JP2005313786A - Control device for vehicular stepped automatic transmission - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a control device for a vehicular stepped automatic transmission capable of suitably restraining a shifting shock when shifting is completed in a vehicle provided with the stepped automatic transmission having an input shaft mechanically connected to an engine. <P>SOLUTION: As a direct-coupled clutch control means 120 is provided for slipping or releasing a direct-coupled clutch Ci at the time of completing the shifting of the automatic transmission 10 in the vehicle, torque vibration generated by a sudden stop of an engine rotation speed change at the time of completing the shifting is absorbed by slipping or releasing of the direct-coupled clutch Ci, which suitably restrains the shifting shock caused by the torque vibration at the time of completing the shifting. <P>COPYRIGHT: (C)2006,JPO&NCIPI

Description

  The present invention relates to a control device for a stepped automatic transmission for a vehicle, and particularly occurs when a shift is performed in a state where an input shaft of the stepped automatic transmission is mechanically coupled to an engine, immediately after the shift. The present invention relates to a technique for suppressing torque vibration.

  There is a vehicle including a stepped automatic transmission that has an input shaft that is mechanically connected to an engine and can be selectively switched to one of a plurality of gear stages. For example, this is the vehicle described in Patent Document 1. In such a vehicle, for example, a fluid transmission device having a large torque vibration absorption function such as a torque converter is not provided between the engine and the stepped automatic transmission. An electric motor is provided for each of the input-side rotating member and the output-side rotating member of the direct coupling clutch, and motor traveling is enabled by releasing the direct coupling clutch, and engine traveling or engine and motor traveling is enabled by engaging the direct coupling clutch. The

US Patent Publication No. 2003 / 0127262A1 JP-A-9-291838 JP-A-9-308008 JP-A-7-331603 Japanese Patent Laid-Open No. 10-24745

  By the way, according to the vehicle as described above, since the input shaft of the stepped automatic transmission and the engine are mechanically coupled, the gear ratio is relatively short when the stepped automatic transmission is shifted. Since the engine speed is changed step by step within the operation period, the engine speed is rapidly changed toward the synchronous speed corresponding to the speed ratio after the shift, and when the engine speed matches the synchronous speed, the speed Because the change is suddenly stopped, an elastic torsional vibration is generated in the power transmission path from the engine to the drive wheel due to the sudden stop of the engine speed change at the end of the shift, and this is due to the elastic torsional vibration. There was a possibility that a shift shock would occur. The same problem occurs when the stepped automatic transmission is shifted while the lockup clutch of the torque converter is engaged.

  The present invention has been made in the background of the above circumstances, and the object of the present invention is to provide an input shaft that is mechanically connected to the engine via a direct coupling clutch and has any one of a plurality of gear stages. Another object of the present invention is to provide a control device for a stepped automatic transmission for a vehicle that suitably suppresses a shift shock at the end of the shift in a vehicle having a stepped automatic transmission that can be switched alternatively.

  The gist of the invention according to claim 1 for achieving this object is a stepped automatic transmission, and a direct coupling clutch interposed between an engine and an input shaft of the stepped automatic transmission. A stepped automatic transmission control device for a vehicle that performs a shift in a state where the input shaft of the stepped automatic transmission is mechanically coupled to the engine via the direct coupling clutch, It is intended to include direct coupling clutch control means for slipping or releasing the direct coupling clutch at the end of shifting of the stepped automatic transmission.

  The gist of the invention according to claim 2 is that in the invention according to claim 1, direct connection clutch operation availability determination means for determining whether or not the direct connection clutch can be slipped or released at the end of the shift. And the torque vibration generated in the power transmission path of the vehicle at the end of shifting of the stepped automatic transmission when it is determined by the direct clutch activation availability determination means that the slip or release of the direct clutch is not possible. And a suppression torque control means for applying a suppression torque for suppressing the power to the power transmission path.

  A gist of the invention according to claim 3 is that, in the invention according to claim 2, the motor includes an electric motor operatively connected to an input shaft of the stepped automatic transmission, and the suppression torque control means includes The electric motor outputs a torque for suppressing the torque vibration.

