gear ratio ρ 0 of the differential planetary gear unit 24. Further, the four vertical lines Y4, Y5, Y6, and Y7 of the automatic transmission unit 20 indicate, in order from the left, the second sun gear S2 corresponding to the fourth rotation element (fourth element) RE4 and the fifth rotation element RE5 ( The first ring gear R1 and the second carrier CA2 connected to each other corresponding to the fifth element) are connected to the first carrier CA1 and the second ring gear R2 connected to each other corresponding to the sixth rotation element (sixth element) RE6. Represents the first sun gear S1 corresponding to the seventh rotation element (seventh element) RE7, and the distance between them is determined according to the gear ratios ρ1, ρ2 of the first and second planetary gear devices 26, 28, respectively. ing. In the relationship between the vertical axes of the nomogram, when 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 section 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 ρ0. Further, in the automatic transmission unit 20, the interval between the sun gear and the carrier is set to correspond to "1" for each of the first and second planetary gear devices 26 and 28, and the interval between the carrier and the ring gear corresponds to ρ. Set to the interval to be

  If expressed using the collinear diagram of FIG. 3 described above, the power transmission device 10 of the present embodiment is configured so that the power distribution mechanism 16 (differential unit 11) has the first rotating element RE1 ( The differential carrier CA0) 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 (differential ring gear R0) RE3 is connected to the transmission member 18 and the second rotating element RE2. It is connected to the electric motor M2, and is configured to transmit (input) the rotation of the input shaft 14 to the automatic transmission unit 20 via the transmission member 18. At this time, the relationship between the rotational speed of the differential section sun gear S0 and the rotational speed of the differential section ring gear R0 is shown 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 the difference indicated by the intersection of the straight line L0 and the vertical line Y3. When the rotational speed of the moving part ring gear R0 is substantially constant by being constrained by the vehicle speed V, the differential part sun gear indicated by the intersection of the straight line L0 and the vertical line Y1 is controlled by controlling the rotational speed of the first electric motor M1. When the rotation of S0 is raised or lowered, the rotational speed, or the engine rotational speed N _E of the carrier CA0, represented by an intersecting point between the straight line L0 and the vertical line Y2 is increased or decreased. The rotation of the differential portion sun gear S0 is the same speed as the engine speed N _E by controlling the rotational speed of the first electric motor M1 such speed ratio γ0 of the differential portion 11 is fixed to "1" If that, the straight line L0 is aligned with the horizontal line X2, the rotational speed, i.e., the power transmitting member 18 of the differential portion ring gear R0 at a speed equal to the engine speed N _E is rotated. Alternatively, by controlling the rotational speed of the first electric motor M1 so that the speed ratio γ0 of the differential section 11 is fixed to a value smaller than “1”, for example, about 0.7, the rotation of the differential section sun gear S0 becomes zero. Once, the straight line L0 is the state shown in FIG. 3, it is higher than the engine speed N _E and the power transmitting member 18 is rotated.

  Further, in the automatic transmission unit 20, the fourth rotation element RE4 is selectively connected to the transmission member 18 via the first clutch C1, the fifth rotation element RE5 is connected to the output shaft 22, and the sixth rotation element RE6 is the sixth rotation element RE6. It is selectively connected to the transmission member 18 via the second clutch C2 and selectively connected to the case 12 via the second brake B2, and the seventh rotating element RE7 is connected to the transmission member 18 via the third clutch C3. It is selectively connected to the case 12 via the first brake B1.

  In the automatic transmission unit 20, as shown in FIG. 3, when the first clutch C1 and the second brake B2 are engaged, the intersection of the vertical line Y4 indicating the rotational speed of the fourth rotation element RE4 and the horizontal line X3. And an oblique straight line L1 passing through the intersection of the vertical line Y6 indicating the rotational speed of the sixth rotational element RE6 and the horizontal line X1, and a vertical line Y5 indicating the rotational speed of the fifth rotational element RE5 connected to the output shaft 22. The rotational speed of the output shaft 22 at the first speed (1st) is shown at the intersection of. Similarly, at an intersection of an oblique straight line L2 determined by engaging the first clutch C1 and the first brake B1, and a vertical line Y5 indicating the rotational speed of the fifth rotating element RE5 connected to the output shaft 22. The rotation speed of the output shaft 22 at the second speed (2nd) is shown, and the fifth rotation connected to the output shaft 22 and the horizontal straight line L3 determined by engaging the first clutch C1 and the second clutch C2. The rotation speed of the output shaft 22 of the third speed (3rd) is indicated by the intersection with the vertical line Y5 indicating the rotation speed of the element RE5, and is determined by the engagement of the second clutch C2 and the first brake B1. 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 Y5 indicating the rotation speed of the fifth rotation element RE5 connected to the output shaft 22.

  FIG. 4 illustrates a signal input to the electronic control device 80 that is a control device for controlling the power transmission device 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 advance in the ROM while using a temporary storage function of the RAM. By performing the above, various controls such as the hybrid drive control for the engine 8 and each electric motor M and the shift control of the automatic transmission unit 20 are executed.

The electronic control unit 80 receives a signal representing the engine water temperature TEMP _W that is the temperature of the cooling fluid of the engine 8 and the shift position P _{SH of the} shift lever 52 (see FIG. 6) from each sensor and switch as shown in FIG. and a signal representative of the number of operations such as in the "M" position, a signal indicative of engine rotational speed N _E is the rotational speed of the engine 8, a signal for commanding the M mode (manual shift running mode), a signal representing the operation of an air conditioner, a vehicle speed A signal representing the vehicle speed V corresponding to the rotational speed N _OUT of the output shaft 22 detected by the sensor 72 and the traveling direction of the vehicle, a signal representing the hydraulic oil temperature T _OIL of the automatic transmission unit 20, a signal representing the side brake operation, wheels A well-known foot brake device (hoist wheel) as a braking device that applies brake torque (braking force) to (the drive wheels 34, driven wheels not shown). During operation of Le braking device) (i.e. the operation of the brake pedal indicating the foot during braking operation) (on) a brake operation signal representing a B _ON, a signal representative of the catalyst temperature, the driver detected by the accelerator opening sensor 78 output Accelerator opening signal indicating the accelerator opening Acc corresponding to the requested amount, an accelerator opening signal indicating the cam angle, a signal indicating the cam angle, a signal indicating the snow mode setting, a signal indicating the longitudinal acceleration G of the vehicle, and an auto cruise traveling A signal, a signal representing the weight of the vehicle (vehicle weight), a signal representing the wheel speed of each wheel, a rotational speed N _M1 of the first electric motor M1 detected by the M1 rotational speed sensor 74 comprising a resolver or the like (hereinafter referred to as “first” the second electric motor M detected by the motor rotation speed is expressed as N _M1 ") and a signal indicating the direction of rotation, consisting of a resolver M2 rotational speed sensor 76 Rotational speed N _{M2 (hereinafter,} "second electric motor speed N _M2" represent) and of a signal indicating the direction of rotation, the electric motor M1, charging and discharging via an inverter 54 with the M2 i.e. electric motor M1, A signal indicating the charging capacity (charging state) SOC of the power storage device 56 (see FIG. 7) supplying and supplying power to M2, a signal indicating the battery temperature TH _BAT of the power storage device (battery) 56, and the like are supplied.

From the electronic control unit 80, an engine output control unit 58 (see FIG. 7) for controlling the output P _{E of the} engine 8 (the unit is, for example, “kW”; hereinafter referred to as “engine output P _E ”). Control signal, for example, a drive signal to the throttle actuator 64 for operating the throttle valve opening θ _TH of the electronic throttle valve 62 provided in the intake pipe 60 of the engine 8, the intake pipe 60 by the fuel injection device 66 or the in-cylinder of the engine 8 A fuel supply amount signal for controlling the fuel supply amount to the engine, an ignition signal for instructing the ignition timing of the engine 8 by the ignition device 68, a supercharging pressure adjustment signal for adjusting the supercharging pressure, and an electric motor for operating the electric air conditioner Air conditioner drive signal, command signal for commanding operation of motors M1 and M2, shift position (operation position) display signal for operating shift indicator A gear ratio display signal for displaying the gear ratio, a snow mode display signal for displaying that the current mode is the snow mode, a wheel brake operation signal for operating the wheel brake device, and the M mode being selected. An electromagnetic valve (solenoid) included in the hydraulic control circuit 70 (see FIGS. 5 and 7) for controlling the M mode display signal to be displayed and the hydraulic actuator of the hydraulic friction engagement device of the differential unit 11 and the automatic transmission unit 20. valve command signals for actuating the valve) or the like, the original for the hydraulic control circuit signals for pressure regulating the line pressure P _L by a regulator valve (pressure regulating valve) provided in 70, the line pressure P _L is pressure adjusted Drive command signal for operating the electric hydraulic pump, which is the hydraulic pressure source, signal for driving the electric heater, cruise control system Signal or the like to use the computer, is outputted.