  Further, the gist of the invention according to claim 4 is that, in the invention according to claim 3, the suppression torque control means has a phase opposite to that of the torque vibration as a suppression torque for suppressing the torque vibration. It includes an anti-phase torque output control means for outputting phase torque vibration from the electric motor.

  A gist of the invention according to claim 5 is that, in the invention according to any one of claims 1 to 4, the output torque of the engine is temporarily reduced at the end of shifting of the stepped automatic transmission. It further includes engine output lowering means.

  The gist of the invention according to claim 6 is that, in the invention according to claim 4, the antiphase torque output control means depends on the amount of inertia torque reduction generated at the end of the shift. It is characterized by changing the magnitude of vibration.

  The gist of the invention according to claim 7 is that, in the invention according to claim 4, the antiphase torque output control means changes the magnitude of the antiphase torque vibration depending on the type of the shift. It is a thing to do.

  Further, the gist of the invention according to claim 8 is that, in the invention according to any one of claims 1 to 7, the direct coupling clutch control means depends on an inertia torque reduction amount generated at the end of the shift. The slip amount of the direct clutch is changed.

  A gist of the invention according to claim 9 is that, in the invention according to any one of claims 1 to 7, the direct clutch control means determines the slip amount of the direct clutch depending on the type of the shift. It is a thing to change.

  The invention according to claim 1 includes a stepped automatic transmission and a direct coupling clutch interposed between the engine and an input shaft of the stepped automatic transmission. A control device for a stepped automatic transmission for a vehicle that executes a shift in a state where an input shaft is mechanically coupled to an engine via a direct coupling clutch, and at the end of the shift of the stepped automatic transmission Since direct coupling clutch control means for slipping or releasing the direct coupling clutch is included, torque vibration generated by sudden stop of engine rotation speed change at the end of shifting is absorbed by slipping or releasing of the direct coupling clutch. The shift shock due to the torque vibration is suitably suppressed.

  In the invention according to claim 2, the direct coupling clutch operation availability determination means for determining whether slip or release control by the direct coupling clutch is possible at the end of the shift, and the direct coupling clutch operation availability determination means. When it is determined that the clutch cannot be slipped or released, a suppression torque for suppressing torque vibration generated in the power transmission path of the vehicle at the end of shifting of the stepped automatic transmission is applied to the power transmission path. In addition, when the torque vibration generated by the sudden stop of the engine speed change at the end of the shift is not absorbed by the slip or release of the direct clutch, the power is replaced by The torque vibration is suppressed by the suppression torque for suppressing the torque vibration applied to the transmission path. Since the, shift shock due to torque fluctuation in the gear shifting completion it is suitably suppressed.

  The invention according to claim 3 further includes an electric motor operatively connected to the input shaft of the stepped automatic transmission, and the suppression torque control means suppresses the torque vibration from the electric motor. Since the torque is output, the torque vibration is suppressed by the suppression torque applied from the electric motor to the power transmission path, so that the shift shock due to the torque vibration at the end of the shift is preferably suppressed.

  Further, in the invention according to claim 4, the suppression torque control means outputs an antiphase torque vibration having an antiphase torque vibration opposite in phase to the torque vibration from the electric motor as the suppression torque for suppressing the torque vibration. Since the control means is included, the torque vibration is canceled by the antiphase torque vibration applied from the electric motor to the power transmission path, so that the shift shock due to the torque vibration at the end of the shift is suitably suppressed.

  The invention according to claim 5 further includes engine output reduction means for temporarily reducing the output torque of the engine at the end of shifting of the stepped automatic transmission. Therefore, the slip control amount of the direct coupling clutch for absorbing it and the antiphase torque vibration for canceling it are reduced.

  Further, in the invention according to claim 6, the reverse phase torque output control means changes the magnitude of the reverse phase torque vibration depending on the amount of inertia torque reduction generated at the end of the shift. Torque vibration generated at the end of shifting is preferably canceled out by an antiphase torque vibration of an appropriate magnitude.

  Further, in the invention according to claim 7, since the anti-phase torque output control means changes the magnitude of the anti-phase torque vibration depending on the type of the shift, The torque vibration generated at the end of the shift is preferably canceled by the phase torque vibration.