FIG. 5 shows linear solenoid valves SL1 to SL5 for controlling the operation of the hydraulic actuators (hydraulic cylinders) AC1, AC2, AC3, AB1, AB2 of the clutches C1, C2, C3 and brakes B1, B2 in the hydraulic control circuit 70. FIG. In FIG. 5, each hydraulic actuator AC1, AC2, AC3, AB1, AB2 has an engagement pressure (engagement hydraulic pressure) corresponding to a command signal from the electronic control unit 80 by the linear solenoid valves SL1-SL5. The pressure is regulated by PC1, PC2, PC3, PB1, and PB2 and supplied directly. This line hydraulic pressure PL is based on the hydraulic pressure generated from an electric oil pump (not shown) or a mechanical oil pump that is driven to rotate by the engine 8 as an original pressure, for example, by a relief type pressure regulating valve (regulator valve), and the accelerator opening Acc or throttle valve opening. The pressure is adjusted to a value corresponding to the engine load or the like represented by the degree θ _TH .

  The linear solenoid valves SL1 to SL5 have basically the same configuration and are excited and de-energized independently by the electronic control unit 80, and the hydraulic pressures of the hydraulic actuators AC1, AC2, AC3, AB1, and AB2 are independently regulated. Thus, the engagement pressures PC1, PC2, PC3, PB1, and PB2 of the clutches C1 to C3 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. Further, in the shift control of the automatic transmission unit 20, for example, the release and engagement of the clutch C and the brake B involved in the shift, that is, the release of the release side engagement device and the engagement of the engagement side engagement device are performed simultaneously. A so-called clutch-to-clutch shift to be controlled is executed.

FIG. 6 is a diagram showing 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 placed in a neutral state, that is, a neutral state in which the power transmission path in the power transmission device 10, that is, the automatic transmission unit 20 is interrupted, and is a parking position “P (” for locking the output shaft 22 of the automatic transmission unit 20. Parking) ”, reverse travel position“ R (reverse) ”for reverse travel, neutral position“ N (neutral) ”for achieving a neutral state in which the power transmission path in the power transmission device 10 is interrupted, power transmission device In the automatic shift control, a forward automatic shift travel position “D (drive)” for executing automatic shift control within a change range of 10 shiftable total speed ratio γT or a manual shift travel mode (manual mode) is established. Forward manual shift travel position “M (manual) for setting a so-called shift range that limits the high-speed gear stage It is provided so as to be manually operated to.

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 each shift position P _SH shown in the “P” to “M” positions (ranges), the “P” position and the “N” position are non-traveling positions (ranges) selected when the vehicle is not traveling. This is a non-driving position for selecting the switching of the power transmission path that disables driving of the vehicle in which the power transmission path in the automatic transmission unit 20 is blocked to the power transmission cut-off state. The “R” position, the “D” position, and the “M” position are travel positions (ranges) selected when the vehicle travels, and the vehicle to which the power transmission path in the automatic transmission unit 20 is connected. It is also a drive position for selecting switching to a power transmission enabled state of the power transmission path that enables driving.

  Specifically, when the shift lever 52 is manually operated to the “P” position, both the clutch C and the brake B are released, and the power transmission path in the automatic transmission unit 20 is set to a power transmission cutoff state. When the output shaft 22 of the automatic transmission unit 20 is locked and manually operated to the “N” position, both the clutch C and the brake B are released, and the power transmission path in the automatic transmission unit 20 is in the power transmission cut-off state. Then, by manually operating to any of the “R”, “D”, and “M” positions, any gear stage corresponding to each position is established, and the power transmission path in the automatic transmission unit 20 is established. Power transmission is possible.

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 unit, that is, the stepped shift control unit 82 functions as a shift control unit that shifts the automatic transmission unit 20. For example, the stepped shift control means 82 is stored in advance in the storage unit, that is, the storage means 84 with the vehicle speed V and the output torque T _OUT (or the accelerator opening Acc, etc.) of the automatic transmission unit 20 as shown in FIG. The required output torque T _OUT of the automatic transmission unit 20 corresponding to the actual vehicle speed V, accelerator opening Acc, etc. from the relationship (shift diagram, shift map) having an upshift line (solid line) and a downshift line (one-dot chain line). Based on the vehicle state indicated by the above, 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 determined shift stage is obtained. The automatic transmission control of the automatic transmission unit 20 is executed.

  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. A linear solenoid valve is actuated to actuate a hydraulic actuator of a hydraulic friction engagement device that is involved in the speed change.

  The hybrid control unit, that is, the hybrid control means 86, functions as engine drive control means for controlling the drive of the engine 8 via the engine output control device 58, and is driven by the first electric motor M1 and the second electric motor M2 via the inverter 54. A function as a motor operation control means for controlling the operation as a power source or a generator is included, and hybrid drive control by the engine 8, the first motor M1, and the second motor M2 is executed by these control functions.

Further, the hybrid control means 86 operates the engine 8 in an efficient operating range, while optimizing the reaction force due to the distribution of the driving force between the engine 8 and the second electric motor M2 and the power generation of the first electric motor M1. To change the gear ratio γ0 of the differential section 11 as an electrical continuously variable transmission. 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. The target output is calculated, and the target engine output (required engine output) _PER is calculated in consideration of transmission loss, auxiliary load, assist torque of the second electric motor M2, etc. so as to obtain the total target output. controlling the output or power of the electric motor M to control the engine 8 so that the output torque (engine torque) T _E of the engine rotational speed N _E and the engine 8 by the engine output P _ER is obtained.

Thus, overall speed ratio γT is the transmission ratio of the whole of the power transmission device 10 includes a gear ratio gamma _AT of the automatic transmission portion 20 controlled by the step-variable shifting control means 82 is controlled by the hybrid control means 86 The speed ratio γ0 of the differential unit 11 is determined. That is, the hybrid control means 86 and the stepped speed change control means 82 are within the range of the shift range corresponding to the shift position P _SH , the hydraulic control circuit 70, the engine output control device 58, the first electric motor M1, and the second electric motor M2. And the like, and functions as a transmission control means for controlling the overall transmission ratio γT, which is the transmission ratio of the power transmission device 10 as a whole.

For example, the hybrid control means 86 executes control of the engine 8 and each electric motor M in consideration of the gear position of the automatic transmission unit 20 in order to improve 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 86, for example, experimentally in advance as to achieve both drivability and fuel efficiency when continuously-variable shifting control in a two-dimensional coordinate composed of the engine rotational speed N _E and engine torque T _E The operating point of the engine 8 is calculated in an optimum fuel consumption rate curve (fuel consumption map, relationship) as an optimum fuel consumption line, which is a kind of operation curve of the engine 8 as shown by a broken line in FIG. (hereinafter, referred to as "engine operating point") to generate an engine output P _E required to satisfy as the engine 8 while is along is actuated, for example, the target output (total target output, required driving force) The target value of the total gear ratio γT of the power transmission device 10 is determined so that the engine torque T _E and the engine speed N _E are equal, and the target value is obtained. The output torque (hereinafter referred to as “first motor torque”) T _M1 of the first electric motor _M1 is changed by feedback control to control the transmission gear ratio γ0 of the differential section 11, and the total transmission gear ratio γT can be changed. Control within range. Here, the above-mentioned engine operating point, indicating the operating state of the engine rotational speed N _E and the engine 8 in a two-dimensional coordinates with coordinate axes state quantity indicating the operating state of the engine 8 is exemplified by such engine torque T _E operation Is a point. In the present embodiment, the fuel efficiency is, for example, a travel distance per unit fuel consumption, a fuel consumption rate (= fuel consumption / drive wheel output) of the entire vehicle, or the like.

  At this time, for example, the hybrid control means 86 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 a transmission member. However, a part of the motive power of the engine 8 is consumed for the electric power generation of the electric motor M, and is converted into electric energy there, and the electric energy is supplied to the other electric motor M through the inverter 54, and is supplied by the electric energy. The driving force output from the electric motor M is transmitted to the transmission member 18. A part of the motive power of the engine 8 is converted into electric energy by equipment related from generation of electric energy by the electric motor M related to power generation to consumption by the electric motor M related to driving, and the electric energy is converted into mechanical energy. An electrical path is formed until conversion.

Moreover, the hybrid control means 86 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. It controls the rotation of the engine rotational speed N _E to any rotational speed or maintained substantially constant. In other words, the hybrid control means 86, 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 86 as can be seen from the diagram of FIG. 3 when raising the engine rotation speed N _E during running of the vehicle, the second electric motor rotation speed N which depends on the vehicle speed V (driving wheels 34) The first motor rotation speed N _M1 is increased while maintaining _M2 substantially constant. The hybrid control means 86 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 .