  Further, in the invention according to claim 8, the direct coupling clutch control means changes the slip amount of the direct coupling clutch depending on the amount of inertia torque reduction generated at the end of the shift. The torque vibration generated at the end of the shift is suitably absorbed by the slip amount.

  Further, in the invention according to claim 9, since the direct clutch control means changes the slip amount of the direct clutch depending on the type of the shift, the shift is completed with an appropriate amount of slip. Torque vibrations sometimes generated are favorably absorbed.

  Here, the stepped automatic transmission for a vehicle is a planetary gear type multi-stage transmission in which gear stages are switched by selectively connecting rotating elements of a plurality of planetary gear apparatuses by a hydraulic friction engagement device. A synchronous mesh parallel twin-shaft system in which a plurality of pairs of transmission gears that are always meshed are provided between two shafts, and one of the plurality of transmission gears is selectively transmitted to a power by a synchronization device driven by a hydraulic actuator. It consists of an automatic transmission. The mounting posture of the above-mentioned stepped automatic transmission for a vehicle may be a horizontal installation type such as an FF (front engine / front drive) vehicle in which the transmission axis is in the width direction of the vehicle. It may be a vertical installation type such as an FR (front engine / rear drive) vehicle in the front-rear direction.

  The stepped automatic transmission only needs to be able to achieve a plurality of gears alternatively. For example, the stepped automatic transmission has four forward speeds, five forward speeds, six forward speeds, seven forward speeds, eight forward speeds, etc. Various multi-stage automatic transmissions can be used.

  Between the engine and the input shaft of the automatic transmission, a direct coupling clutch, a damper with a direct coupling clutch, or a lockup clutch (direct coupling clutch) with a torque converter may be interposed. In short, any vehicle may be used as long as the stepped automatic transmission is shifted in a state where the input shaft of the stepped automatic transmission is mechanically coupled to the engine.

  Preferably, a direct coupling clutch is provided between the engine and the input shaft of the automatic transmission. In this case, a first electric motor and a second electric motor are provided on the input side rotating member of the direct coupling clutch and the output side rotating member of the direct coupling clutch, respectively, and the phases of the first electric motor and / or the second electric motor are opposite to each other. Torque vibration is applied. The second electric motor may be provided on the output shaft of the stepped automatic transmission.

  FIG. 1 is a skeleton diagram illustrating a configuration of a stepped automatic transmission (hereinafter referred to as an automatic transmission) 10 provided in a vehicle to which the present invention is applied. The automatic transmission 10 mainly includes an input shaft 16 and a first planetary gear unit 18 connected to a crankshaft 9 of the engine 8 via a direct coupling clutch Ci on a common shaft center in a transmission case 12 attached to the vehicle body. The first transmission unit 20 configured as described above, the second transmission unit 26 configured mainly by the second planetary gear device 22 and the third planetary gear device 24, and the output shaft 28 are sequentially arranged. The input shaft 16 functions as an output side rotation member of the direct coupling clutch Ci, and also functions as an input rotation member of the automatic transmission 10. The output shaft 28 corresponds to an output rotating member of the automatic transmission 10, and the transmission case 12 corresponds to a non-rotating member. The output shaft 28 rotationally drives the left and right drive wheels via, for example, a differential gear device (not shown). Since the automatic transmission 10 is configured symmetrically with respect to its axis, the lower side is omitted in the skeleton diagram of FIG.

  The first planetary gear unit 18 is a double pinion type planetary gear unit, and includes a sun gear S1, a plurality of pairs of pinion gears P1 that mesh with each other, a carrier CA1 that supports the pinion gears P1 so as to rotate and revolve, and a sun gear S1 via the pinion gears P1. A ring gear R1 meshing with the ring gear R1. The carrier CA1 is coupled to the input shaft 16 and driven to rotate, and the sun gear S1 is fixed to the transmission case 12 so as not to rotate. The ring gear R <b> 1 functions as an intermediate output member, is rotated at a reduced speed with respect to the input shaft 16, and transmits the rotation to the second transmission unit 26. In the present embodiment, the path for transmitting the rotation of the input shaft 16 to the second transmission unit 26 at the same speed is the first intermediate output path for transmitting the rotation at a predetermined constant speed ratio (= 1.0). The first intermediate output path PA1 includes a direct connection path PA1a that transmits rotation from the input shaft 16 to the second transmission unit 26 without passing through the first planetary gear unit 18, and a first planetary gear from the input shaft 16. There is an indirect path PA1b that transmits the rotation to the second transmission unit 26 via the carrier CA1 of the device 18. Further, the transmission ratio from the input shaft 16 to the second transmission 26 via the carrier CA1, the pinion gear P1 disposed on the carrier CA1, and the ring gear R1 is larger than the first intermediate output path PA1 (> 1). .0) is a second intermediate output path PA2 that transmits the rotation of the input shaft 16 at a reduced speed (deceleration).