The hybrid control means 86 controls the opening and closing of the electronic throttle valve 62 by the throttle actuator 64 for throttle control, and controls the fuel injection amount and injection timing by the fuel injection device 66 for fuel injection control. a command to control the ignition timing by the ignition device 68 such as an igniter for controlling alone or in combination with output to the engine output control device 58, an output control of the engine 8 so as to generate the necessary engine output P _E Execute. That is, it functions as an engine drive control means for controlling the drive of the engine 8.

For example, the hybrid controller 86 basically drives the throttle actuator 64 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 In addition, 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 86, and also performs 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 86 drives the second electric motor M2 for traveling without using the engine 8, for example, by the electric CVT function (differential action) of the differential section 11 regardless of whether the engine 8 is stopped or in an idle state. Motor traveling (EV mode traveling) can be performed as a power source. For example, the solid line A in FIG. 8 is used for switching the driving force source for starting / running the vehicle (hereinafter referred to as running) between the engine 8 and the electric motor, for example, the second electric motor M2, in other words, for running the engine 8. An engine travel region for switching between so-called engine travel for starting / running (hereinafter referred to as travel) the vehicle as a driving force source for the vehicle and so-called motor travel for traveling the vehicle using the second electric motor M2 as a drive power source for travel; It is a boundary line with a motor travel area. The pre-stored relationship having a boundary line (solid line A) for switching between engine running and motor running shown in FIG. 8 is a two-dimensional coordinate having the vehicle speed V and the output torque T _OUT of the automatic transmission unit 20 as variables. It is an example of the driving force source switching diagram (driving force source map) comprised by these. This driving force source switching diagram is stored in advance in the storage means 84 together with a shift diagram (shift map) indicated by, for example, the solid line and the alternate long and short dash line in FIG.

Then, the hybrid control means 86, for example, 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 driving force source switching diagram of FIG. And the motor running or the engine running is executed. As described above, the motor traveling by the hybrid control means 86 is relatively low output torque T _OUT (relatively low accelerator opening), which is generally considered to be poor in engine efficiency as compared with the high torque region as apparent from FIG. degree Acc) range, that is, a low engine torque T _E region, or is performed at a relatively low speed drive, that is, a low load region of the vehicle speed V.

Further, the hybrid control means 86 controls the first motor rotation speed N _M1 at a negative rotation speed so as to suppress dragging of the stopped engine 8 and improve fuel efficiency 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).

  In addition, the hybrid control means 86 is an electric energy and / or power storage device 56 from the first electric motor M1 by the electric path described above even in an engine driving region where the engine 8 is driven using the engine 8 as a driving power source for driving. The so-called torque assist for assisting the power of the engine 8 is possible by supplying the electric energy from the second motor M2 and driving the second motor M2 to apply torque to the drive wheels 34. Therefore, the engine traveling of the present embodiment includes a case where the engine 8 is used as a driving power source for traveling and a case where both the engine 8 and the second electric motor M2 are used as driving power sources for traveling. The motor traveling in this embodiment is traveling that stops the engine 8 and uses the second electric motor M2 as a driving force source for traveling.

  The hybrid control means 86 switches the operating state of the engine 8 between an operating state and a stopped state in order to switch between engine running and motor running, that is, an engine start / stop control unit that starts and stops the engine 8. A start / stop control means 88 is provided. The engine start / stop control means 88 starts or stops the engine 8 when the hybrid control means 86 determines, for example, switching between motor running and engine running based on the vehicle state from the driving force source map of FIG. To do.

For example, the engine start / stop control means 88 is configured such that the required output torque T _OUT increases as the accelerator pedal is depressed as shown by the point a → the point b of the solid line B in FIG. When it is determined that the travel region has changed to the engine travel region and it is determined that the motor travel is switched to the engine travel, that is, when the hybrid control means 86 determines that the engine is started, the first motor M1 is energized. to raising the first electric motor speed N _M1, i.e. it to function first electric motor M1 as a starter, complete combustion can be predetermined rotational speed N _{E 'for} example the idle speed more autonomous engine rotational speed N _E The engine rotation drive control is performed to increase the rotational speed to a predetermined autonomous rotational speed N _EIDL or more and Supplying fuel by the fuel injection device 66 (injector) and was ignited by the ignition device 68 to start the engine 8 by performing the engine torque generation control that generates engine torque T _E by the rolling speed N _{E 'above,} the motor driving Switch from engine to running. Further, the engine start / stop control means 88, as shown by the point b → point a of the solid line B in FIG. 8, the accelerator pedal is returned to reduce the required output torque T _OUT and the vehicle state is changed from the engine travel region to the motor travel region. In the case of changing to, the fuel supply is stopped by the fuel injection device 66, that is, the engine 8 is stopped by fuel cut, and the engine running by the hybrid control means 86 is switched to the motor running.

  Further, the hybrid control means 86 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 blocked. 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 86 can bring the differential unit 11 into 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 86 does not operate the engine 8 in order to improve fuel consumption (reduce the fuel consumption rate) during inertial running with the accelerator off (coast running) or wheel brake operation by operating the brake pedal. In the driving state, regenerative control is performed in which the kinetic energy of the vehicle transmitted from the drive wheels 34 is converted into electric energy by the differential unit 11. Specifically, the second motor M2 is rotationally driven by the reverse driving force transmitted from the drive wheel 34 to the engine 8 side to operate as a generator, and the electric energy, that is, the second motor generated current is passed through the inverter 54. Regenerative control for charging power storage device 56 is executed. That is, the hybrid control means 86 functions as a regeneration control means for executing the regeneration control.

Here, the first motor torque T is executed by the hybrid control means 86 described above so that the engine 8 is operated, for example, along an optimum fuel consumption rate curve (see FIG. 8) so as to set the engine operating point as a target value. _The feedback control for controlling _M1 will be described below.

While the engine is running, the hybrid control means 86 controls the shift of the differential portion 11, that is, the speed ratio γ 0 of the power distribution mechanism 16 so that the engine 8 operates along the optimum fuel efficiency curve of the engine 8, for example. . For this purpose, the hybrid control means 86 includes a target engine rotation speed determination unit, that is, a target engine rotation determination means 90. The target engine rotational determining means 90, for example, optimum fuel consumption curve, the accelerator opening Acc, based like the gear ratio gamma _AT vehicle speed V, and the automatic transmission portion 20, while the engine operating point is along the optimum fuel consumption curve, so that the engine torque T _E and the engine rotational speed N _E for generating the engine output P _E required to meet the required driving force corresponding to the accelerator opening Acc, the target value of the engine rotational speed N _E A certain target engine speed N _E ^* is determined in advance. Then, the hybrid control means 86, first a reaction torque which the engine speed N _E to counteract the engine torque T _E to be the predetermined target engine speed N _E ^* by the target engine rotation determination means 90 Feedback control for controlling the motor torque T _M1 is executed. Thus, the feedback control of the first electric motor torque T _M1 is executed by the hybrid control means 86 so that the engine rotational speed _NE becomes the target engine rotational speed _NE ^* , whereby the engine operation is displayed on the optimum fuel consumption rate curve. The engine 8 is operated so that the points follow.

For example, the first electric motor M1 has a feedback torque TFB _M1 (hereinafter referred to as “first motor”) by feedback control based on the control expression of the following expression (1) in order to converge the engine rotation speed _NE to the target engine rotation speed _NE ^* . Motor feedback torque TFB _M1 ") is generated and changed. In the control expression of the following formula (1), the first term on the right side is a proportional term, and the second term on the right side is an integral term. In the following formula (1), “KP” represents a proportional gain, and “KI” represents an integral gain. The proportional gain KP and the integral gain KI are obtained experimentally in advance and stored in the storage means 84 so that, for example, the response and stability of the first motor feedback torque TFB _M1 are compatible.
TFB _M1 = KP × (N _E ^* −N _E ) + KI × _Ｅ (N _E ^* −N _E ) dt (1)

The first electric motor torque T _M1 is a driving torque for transmitting the engine torque _TE to the drive wheels 34 in addition to the purpose of converging the engine rotational speed _NE to the target engine rotational speed _NE ^* as described above. That is, the first electric motor torque T _M1 is the reaction torque against the engine torque T _E to maintain the engine speed N _E when the engine rotational speed N _E is equal to the target engine speed N _E ^* it can be considered to be represented by the sum of the first electric motor feedback torque TFB _M1 which becomes zero when the engine rotational speed N _E is equal to the target engine rotational speed N _E ^*. Therefore, the hybrid control means 86 determines the first motor torque T _M1 including the first motor feedback torque TFB _M1 based on the equation (1), so that the engine rotational speed _NE becomes the target engine rotational speed N _E ^*. It can be said that the feedback control for controlling the first motor torque T _M1 is executed so that