  The second planetary gear unit 22 is a single-pinion type planetary gear unit, and includes a sun gear S2, a carrier CA2 that supports the pinion gear P2 so as to rotate and revolve, and a ring gear R2 that meshes with the sun gear S2 via the pinion gear P2. The third planetary gear unit 24 is a double-pinion type planetary gear unit, and includes a sun gear S3, a plurality of pairs of pinion gears P2 and P3 that mesh with each other, a carrier CA3 that supports the pinion gears P2 and P3 so as to rotate and revolve, and pinion gears P2 and P3. Is provided with a ring gear R3 that meshes with the sun gear S3.

  In the second planetary gear device 22 and the third planetary gear device 24, the pinion gear P2, the carriers CA2 and CA3 that rotatably support the pinion gear P2, and the ring gears R2 and R3 are shared with each other, thereby four rotation elements RM1 to RM4. Is configured. That is, the first rotating element RM1 is configured by the sun gear S2 of the second planetary gear unit 22, and the carrier CA2 of the second planetary gear unit 22 and the carrier CA3 of the third planetary gear unit 22 are integrally connected to each other to perform the second rotation. The element RM2 is configured, and the third rotating element RM3 is configured by the ring gear R2 of the second planetary gear unit 22 and the ring gear R3 of the third planetary gear unit 24, and the sun gear S3 of the third planetary gear unit 24 is coupled to each other. A four-rotation element RM4 is configured.

  The first rotating element RM1 (sun gear S2) is selectively connected to the transmission case 12 via the first brake B1 and stopped, and the first planetary gear unit 18 which is an intermediate output member via the third clutch C3. Is selectively connected to the ring gear R1 (ie, the second intermediate output path PA2), and is further selectively connected to the input shaft 16 (ie, the direct connection path PA1a of the first intermediate output path PA1) via the fourth clutch C4. ing. The second rotation element RM2 (carriers CA2 and CA3) is selectively connected to the transmission case 12 via the second brake B2 and stopped from rotation, and is selectively connected to the input shaft 16 via the second clutch C2. It is connected. The third rotation element RM3 (ring gears R2 and R3) is integrally connected to the output gear 28 to output rotation. The fourth rotation element RM4 (sun gear S3) is connected to the ring gear R1 via the first clutch C1. Each of the brakes B1 and B2 and the clutches C1 to C4 is a multi-plate hydraulic friction engagement device that is frictionally engaged by a hydraulic cylinder.

  FIG. 2 is a collinear diagram in which the rotational speeds of the rotary elements of the first transmission unit 20 and the second transmission unit 26 can be represented by straight lines. The lower horizontal line indicates the rotational speed “0”. The horizontal line indicates the rotational speed “1.0”, that is, the same rotational speed as the input shaft 16. Further, each vertical line of the first transmission unit 20 represents the sun gear S1, the ring gear R1, and the carrier CA1 in order from the left side, and these intervals are the gear ratio ρ1 (= sun gear S1 of the first planetary gear unit 18). The number of teeth / the number of teeth of the ring gear R1). FIG. 2 shows a case where the gear ratio ρ1 = 0.463, for example. The four vertical lines of the second transmission unit 26 indicate, in order from the left side, the first rotating element RM1 (sun gear S2), the second rotating element RM2 (carrier CA2 and carrier CA3), and the third rotating element RM3 (ring gear R2 and ring gear). R3), the fourth rotation element RM4 (sun gear S3), and their intervals are determined according to the gear ratio ρ2 of the second planetary gear unit 22 and the gear ratio ρ3 of the third planetary gear unit 24. FIG. 2 shows a case where the gear ratio ρ2 = 0.463 and ρ3 = 0.415, for example.