The output side of the differential portion 11 (power distributing mechanism 16) is the engine torque T _E via the differential unit 11 by withstand the reaction force of the engine torque T _E by the first electric motor torque T _M1 (i.e. transmitting member Torque is mechanically transmitted to 18). In this embodiment, it referred to the torque the mechanically transmitted Jikatachi torque T _D. The feedthrough torque T _D is the relationship of the following formula (2) the first electric motor torque T _M1 and the engine torque T _E and the direct torque T _D in the differential portion planetary gear set 24 having a gear ratio .rho.0 ( 1 / ρ0) × T _M1 . Incidentally, the direction of action of the torque with the direct torque T _D and the first electric motor torque T _M1 from the relationship acting direction of the torque of each rotating element of the differential portion planetary gear set 24 is in the opposite direction.
T _M1 : T _E : T _D = ρ0 / (1 + ρ0): 1: 1 / (1 + ρ0) (2)

Further, in the differential unit 11, the input torque (transmission portion input torque) T _IN of the automatic transmission portion 20, for example based on the second electric motor torque T _M2 minutes the direct torque T _D min and the second electric motor M2 to output Is calculated as shown in the following equation (3). This expression (3) is shown as an approximate value of the transmission unit input torque T _IN excluding the inertia component in each torque of the second motor torque T _M2 and the direct torque T _D (first motor torque T _M1 ). ing.
T _IN ≈T _M2 + T _D = T _M2 − (1 / ρ0) × T _M1 (3)

On the other hand, apart from the feedback control for controlling the first motor torque T _M1 , in this embodiment, the change gradient of the transmission unit input rotational speed N _AT (that is, the transmission member rotational speed N ₁₈ ) in the shifting process of the automatic transmission unit 20. Shift control of the automatic transmission unit 20 is executed by feedback control so that is set as a target value. For example, the feedback control related to the shift control of the automatic shifting portion 20, at least one hydraulic involved a change gradient of the shifting portion input rotation speed N _AT in the shift process of the automatic transmission portion 20 to shift to a predetermined target slope The hydraulic control of the type frictional engagement device is executed by feedback control. More specifically, in the automatic shifting portion 20, the hydraulic pressure command value for the hydraulic friction engagement device in response to the magnitude of the shift unit input torque T _IN involved in clutch-to-clutch shift is set (Fig. 11). For example, the hydraulic pressure command value when pressure standby pressure and sweep (increasing or decreasing) the speed change process early inertial phase before the automatic shifting portion 20, experimentally determined in advance based on the shift unit input torque T _IN by the feed forward control Set to the specified value. Then, to a predetermined target slope previously obtained experimentally for the change gradient of the shifting portion input rotation speed N _AT is to suppress the example shift shock in the inertia phase of the automatic shifting portion 20, a sweep in the inertia phase The release side hydraulic pressure command value at that time (that is, the sweep rate of the release side hydraulic pressure command value) is corrected by feedback control.

By the way, the feedback control of the first motor torque _TM1 based on the control expression of the above formula (1) is executed not only during the non-shifting of the automatic transmission unit 20, but also during the shifting of the automatic transmission unit 20. That is, there is a case where a simultaneous shift in which the shift of the differential unit 11 and the shift of the automatic transmission unit 20 overlap is executed. For example, when the shift control of the automatic transmission portion 20 is performed by the step-variable shifting control means 82, the power at the shifting back and forth with that gear ratio gamma _AT of the automatic transmission portion 20 is changed stepwise The total gear ratio γT of the transmission device 10 is changed stepwise. In such speed change control, it is possible to change the drive torque more quickly than the continuous change in the total speed ratio γT. On the other hand, there is a possibility that the shift shock may occur, fuel economy can not control the engine rotational speed N _E along the optimum fuel consumption curve deteriorate. Therefore, the hybrid control means 86 synchronizes with the shift control of the automatic transmission unit 20 so as to continuously change the total transmission ratio γT of the power transmission device 10 before and after the automatic transmission unit 20 shifts. Shift control is executed.

For example, the drivability better during an upshift of the accelerator opening Acc constant of the automatic shifting portion 20 as shown in point c → point d of the solid line C which is the engine rotational speed N _E does not change in FIG. 8 is considered good. Therefore, the hybrid control means 86 does not change the total gear ratio γT of the power transmission device 10 transiently before and after the automatic transmission unit 20 shifts, and the engine operating point (engine torque T _E and engine rotation speed N _E ) is substantially constant. so as to maintain, change the speed ratio γ0 in the opposite direction to the variation only the change direction in synchronization with the shifting action of the automatic transmission portion 20 corresponding to the gradual change of the gear ratio gamma _aT of the automatic shifting portion 20 The shift control of the differential unit 11 is executed. That is, the hybrid control means 86 performs the shift of the differential unit 11 by so-called equal power shift with the shift of the automatic transmission unit 20. On the other hand, for example, a power-on downshift of the automatic shifting portion 20 due to the accelerator pedal is largely stepped operated demanded output torque T _OUT has been increased as shown in point a → point b of the solid line B in FIG. 8, it is necessary to generate an engine output P _E corresponding to the accelerator opening Acc. Further, in view from the engine rotational speed N _E of the responsiveness drivability better to ASAP increase is considered good. Therefore, the hybrid control means 86, for example, when the power-on downshift of the automatic shifting portion 20 is executed, the engine operating point for generating the engine output P _E corresponding to the accelerator opening Acc and (engine torque T _E Shifting of the differential unit 11 is executed to change the engine operating point along the optimum fuel consumption rate curve so that the engine rotational speed N _E ).

As described above, when the simultaneous shift of the shift of the differential unit 11 and the shift of the automatic transmission unit 20 is executed, the feedback control of the first motor torque _TM1 in the differential unit 11 and the inertia phase in the automatic transmission unit 20 are performed. The two feedback controls including the release side hydraulic pressure command value feedback control are operated independently. That is, the differential unit 11 suppresses the deviation of the actual engine rotational speed _{NE from} the target engine rotational speed _NE ^{* when} the transmission unit input rotational speed _NAT is changed with the shift of the automatic transmission unit 20. In addition, since the first motor torque _TM1 is changed by feedback control, the actual transmission unit input torque _TIN is changed. On the other hand, in the automatic transmission unit 20, when the actual transmission unit input torque _TIN is changed during the shift, the precondition of the transmission torque capacity in the feedback control of the release side hydraulic pressure command value is changed, and the actual transmission unit gradient change in the input rotational speed N _aT is likely to rotational change temporarily allowed to deviate significantly from the target gradient actual gear portion input rotation speed N _aT is increased. Therefore, there is a possibility that the change in the first motor torque T _M1 is further increased by the feedback control in the differential unit 11 and the actual transmission unit input torque T _IN is further greatly changed. In addition, the fact, in the automatic shifting portion 20, can change the slope of the actual gear portion input rotation speed N _AT rotation change of being allowed to deviate further increased from the target gradient actual gear portion input rotation speed N _AT is further increased There is sex. As described above, the two feedback controls that are operated independently interfere with each other, so that the control is diverged, and the target engine rotational speed _NE ^* and the target input rotational speed _NAT are changed to the target gradient at the time of simultaneous shift. It may become impossible to control properly. This can increase the shift shock. Further, the change in the first electric motor torque T _M1 is further increased by the feedback control in the differential unit 11, for example, due to the electric power (power) control in the power storage device 56, the supply and demand of electric power to the first electric motor M 1 is temporary. Therefore, the input / output limit of the power storage device 56 may be exceeded. As a result, there is a possibility that the durability of the equipment related to the power supply and demand such as the power storage device 56 and the inverter 54 may decrease.

Therefore, in this embodiment, when the shift of the differential unit 11 and the shift of the automatic transmission unit 20 overlap, the feedback control of the first motor torque T _M1 in the differential unit 11 that is operated independently and automatic When the shift does not overlap the feedback control gain related to the torque control of the first electric motor M1 during the shift in order to suppress the interference between the two feedback controls with the feedback control of the release side hydraulic pressure command value in the shift unit 20 Change smaller than For example, when the shift of the differential unit 11 and the shift of the automatic transmission unit 20 overlap, an experiment is performed in advance so that the responsiveness and stability of the first motor feedback torque TFB _M1 are compatible with each other compared to the case where the shift does not overlap. At least one of the proportional gain KP and the integral gain KI in the equation (1) that is obtained and stored is changed to be small.

More specifically, returning to FIG. 7, the traveling state determination unit, that is, the traveling state determination unit 92 determines, for example, whether the vehicle is traveling in the “D” range, which is the forward traveling range, based on the shift position _PSH. To do. In addition, the traveling state determination unit 92 determines whether or not the stepped shift control unit 82 has determined the shift of the automatic transmission unit 20 and has output a hydraulic pressure command value for executing the determined shift of the automatic transmission unit 20. Determine whether. Further, the traveling state determination unit 92 determines whether or not the shift when the hydraulic pressure command value is output by the stepped shift control unit 82 is a power-on downshift, for example, the accelerator opening Acc (or throttle valve opening). It is determined based on whether or not the change in the degree θ _TH or the like is a change equal to or greater than the accelerator-on determination value for determining the power-on downshift that has been experimentally obtained and stored in advance. That is, the traveling state determination unit 92 determines whether or not a hydraulic pressure command value for power-on downshift of the automatic transmission unit 20 has been output. Since the shift determination of the automatic transmission unit 20 by the stepped shift control unit 82 and the output of the hydraulic command value for the shift have a one-to-one relationship, the traveling state determination unit 92 Instead of determining whether or not a value has been output, for example, it may be determined whether or not the shift of the automatic transmission unit 20 has been determined by the stepped shift control means 82.