  As is clear from this nomograph, the first clutch C1 and the second brake B2 are engaged, and the fourth rotating element RM4 rotates at a reduced speed with respect to the input shaft 16 via the first transmission 20. When the rotation of the second rotation element RM2 is stopped, the third rotation element RM3 connected to the output gear 28 is rotated at the rotation speed indicated by “1st”, and the largest gear ratio (= input shaft 16). The first shift speed “1st” is established. The first clutch C1 and the first brake B1 are engaged, and the fourth rotating element RM4 is decelerated and rotated with respect to the input shaft 16 via the first transmission unit 20, and the first rotating element RM1 stops rotating. Then, the third rotation element RM3 is rotated at the rotation speed indicated by “2nd”, and the second speed “2nd” having a smaller gear ratio than the first speed “1st” is established. The first clutch C1 and the third clutch C3 are engaged, and the fourth rotation element RM4 and the first rotation element RM1 are decelerated and rotated with respect to the input shaft 16 via the first transmission unit 20 to perform the second shift. When the part 26 is rotated integrally, the third rotation element RM3 is rotated at the rotation speed indicated by “3rd”, and the third shift stage “3rd” having a smaller speed ratio than the second shift stage “2nd” is established. It is done. The first clutch C1 and the fourth clutch C4 are engaged, the fourth rotating element RM4 is decelerated and rotated with respect to the input shaft 16 via the first transmission unit 20, and the first rotating element RM1 is input to the input shaft. When it is rotated integrally with 16, the third rotation element RM3 is rotated at the rotational speed indicated by “4th”, and the fourth shift stage “4th” having a smaller speed ratio than the third shift stage “3rd” is established. . When the first clutch C1 and the second clutch C2 are engaged, the fourth rotation element RM4 is rotated at a reduced speed with respect to the input shaft 16 via the first transmission unit 20, and the second rotation element RM2 is input to the input shaft 16. The third rotation element RM3 is rotated at the rotation speed indicated by “5th”, and the fifth shift stage “5th” having a smaller gear ratio than the fourth shift stage “4th” is established. When the second clutch C2 and the fourth clutch C4 are engaged and the second transmission unit 26 is rotated integrally with the input shaft 16, the third rotational element RM3 is rotated at the rotational speed indicated by "6th", that is, with the input shaft 16. The sixth shift stage “6th”, which is rotated at the same rotational speed and has a smaller speed ratio than the fifth shift stage “5th”, is established. The gear ratio of the sixth gear stage “6th” is 1. The second clutch C2 and the third clutch C3 are engaged, and the first rotating element RM1 is decelerated and rotated with respect to the input shaft 16 via the first transmission unit 20, and the second rotating element RM2 is input to the input shaft. When it is rotated integrally with 16, the third rotation element RM3 is rotated at the rotational speed indicated by “7th”, and the seventh shift stage “7th” having a smaller gear ratio than the sixth shift stage “6th” is established. . When the second clutch C2 and the first brake B1 are engaged, the second rotating element RM2 is rotated integrally with the input shaft 16, and the first rotating element RM1 is stopped rotating, the third rotating element RM3 is “ The eighth speed “8th” is established with a lower speed ratio than the seventh speed “7th”.

  When the third clutch C3 and the second brake B2 are engaged, the first rotating element RM1 is rotated at a reduced speed via the first transmission unit 20, and the second rotating element RM2 is stopped from rotating. The third rotation element RM3 is reversely rotated at the rotation speed indicated by “Rev1”, and the first reverse shift stage “Rev1” having the largest speed ratio in the reverse rotation direction is established. When the fourth clutch C4 and the second brake B2 are engaged, the first rotation element RM1 is rotated integrally with the input shaft 16, the second rotation element RM2 is stopped, and the third rotation element RM3 is “ The second reverse shift speed “Rev2”, which is reversely rotated at the rotation speed indicated by “Rev2” and has a smaller gear ratio than the first reverse shift speed “Rev1”, is established. The first reverse speed “Rev1” and the second reverse speed “Rev2” correspond to the first speed and the second speed in the reverse rotation direction, respectively.