Further, the traveling state determination unit 92 determines whether or not the hybrid control unit 86 executes the shift control of the differential unit 11. For example, when the travel state determination unit 92 determines that the hydraulic command value for executing the shift of the automatic transmission unit 20 is output by the stepped shift control unit 82, the automatic transmission unit by the stepped shift control unit 82 It is determined whether or not the shift control of the differential section 11 is executed by the hybrid control means 86 in synchronization with the 20 shift control. Specifically, when the shift state when the traveling state determination unit 92 determines that the hydraulic pressure command value is output by the stepped shift control unit 82 is a power-on downshift, the hybrid control unit 86 controls the accelerator opening degree. along the optimum fuel consumption curve so that the engine operating point for generating the engine output P _E corresponding to the Acc determines whether the shift of the differential portion 11 to operate the engine 8 is executed. On the other hand, when the shift state when the hydraulic pressure command value is determined to be output by the stepped shift control unit 82 is a shift other than the power-on downshift, the running state determination unit 92 uses the hybrid control unit 86 to change the automatic transmission unit. It is determined whether or not the simultaneous shift of the differential unit 11 accompanying the 20 shifts is executed by a so-called equal power shift.

  The gain changing unit, that is, the gain changing unit 94, determines that the hydraulic pressure command value for executing the shift of the automatic transmission unit 20 is output by the traveling state determination unit 92 and the shift control of the differential unit 11 is executed. In such a case, a gain change command for changing the proportional gain KP and the integral gain KI in the equation (1) to be smaller than when the shift of the differential unit 11 and the shift of the automatic transmission unit 20 do not overlap. Output to the hybrid control means 86. In accordance with the gain change command, the hybrid control unit 86, for example, sets the proportional gain KP and the integral gain KI in the formula (1) to a predetermined second proportional gain KP ′ and a second integral smaller than the proportional gain KP and the integral gain KI. Change to gain KI '. The predetermined second proportional gain KP ′ and second integral gain KI ′ are previously tested for suppressing the divergence of control by suppressing interference between the two feedback controls in the differential unit 11 and the automatic transmission unit 20, for example. The feedback control gain is smaller than the proportional gain KP and the integral gain KI that are calculated and stored in the storage means 84.

  FIG. 10 shows the main part of the control operation of the electronic control unit 80, that is, the interference of each feedback control executed independently by the differential unit 11 and the automatic transmission unit 20 when the differential unit 11 and the automatic transmission unit 20 are simultaneously shifted. 5 is a flowchart for explaining a control operation for suppressing the above-mentioned, and is repeatedly executed with an extremely short cycle time of, for example, about several milliseconds to several tens of milliseconds. FIG. 11 is a time chart corresponding to the control operation of FIG. 10, and is an example in the case where the simultaneous shift of the 3 → 2 power-on downshift of the automatic transmission unit 20 and the shift of the differential unit 11 is performed.

In FIG 10, first, steps corresponding to the running state determining means 92 (hereinafter, omitted step) In S10, whether the vehicle is traveling in a forward running range based on the shift position P _SH "D" range Is determined. If the determination in S10 is affirmative, in step S20 corresponding to the traveling state determination unit 92, the stepped shift control unit 82 determines the shift of the automatic transmission unit 20, and the determined shift of the automatic transmission unit 20 is changed. It is determined whether or not a hydraulic pressure command value for execution is output. For example, it is determined whether a power-on downshift of the automatic transmission unit 20 is determined and a hydraulic pressure command value for executing the determined power-on downshift of the automatic transmission unit 20 is output. That is, in S20, it is determined whether or not a hydraulic pressure command value for power-on downshift of the automatic transmission unit 20 has been output. If the determination in S20 is affirmative, in S30 corresponding to the traveling state determination unit 92, the hybrid control unit 86 determines whether or not the shift control of the differential section 11 is executed. For example, when the hydraulic pressure command value for the power-on downshift of the automatic transmission portion 20 in the above S20 is judged to have been output, the engine operation to generate an engine output P _E corresponding to the accelerator opening Acc It is determined whether or not the shift of the differential unit 11 that operates the engine 8 is performed along the optimum fuel consumption rate curve so as to be a point. If the determination in S30 is affirmative, in S40 corresponding to the gain changing means 94, the engine speed _NE is set to the target as compared with the case where the shift of the differential unit 11 and the shift of the automatic transmission unit 20 do not overlap. The feedback control gain when the feedback control of the first motor torque T _M1 is executed is changed to be small so that the engine rotation speed N _E ^* is obtained. For example, a gain change command for changing the proportional gain KP and the integral gain KI in the equation (1) smaller than when the shift of the differential unit 11 and the shift of the automatic transmission unit 20 do not overlap is output. Then, in accordance with the gain change command, the proportional gain KP and integral gain KI in the equation (1) are changed to predetermined second proportional gain KP ′ and second integral gain KI ′ that are smaller than the proportional gain KP and integral gain KI. Is done. If any of the determination in S10, the determination in S20, and the determination in S30 is negative, in S50, for example, the proportional gain KP and the integral gain KI in the equation (1) are set to the second proportionality. Other normal control other than changing to the gain KP ′ and the second integral gain KI ′ is executed. For example, the shift control of the differential unit 11 alone is executed using the proportional gain KP and the integral gain KI.

In FIG. 11, the solid line indicates the second proportional gain KP ′ and the second integral gain KI ′ in the feedback control of the first motor torque _TM1 , in which the feedback control gain in the equation (1) is smaller than the proportional gain KP and the integral gain KI. It is an example when it is changed to. Further, the broken line is an example in the case where the proportional gain KP and the integral gain KI are used as they are for the feedback control gain in the equation (1) in the feedback control of the first motor torque T _M1 .

The time point t1 indicates that a power-on downshift of the automatic transmission unit 20 has been determined and a shift hydraulic pressure command for a power-on downshift by clutch-to-clutch control has been output. The solid line indicates the engagement side hydraulic pressure command value in the shift hydraulic pressure command, and the two-dot chain line indicates the release side hydraulic pressure command value. Then, after the torque phase is started as shown at time t2, the inertia phase is started as shown at time t3. In this inertia phase, for example, the sweep rate of the release side hydraulic pressure command value is fed back so that the change gradient of the transmission unit input rotation speed N _AT (transmission member rotation speed N ₁₈ , second motor rotation speed N _M2 ) becomes the target gradient. It is corrected by control. At this time, the engine rotational speed N _E with the increase of the shifting portion input rotation speed N _AT is raised. Then, the first electric motor torque T _M1 is feedback-controlled so that the engine rotational speed _NE becomes the target engine rotational speed _NE ^* .

In the feedback control of the first motor torque T _M1 , when the proportional gain KP and the integral gain KI are used as they are for the feedback control gain in the equation (1), as shown by the broken line, the first motor torque that is changed by the feedback control. T _M1 is increased. Since the change in the actual speed section input torque T _IN and the first electric motor torque T _M1 is increased is increased, variation gradient of the actual shifting-portion input rotation speed N _AT is temporarily allowed to deviate significantly from the target gradient. Therefore, the actual gear portion input torque T _IN changes is further increased in the first electric motor torque T _M1 is varied even greater by the feedback control in the differential unit 11, a change gradient of the actual gear portion input rotation speed N _AT rotation change of the actual gear portion input rotation speed N _AT is further increased is to deviate further increased from the target gradient. In this way, the two feedback controls that are independently operated by the differential unit 11 and the automatic transmission unit 20 interfere with each other, so that the control diverges, and the differential unit 11 and the automatic transmission unit 20 It can not be properly controlled to a target slope of the target engine speed N _E ^* and the transmission portion input rotation speed N _aT during concurrent shifting.

On the other hand, in the feedback control of the first motor torque T _M1 , the second proportional gain KP ′ and the second integral gain KI ′, which are smaller than the proportional gain KP and the integral gain KI, in the feedback control gain in the equation (1). When used, as shown by the solid line, the first motor torque T _{M1 that} is changed by feedback control is suppressed. Since the change in the actual speed section input torque T _IN is suppressed and the first electric motor torque T _M1 is suppressed, the variation gradient of the actual shifting-portion input rotation speed N _AT deviates from the target gradient is suppressed. Therefore, the change of the first electric motor torque T _M1 change is further suppressed of actual gear unit input torque T _IN in the feedback control in the differential unit 11 is also further suppressed, variation gradient of the actual gear portion input rotation speed N _AT the target Deviation from the gradient is further suppressed. That is, change of the actual shifting-portion input rotation speed N _AT is allowed to stabilize, the feedback control in the automatic shifting portion 20 is carried out stably. In this way, interference of feedback control executed independently by the differential unit 11 and the automatic transmission unit 20 is suppressed. In the embodiment of FIG. 11, the second proportional gain KP ′ and the second proportional gain KP and the integral gain KI are smaller than the proportional gain KP and the integral gain KI during the inertia phase in the shifting process of the automatic transmission unit 20. The integral gain is changed to KI ′.