  FIG. 3 is an operation table for explaining the engagement elements and the gear ratios when the above gear positions are established. “◯” indicates the engaged state, and the blank is released. The gear ratio of each gear stage is appropriately determined by the gear ratios ρ1 to ρ3 of the first planetary gear device 18, the second planetary gear device 22, and the third planetary gear device 24, for example, ρ1 = 0.463, ρ2 = 0. .463, ρ3 = 0.415, the value of the gear ratio step (the gear ratio between the gears) is substantially appropriate and the total gear ratio width (= 4.532 / 0.667) is also obtained. The gear ratio of the reverse gears “Rev1” and “Rev2” is also appropriate, and an appropriate gear ratio characteristic is obtained as a whole.

  As described above, the automatic transmission 10 according to the present embodiment includes the first transmission unit 20 having the two intermediate output paths PA1 and PA2 having different transmission ratios and the second transmission unit 26 having the two planetary gear units 22 and 24. Since the forward shift 8-speed gear stage is achieved by switching the engagement of the four clutches C1 to C4 and the two brakes B1 and B2, the structure is reduced in size and the mountability to the vehicle is improved. Further, as shown in FIG. 3, the automatic transmission 10 according to the present embodiment can have a large speed ratio width and an appropriate speed ratio step. In addition, as is apparent from FIG. 3, since it is possible to perform shifts at each shift stage by simply grasping any one of the clutches C1 to C4 and the brakes B1 and B2, shift control is easy and shift shock is generated. Is suppressed.

FIG. 4 is a block diagram illustrating a control system for controlling the engine 8, the direct coupling clutch Ci, the gear stage of the automatic transmission 10, the motor generators MG1 and MG2, and the like in the vehicle of the present embodiment. In FIG. 4, the operation amount Acc of the accelerator pedal 50 is detected by an accelerator operation amount sensor 51. The accelerator pedal 50 is largely depressed according to the driver's required output amount, and therefore corresponds to an accelerator operation member, and the accelerator operation amount Acc corresponds to an output request amount. The intake pipe of the engine 8 is provided with an electronic throttle valve 56 that has an opening angle (opening) θ _TH corresponding to the accelerator operation amount Acc by a throttle actuator 54. Further, in the bypass passage 52 that bypasses the electronic throttle valve 56 for idle rotation speed control, the intake air amount when the electronic throttle valve 56 is fully closed is controlled in order to control the idle rotation speed NE _IDL of the engine 8. An ISC (idle rotational speed control) valve 53 is provided. In addition, an engine rotation speed sensor 58 for detecting the rotation speed NE of the engine 8 (= the rotation speed of the first motor generator MG1), an intake air amount sensor 60 for detecting the intake air amount Q of the engine 8, for detecting the fully closed state of the electronic throttle valve 56 (idle state) and the idle throttle valve with switch opening sensor 62 for detecting the opening degree theta _TH, (corresponding to the rotational speed Nout of the output shaft 28) speed V Vehicle speed sensor 64, input shaft rotational speed sensor 66 for detecting rotational speed Nin of input shaft 16 of automatic transmission 10 (= rotational speed of second motor generator MG2), presence / absence of operation of foot brake which is a service brake a brake switch 68 for detecting a lever position (operating position) of the shift lever 72 Rebapo for detecting P _SH Sensor 74, a coolant temperature sensor 76 for detecting the coolant temperature of the engine 8, an oil temperature sensor 80 for detecting the operating oil temperature of the automatic transmission 10, and a temperature of a catalyst for purifying exhaust gas. A catalyst temperature sensor 82, an acceleration sensor 84 for detecting vehicle acceleration, and the like are provided. The engine speed NE, the intake air amount Q, the throttle valve opening θ _TH , the vehicle speed V are detected from these sensors and switches. , Input shaft rotation speed Nin, presence / absence of brake operation, lever position P _SH of the shift lever 72, coolant temperature Iw, oil temperature Toil, catalyst temperature Tre, vehicle acceleration G, and the like are supplied to the electronic control unit 90. It is like that.