As described above, according to this embodiment, in the power transmission device 10 including the differential unit 11 and the automatic transmission unit 20, the simultaneous transmission of the differential unit 11 and the automatic transmission unit 20 is performed during the shift. 1 Since the feedback control gain related to the torque control of the electric motor M1 is made smaller than when the shifts do not overlap, the differential unit 11 performs the first control by the feedback control compared to when the feedback control gain is not made smaller. variations in motor torque T _M1 is suppressed fluctuation of the actual gear portion input torque T _iN is suppressed. Then, when the variation of the actual gear portion input torque T _IN is suppressed, the automatic transmission portion 20, in the shift control to the target slope gradient change of the transmission portion input rotation speed N _AT by the feedback control, the feedback control gain compared with when not small variations in changes in the gradient of the actual shifting-portion input rotation speed N _AT is suppressed. Accordingly, in the feedback control in the differential unit 11, the variation in the first motor torque T _M1 is further suppressed and the variation in the actual transmission unit input torque T _IN is further suppressed, so in the feedback control in the automatic transmission unit 20, the actual transmission unit the variation gradient of the input rotational speed N _AT is allowed to deviate from the target gradient is further suppressed. That is, change of the actual shifting-portion input rotation speed N _AT is allowed to stabilize, the feedback control of the automatic transmission portion 20 is implemented stably. In this way, it is possible to suppress interference of feedback control executed independently between the differential unit 11 and the automatic transmission unit 20. Therefore, the divergence of the control is inhibited, are appropriately controlled to change the gradient of the desired engine speed N _E and the transmission portion input rotation speed N _AT during a shift control, the shift shock is suppressed. In addition, since interference between the feedback controls executed independently by the differential unit 11 and the automatic transmission unit 20 is suppressed and fluctuations in the first motor torque _TM1 are further suppressed, for example, the first in the power storage device 56 The supply and demand of electric power to the electric motor M1 is suppressed, and the durability of the devices related to the electric power supply and demand such as the power storage device 56 and the inverter 54 is improved.

Further, according to this embodiment, the feedback control related to the shift control of the automatic transmission portion 20, automatic transmission portion 20 changes the gradient of the shifting portion input rotation speed N _AT in the speed change process of at least one to a target slope of The hydraulic control of the hydraulic friction engagement device is executed by feedback control. In this manner, when the differential unit 11 and the automatic transmission unit 20 are simultaneously shifted, the first motor torque T _M1 is reduced by reducing the feedback control gain related to the torque control of the first motor M1 in the differential unit 11. the variation is suppressed in the variation of the actual gear portion input torque T _iN is suppressed, as compared to when the feedback control gain related to torque control is not reduced, the hydraulic friction engagement device in the automatic shifting portion 20 variations in gradient of change actual speed portion input rotation speed N _aT is suppressed in the feedback control related to the hydraulic control.

  Further, according to the present embodiment, the shift of the automatic transmission unit 20 is a power-on downshift accompanying the increase in the required output amount for the vehicle, and the shift of the differential unit 11 is preset with the engine operating point. It will be changed along the optimum fuel consumption line. In this way, at the time of the simultaneous shift in which the power-on downshift in the automatic transmission unit 20 and the shift in which the engine operating point is changed along the optimum fuel consumption line overlap in the differential unit 11, the differential unit 11 and the automatic transmission unit 20 The interference of each feedback control executed independently can be suppressed.

  Next, another embodiment of the present invention will be described. In the following description, parts common to the embodiments are denoted by the same reference numerals and description thereof is omitted.

  In the present embodiment, the timing (timing) at which the feedback control gain in Equation (1) is changed to the second proportional gain KP ′ and the second integral gain KI ′ smaller than the proportional gain KP and the integral gain KI will be considered.

By changing the feedback control gain from the proportional gain KP and the integral gain KI to the second proportional gain KP ′ and the second integral gain KI ′, the first motor torque T _M1 is changed (varied) in the feedback control and the actual speed change is performed. The part input torque _{TIN is} also changed. Thereby, the drive torque (drive force) in the drive wheel 34 is also changed. Change in the drive torque accompanying the change of such actual gear unit input torque T _IN (driving force) is capable suppressed as compared to the other during the inertia phase if the inertia phase in the shift process of the automatic transmission portion 20 it is conceivable that. Therefore, in this embodiment, in order to suppress a change in the driving torque (driving force), the feedback control gain related to the torque control of the first electric motor M1 is changed after the inertia phase in the shifting process of the automatic transmission unit 20 is started. . For example, it is performed during the inertia phase after the start of the inertia phase.

On the other hand, by returning the feedback control gain changed to the second proportional gain KP ′ and the second integral gain KI ′ to the proportional gain KP and the integral gain KI, the first motor torque _TM1 changes (varies) in the feedback control. As a result, the actual transmission input torque _TIN is also changed. Therefore, when the feedback control gain is returned to the proportional gain KP and the integral gain KI during the shift of the automatic transmission unit 20, the actual transmission unit input torque _TIN is changed. Thus, variation gradient of the shifting portion input rotation speed N _AT varies with the variation of the actual gear portion input torque T _IN, which may lead to shift shock. Therefore, in this embodiment, in order to suppress the shift shock, the change of the feedback control gain related to the torque control of the first electric motor M1 is released after the shift of the automatic transmission unit 20 is completed.

More specifically, returning to FIG. 7, the traveling state determination unit 92 further determines whether an inertia phase in the shifting process of the automatic transmission unit 20 has been started. For example, after determining that the hydraulic pressure command value for executing the shift of the automatic transmission unit 20 is output by the stepped shift control unit 82, the traveling state determination unit 92 determines that the actual transmission unit input rotational speed _NAT is the inertia phase. based on whether the changes stored predetermined rotation speed variation .DELTA.N _aT or previously experimentally determined for determining the start, it is determined whether the inertia phase has started. Further, the traveling state determination unit 92 determines whether or not the shift of the automatic transmission unit 20 has been completed. For example, the running state determining means 92, the shifting completion of the actual gear portion input rotation speed N _AT and the automatic transmission portion 20 the rotational speed difference is the automatic shifting portion 20 with the synchronous rotational speed of the shifting portion input rotation speed N _AT after shift It is determined whether or not the shift of the automatic transmission unit 20 has been completed based on whether or not it is within a predetermined synchronous rotation determination speed that is experimentally obtained and stored in advance.

  When the traveling state determining unit 92 determines that the inertia phase in the shifting process of the automatic transmission unit 20 has started, the gain changing unit 94 does not overlap the shift of the differential unit 11 and the shifting of the automatic transmission unit 20. Compared to the time, a gain change command for changing the proportional gain KP and the integral gain KI in the equation (1) smaller is output to the hybrid control means 86. Further, the gain changing unit 94 changes the gain gain KP and the integral gain KI in the above equation (1) to be small when the traveling state determining unit 92 determines that the shift of the automatic transmission unit 20 has been completed. Release (end) the command.

FIG. 12 is a flowchart corresponding to the flowchart of FIG. 10, and steps other than step S <b> 40 are not shown because they are the same as FIG. 10. In FIG. 12, it is determined in S401A corresponding to the traveling state determination means 92 whether or not the inertia phase in the shifting process of the automatic transmission unit 20 has been started. If the determination in S401A is negative, the determination is repeated until the determination is positive. Then, in S402A corresponding to the gain changing means 94 when the determination in S401A is positive, as compared to when no overlap and shifting of the automatic shifting portion 20 of the differential portion 11, the engine rotational speed N _E The feedback control gain when the feedback control of the first motor torque T _M1 is executed is changed to be small so that becomes the target engine speed N _E ^* . For example, a gain change command for changing the proportional gain KP and the integral gain KI in the equation (1) smaller than when the shift of the differential unit 11 and the shift of the automatic transmission unit 20 do not overlap is output.

  Next, in S403A corresponding to the traveling state determination unit 92, it is determined whether or not the shift of the automatic transmission unit 20 has been completed. If the determination in S403A is negative, the steps after S402A are repeatedly executed until the determination is positive. If the determination in S403A is affirmative, in S404A corresponding to the gain changing means 94, the feedback control gain that has been changed slightly in S402A is returned to the original feedback control gain before the change. For example, the gain change command for changing the proportional gain KP and the integral gain KI in Equation (1) to be small is canceled.