  The electronic control unit 90 includes a so-called microcomputer having a CPU, a RAM, a ROM, an input / output interface, and the like. The CPU uses a temporary storage function of the RAM, and signals according to a program stored in the ROM in advance. By performing the process, the engine 8 output control, the automatic transmission 10 shift control, the shift end shock mitigation control, the power control and regeneration control of the first motor generator MG1 and the second motor generator MG2, etc. It is configured to be divided into an engine control unit, a shift control unit, a hybrid control unit, and the like as necessary.

In the output control of the engine 8 by the electronic control device 90, the throttle actuator 54 controls the opening and closing of the electronic throttle valve 56, the fuel injection device 92 is controlled for the fuel injection amount control, and the igniter is used for the ignition timing control. The ISC valve 53 is controlled for idle rotation speed control. Control of the electronic throttle valve 56, for example, drives the throttle actuator 54 based on the actual accelerator operation amount Acc from the relationship shown in FIG. 5, the accelerator operation amount Acc increases the throttle valve opening theta _TH enough to increase. When the engine 8 is started, for example, the crankshaft 9 of the engine 8 is rotationally driven (cranked) by operating the first motor generator MG196 as an electric motor with the direct clutch Ci released.

In the shift control of the automatic transmission 10 by the electronic control unit 90, for example, a shift determination is made based on the actual vehicle speed V and the throttle valve opening θ _TH from the relationship stored in advance as shown in FIG. Shift control of the automatic transmission 16 is performed by switching the hydraulic circuit by exciting and de-energizing the AT shift solenoid in the hydraulic control circuit 99 for shifting so as to obtain a shift. This shift is normally performed at least during power-on running with the direct clutch Ci being engaged by the direct clutch control valve 96 provided in the hydraulic control circuit 98. The AT shift solenoid switches the engagement and disengagement states of the clutches C0 to C2 and the brakes B0 to B4 by changing the operation state of the switching valve and the like and switching the hydraulic control circuit 98, thereby changing the plurality of shift stages And neutral “N” are established, and a plurality of them are provided. The electronic control unit 90 uses the AT shift solenoid to provide shift control means for performing shift control according to a shift map (shift condition) stored in advance using the vehicle speed V and the throttle valve opening θ _TH illustrated in FIG. 6 as parameters. It is functionally provided so that a low-speed gear stage with a large gear ratio is established as the vehicle speed V decreases or the throttle valve opening _θTH increases. Further, the transitional hydraulic pressure of the clutches C0 to C2 and the brakes B0 to B4 that are engaged or released at the time of the shift is controlled by an AT line pressure control solenoid, a linear solenoid valve SLN, or the like, and a smooth shift is executed.

  As shown in FIG. 7, the shift lever 72 releases the power transmission path in the automatic transmission 10 and locks the rotation of the output shaft 28, and reversely rotates the output shaft 28 of the automatic transmission 10. Automatic shift within the R position for releasing the power transmission path in the automatic transmission 10, the N position for releasing the power transmission path in the automatic transmission 10, and the shift range (D range) that allows the first to eighth shifts of the automatic transmission 10. D position for executing control, M position for switching to manual shift mode, “+” position for shifting the shift range or gear up for each operation in manual shift mode, and for each operation in manual shift mode The shift position or gear stage is operated to the “−” position for shifting down, and the lever position sensor 74 is shifted to the shift position. Detecting the operating position of the lever 72. In the shift control of the automatic transmission 10 by the electronic control unit 90, the shift range or gear stage of the automatic transmission 10 is changed in response to a manual operation in the manual shift mode.

  By the way, within the shift period, the rotation speed of the engine 8 is changed at a predetermined change speed from the rotation speed before the shift to the rotation speed after the shift, and then synchronized with the rotation speed after the shift. When the rotational speed change of the engine 8 is forced to be maintained at the rotational speed after the shift and the rotational speed change is stopped, an elastic torsional vibration is generated in the power transmission path, which causes a shift shock. Such torsional vibration is not noticeable in a power transmission system in which a fluid power transmission device such as a torque converter is interposed, but the crankshaft 9 and the input shaft 16 are connected via a direct coupling clutch Ci as shown in FIG. This occurs remarkably in a power transmission system in which the crankshaft 9 and the input shaft 16 are mechanically connected as in the case of direct connection. In the end-of-shift shock mitigation control of the automatic transmission 10 by the electronic control unit 90, in order to mitigate such torsional vibration, torque vibration having a phase opposite to that of the torsional vibration is transmitted from the motor generator MG1 and / or the motor generator MG2. When the control is applied, if necessary, the engine output reduction control for reducing the engine output by retarding the ignition timing during the shift transition period, and the direct connection clutch Ci for releasing or sliding the direct connection clutch Ci at the shift end timing. Control is executed.