As described above, according to the present embodiment, the feedback control gain change related to the torque control of the first electric motor M1 is performed after the start of the inertia phase (for example, during the inertia phase) in the shift process of the automatic transmission unit 20. For example, in comparison with changing the feedback control gain related to the torque control of the first electric motor M1 before the start of the inertia phase in the shift process, the fact that the first electric motor torque T _M1 is changed by the change of the feedback control gain. It is possible to suppress fluctuations in driving torque (driving force) accompanying fluctuations in the transmission unit input torque _TIN .

Further, according to the present embodiment, the change of the feedback control gain related to the torque control of the first electric motor M1 is canceled after the shift of the automatic transmission unit 20 is completed (that is, the original feedback control gain is restored). Compared to canceling the change of the feedback control gain related to the torque control of the first motor M1 during the shift of the unit 20, the first motor torque T _M1 is changed by releasing the change of the feedback control gain. variation gradient of the shifting portion input rotation speed N _AT with the variation of the transmission unit input torque T _iN is possible to avoid that varies. Therefore, it is possible to suppress the shift shock caused by the fluctuations of the change gradient of the transmission portion input rotation speed N _AT. Further, since the change of the feedback control gain is canceled after the shift is completed, not during the shift of the automatic transmission unit 20, the effect of changing the feedback control gain during the shift of the automatic transmission unit 20 (particularly during the inertia phase) can be sufficiently obtained. .

  In this embodiment, regarding the timing (time) when the feedback control gain in the equation (1) is changed to the second proportional gain KP ′ and the second integral gain KI ′ smaller than the proportional gain KP and the integral gain KI, An alternative embodiment to Example 2 is proposed.

By changing the feedback control gain from the proportional gain KP and the integral gain KI to the second proportional gain KP ′ and the second integral gain KI ′, the first motor torque T _M1 is changed (varied) in the feedback control and the actual speed change is performed. The part input torque _{TIN is} also changed. Therefore, changing the feedback control gain in the inertia phase of the automatic shifting portion 20 to the second proportional gain KP 'and the second integral gain KI', shifting portion input rotation speed with the variation of the actual gear portion input torque T _IN feedback control of the gradient of change N _aT could be temporarily reduced. Therefore, in the present embodiment, in order to suppress a decrease in feedback controllability of the change gradient of the transmission unit input rotation speed _NAT , the change of the feedback control gain related to the torque control of the first electric motor M1 is changed. Perform before starting the inertia phase in the process.

On the other hand, by returning the feedback control gain changed to the second proportional gain KP ′ and the second integral gain KI ′ to the proportional gain KP and the integral gain KI, the first motor torque _TM1 changes (varies) in the feedback control. As a result, the actual transmission input torque _TIN is also changed. Therefore, if the feedback control gain is returned to the proportional gain KP and the integral gain KI after the shift of the automatic transmission unit 20 is completed, a change in drive torque (drive force) occurs after the shift. Therefore, in the present embodiment, in order to suppress the change in the driving torque (driving force) after the shifting of the automatic transmission unit 20 is finished, the feedback control gain change related to the torque control of the first electric motor M1 is changed. Cancel at the end of shifting.

More specifically, returning to FIG. 7, the traveling state determination unit 92 further determines whether or not it is before the start of the inertia phase in the shift process of the automatic transmission unit 20. For example, after determining that the hydraulic pressure command value for executing the shift of the automatic transmission unit 20 is output by the stepped shift control unit 82, the traveling state determination unit 92 determines that the actual transmission unit input rotational speed _NAT is the inertia phase. based on whether a previously experimentally below sought predetermined rotation speed variation amount stored in .DELTA.N _aT for determining the start, it is determined whether prior to the start of the inertia phase. Further, the traveling state determination unit 92 determines whether or not the automatic transmission unit 20 is at the end of the shift. For example, the running state determining means 92, shifting the end of the actual gear portion input rotation speed N _AT and the automatic transmission portion 20 the rotational speed difference is the automatic shifting portion 20 with the synchronous rotational speed of the shifting portion input rotation speed N _AT after shift It is determined whether or not it is the end of shift of the automatic transmission unit 20 based on whether or not it is within a predetermined shift end determination speed that is experimentally obtained and stored in advance.

  The gain changing unit 94, instead of the second embodiment, determines that the inertia phase in the shifting process of the automatic transmission unit 20 is before the start by the running state determination unit 92, A gain change command for changing the proportional gain KP and the integral gain KI in the equation (1) to be smaller than that when the shift of the automatic transmission unit 20 does not overlap is output to the hybrid control means 86. Further, the gain changing means 94, instead of the second embodiment, when the traveling state determining means 92 determines that the shift end of the automatic transmission unit 20 is reached, the proportional gain KP in the equation (1) and Release (end) the gain change command to change the integral gain KI small.

FIG. 13 is a flowchart corresponding to the flowchart of FIG. 10, and steps other than step S <b> 40 are not shown because they are the same as FIG. 10. In FIG. 13, it is determined in S401B corresponding to the traveling state determination unit 92 whether the inertia phase in the shifting process of the automatic transmission unit 20 is before the start. If the determination in S401B is negative, this routine is terminated. If the determination is positive, in S402B corresponding to the gain changing means 94, the shift of the differential unit 11 and the shift of the automatic transmission unit 20 do not overlap. The feedback control gain when the feedback control of the first electric motor torque T _M1 is executed is changed to be smaller so that the engine rotational speed _NE becomes the target engine rotational speed _NE ^* . For example, a gain change command for changing the proportional gain KP and the integral gain KI in the equation (1) smaller than when the shift of the differential unit 11 and the shift of the automatic transmission unit 20 do not overlap is output.

  Next, in S403B corresponding to the traveling state determination unit 92, it is determined whether or not the automatic transmission unit 20 is shifting. For example, it is determined whether or not the shift end of the automatic transmission unit 20 is reached. If the determination in S403B is negative, the steps after S402B are repeatedly executed until the determination is positive. If the determination in S403B is affirmative, in S404B corresponding to the gain changing means 94, the feedback control gain that has been changed slightly in S402B is returned to the original feedback control gain before the change. For example, the gain change command for changing the proportional gain KP and the integral gain KI in Equation (1) to be small is canceled.

As described above, according to the present embodiment, the feedback control gain change related to the torque control of the first electric motor M1 is performed before the start of the inertia phase in the shifting process of the automatic transmission unit 20, so that, for example, after the start of the inertia phase ( for example, compared to changing the feedback control gain related to the torque control of the first electric motor M1 in the inertia phase), the actual gear portion input by the by changing the feedback control gain first-motor torque T _M1 is varied it is possible to prevent the feedback control of the change gradient of the transmission portion input rotation speed N _aT with the variation of the torque T _iN is temporarily reduced. That is, it is possible to improve the feedback control of the change gradient of the transmission portion input rotation speed N _AT.

Further, according to the present embodiment, the change of the feedback control gain related to the torque control of the first electric motor M1 is released at the end of the shift of the automatic transmission unit 20, so that, for example, after the shift of the automatic transmission unit 20 is completed, the first electric motor M1 compared to cancel the change of the feedback control gain related to torque control, due to the variation of the actual gear portion input torque T _iN by the first electric motor torque T _M1 by the release of change of the feedback control gain is varied Driving force fluctuations are prevented from occurring after a shift. In addition, since the change of the feedback control gain is canceled especially during the shift end of the automatic transmission unit 20, the effect of changing the feedback control gain during the shift of the automatic transmission unit 20 (particularly during the inertia phase) is sufficient. can get.

  As mentioned above, although the Example of this invention was described in detail based on drawing, this invention can be implemented combining an Example mutually and is applied also in another aspect.

  For example, in the above-described embodiment, the engine operating point is changed along the optimal fuel consumption rate curve along with the power-on downshift of the automatic transmission unit 20 and the power-on downshift in the flowchart of FIG. 10 and the time chart of FIG. Although the present invention has been described by exemplifying the simultaneous shift with the shift of the differential unit 11, the present invention is not limited to such a simultaneous shift and can be applied. For example, the present invention can also be applied to a simultaneous shift in which the shift of the differential unit 11 is performed by a so-called equal power shift with the shift of the automatic transmission unit 20.

In the above-described embodiment, when the feedback control gain is changed, the proportional second gain KP ′ and the second integral gain KI that are set to be smaller than those when the proportional gain KP and the integral gain KI are not simultaneously shifted are set. It has been changed to ', but it does not have to be a uniform predetermined gain. For example, when changing the feedback control gain, the feedback control gain is changed depending on the type of shift of the automatic transmission unit 20, such as changing the feedback control gain smaller as the change in the gear ratio γ _AT between the gears of the automatic transmission unit 20 is larger. The value of may be changed.

In the illustrated embodiment, the feedback control related to the shift control of the automatic shifting portion 20, the disengagement hydraulic pressure during the sweep such that the change gradient of the shifting portion input rotation speed N _AT in the inertia phase becomes a predetermined target slope Although the command value is corrected by feedback control, the present invention can be applied even if the engagement side hydraulic pressure command value at the time of sweep is corrected by feedback control.

  Further, in the above-described embodiment, by controlling the operating state of the first electric motor M1, the differential unit 11 (power distribution mechanism 16) continuously changes its speed ratio γ0 from the minimum value γ0min to the maximum value γ0max. For example, the gear ratio γ0 of the differential unit 11 may be changed stepwise using a differential action instead of continuously. Good.

  Further, in the power transmission device 10 of the above-described embodiment, the engine 8 and the differential unit 11 are directly connected, but the engine 8 may be connected to the differential unit 11 via an engagement element such as a clutch. Good.

  In the power transmission device 10 of the above-described embodiment, the first electric motor M1 and the second rotating element RE2 are directly connected, and the second electric motor M2 and the third rotating element RE3 are directly connected. The electric motor M1 may be connected to the second rotating element RE2 via an engaging element such as a clutch, and the second electric motor M2 may be connected to the third rotating element RE3 via an engaging element such as a clutch.

  In the above-described embodiment, the automatic transmission unit 20 is connected next to the differential unit 11 in the power transmission path from the engine 8 to the drive wheel 34, but the differential unit 11 next to the automatic transmission unit 20. May be in the order in which they are connected. In short, the automatic transmission unit 20 may be provided so as to constitute a part of the power transmission path from the engine 8 to the drive wheel 34, and the motor and the engine 8 may be connected so that power can be transmitted to the input side rotation member. .

  Further, according to FIG. 1 of the above-described embodiment, the differential unit 11 and the automatic transmission unit 20 are connected in series, but the electrical difference that can electrically change the differential state as the entire power transmission device 10. The present invention can be applied even if the differential unit 11 and the automatic transmission unit 20 are not mechanically independent as long as the function and the function of shifting by a principle different from the shift by the electric differential function are provided. Is done.

  In the above-described embodiment, the power distribution mechanism 16 is a single planetary, but may be a double planetary.

  The power distribution mechanism 16 serving 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). It may be a differential gear device operatively connected to the motor.

  In the above-described embodiment, the engine 8 is connected to the first rotating element RE1 constituting the differential planetary gear unit 24 so that power can be transmitted, and the first motor M1 is transmitted to the second rotating element RE2. The power transmission path to the drive wheel 34 is connected to the third rotating element RE3, but, for example, two or more planetary gear devices are connected to each other by some rotating elements constituting the third rotating element RE3. The engine, the electric motor, and the driving wheel are connected to the rotating element of the planetary gear device so as to be able to transmit power, and the stepped transmission is controlled by the clutch or brake connected to the rotating element of the planetary gear device. The present invention is also applied to a configuration that can be switched to a continuously variable transmission.

  In the above-described embodiment, the automatic transmission unit 20 is inserted in the power transmission path between the transmission member 18 that is the output member of the differential unit 11, that is, the power distribution mechanism 16, and the drive wheel 34. For example, a continuously variable transmission (CVT), which is a kind of automatic transmission, and a continuously meshing parallel two-shaft type well known as a manual transmission, the gear stage can be automatically switched by a select cylinder and a shift cylinder. Other types of power transmission devices (transmissions) such as possible automatic transmissions may be provided. In the case of the continuously variable transmission (CVT), for example, a plurality of fixed gear ratios are stored in advance so as to correspond to the gear stages in the stepped transmission, and the gears are shifted using the plurality of fixed gear ratios. May be executed.

  Further, in the above-described embodiment, the second electric motor M2 is directly connected to the transmission member 18, but the connecting position of the second electric motor M2 is not limited to this, but directly or a transmission, a planetary gear device, an engagement It may be indirectly connected through a device or the like.

  In the power distribution mechanism 16 of the above-described embodiment, the differential carrier CA0 is connected to the engine 8, the differential sun gear S0 is connected to the first electric motor M1, and the differential ring gear R0 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 included in the three elements CA0, S0, and R0 of the differential planetary gear unit 24. It may be connected to any of these.

  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, for example, via a gear, a belt, or the like, and needs to be disposed on a common axis. Absent.

  In the above-described embodiment, the first motor M1 and the second motor M2 are disposed concentrically with the input shaft 14, the first motor M1 is connected to the differential sun gear S0, and the second motor M2 is connected to the transmission member 18. The first motor M1 is operatively connected to the differential unit sun gear S0, for example, via a gear, a belt, a speed reducer, and the like, and is not necessarily arranged as such. May be coupled to the transmission member 18.

  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 of the above-described embodiment is composed of a pair of differential planetary gear devices 24, but is composed of two or more planetary gear devices in a non-differential state (constant shift state). It may function as a transmission having three or more stages.

  In addition, the second electric motor M2 of the above-described embodiment is connected to the transmission member 18 constituting a part of the power transmission path from the engine 8 to the drive wheel 34, but the second electric motor M2 is connected to the power transmission path. In addition, the power distribution mechanism 16 can be connected via an engagement element such as a clutch, and the differential state of the power distribution mechanism 16 is changed by the second electric motor M2 instead of the first electric motor M1. The power transmission device 10 may be configured to be controllable.

  In the above-described embodiment, the differential unit 11 includes the first electric motor M1 and the second electric motor M2, but the first electric motor M1 and the second electric motor M2 transmit power separately from the differential unit 11, respectively. The apparatus 10 may be provided.

  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 above-described embodiment, the hydraulic friction engagement device such as the first clutch C1 and the second clutch C2 is a magnetic type such as a powder (magnetic powder) clutch, an electromagnetic clutch, an engagement type dog clutch, an electromagnetic type, You may be comprised from the mechanical engagement apparatus. For example, in the case of an electromagnetic clutch, the hydraulic control circuit 70 is constituted by a switching device, an electromagnetic switching device, or the like that switches an electrical command signal circuit to the electromagnetic clutch, not a valve device that switches an oil passage.

  Each of the above-described embodiments can be implemented in combination with each other, for example, by providing a priority order.

  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.

8: Engine 10: Vehicle power transmission device 11: Differential part (electrical differential part)
16: Power distribution mechanism (differential mechanism)
18: Transmission member (output-side rotating member of electric differential section)
20: Automatic transmission unit (transmission unit)
34: Drive wheel 80: Electronic control device (control device)
B: Brake (engagement device)
C: Clutch (engagement device)
M1: First motor (differential motor)
M2: Second electric motor (traveling motor)

Claims (7)

A differential mechanism coupled to the engine so as to be capable of transmitting power; and a differential motor coupled to the differential mechanism so as to be capable of transmitting power. The differential motor is controlled by controlling an operating state of the differential motor. An electric differential unit in which the differential state of the mechanism is controlled, a traveling motor connected to a drive wheel so as to be able to transmit power, and an output side rotating member of the electric differential unit connected in series A control device for a vehicle power transmission device, comprising: a transmission unit that constitutes a part of a power transmission path from an electric differential unit to the drive wheel;
Torque control of the differential motor is executed by feedback control so that the engine rotation speed becomes the target engine rotation speed, and separately from the feedback control in the torque control, the input rotation of the transmission unit in the shifting process of the transmission unit The speed change control of the speed change unit is executed by feedback control so that the speed change gradient becomes the target slope,
When the shift of the electric differential unit and the shift of the transmission unit overlap, the feedback control gain related to torque control of the differential motor during the shift is changed smaller than when the shift does not overlap A control device for a vehicle power transmission device.   The control device for a vehicle power transmission device according to claim 1, wherein the change of the feedback control gain related to the torque control is performed after the start of the inertia phase in the shifting process of the transmission unit.   2. The control device for a vehicle power transmission device according to claim 1, wherein the feedback control gain related to the torque control is changed before the inertia phase is started in the shifting process of the transmission unit.   The control device for a vehicle power transmission device according to any one of claims 1 to 3, wherein the change of the feedback control gain related to the torque control is canceled after the shift of the transmission unit is completed.   4. The control device for a vehicle power transmission device according to claim 1, wherein the change of the feedback control gain related to the torque control is canceled at the end of the shift of the transmission unit. 5. The transmission unit is a stepped automatic transmission in which a clutch-to-clutch shift is performed by releasing the disengagement side engagement device and engaging the engagement side engagement device to selectively establish a plurality of gear stages. Yes,
The feedback control related to the shift control of the transmission unit is feedback control of the engagement pressure control of at least one of the engagement devices so that the change gradient of the input rotation speed of the transmission unit during the shift process of the transmission unit is set as a target gradient. The control device for a vehicle power transmission device according to any one of claims 1 to 5, wherein the control device is executed by the following. The shift of the transmission unit is a power-on downshift associated with an increase in the amount of output required for the vehicle,
7. The vehicle according to claim 1, wherein the shift of the electric differential unit changes the engine operating point along a preset optimum fuel consumption line. Control device for power transmission device.

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