  Further, in the hybrid control by the electronic control unit 90, in order to perform motor travel, engine travel, motor and engine travel, regenerative braking travel, and the like according to the travel state of the vehicle, the opening / closing control of the direct clutch Ci, the first motor Power running control, regenerative control, and the like of the generator MG1 and the second motor generator MG2 are executed. For example, in the motor travel control, the MG2 controller 102 drives the inverter 106 from the inverter 106 in a state in which the valve direct coupling clutch Ci is released by the direct coupling clutch control 100 provided in the hydraulic control circuit 98 for quiet vehicle start-up and traveling. A current is supplied to the second motor generator MG2 to drive it. In the engine running control, the running of the engine 8 is performed even when the remaining amount of charge of the power storage device 108 is reduced. Therefore, the direct coupling clutch Ci is connected by the direct coupling clutch control 100, whereby the output of the engine 8 is changed to the automatic transmission 10. The first motor generator MG1 or the second motor generator MG2 is brought into a power generation state by the MG2 controller 102 or the MG1 controller 104 as necessary, and the generated energy is stored in the power storage device 108. The In the motor and engine running control, the output of the engine 8 and the output of the first motor generator MG1 and / or the second motor generator MG2 are automatically transmitted in the state where the direct coupling clutch Ci is connected for acceleration running. It is directly transmitted to 10 input shafts 16. In the regenerative braking control, the second motor generator MG2 is brought into a power generation state by the MG2 controller 102 in order to obtain a desired braking force at the time of a braking operation in which the brake pedal is operated or during coasting, and the regenerative power consumed for the power generation. A braking force is obtained by the torque, and the generated energy is stored in the power storage device 108 via the inverter 106.

FIG. 8 is a functional block diagram for explaining the main part of the control function of the electronic control unit 90, that is, the shock relaxation control at the end of shifting. In FIG. 8, the shift control means 110 performs shift determination based on the actual vehicle speed V and the throttle opening θ _TH from the shift map stored in advance as shown in FIG. 6, for example, to execute the determined shift. The gear position of the automatic transmission 16 is automatically switched by performing the gear shift output. For example, in the 2 → 3 upshift, the brake B1 is started to be released, and when the engagement torque is maintained to some extent, the engagement of the clutch C3 is started to generate the engagement torque. While shifting from the second speed gear ratio γ ₂ to the third speed gear ratio γ ₃ , the release of the brake B1 and the engagement of the clutch C3 are completed. The shift determination means 112 determines whether or not the shift by the shift control means 110 has started, that is, the time t _{1 in} the time charts of FIGS. 9 and 10, based on the shift output of the shift control means 110, for example. . The shift end determining means 114 determines whether or not the shift is completed, whether or not the change in the rotation speed at which the rotation speed Nin of the input shaft 16 changes from the value before the shift to the value after the shift is completed, that is, the rotation speed Nin of the input shaft 16 is shifted. Judgment is made based on whether or not it coincides with the subsequent synchronous rotation speed. Further, the shift end determination means 114 determines immediately before the end of the shift based on the fact that the rotational speed Nin of the input shaft 16 is close to the synchronous rotational speed. For example, in the 2 → 3 upshift in FIG. 9, it is time t ₅ when the rotational speed Nin of the input shaft 16 coincides with the third speed synchronous rotational speed (= rotational speed Nout × γ ₃ ) or the time t ₄ immediately before that. Is judged. Further, in the 3 → 2 downshift in FIG. 10, at the time t ₅ when the rotational speed Nin of the input shaft 16 coincides with the _second synchronous speed (= rotational speed Nout × γ ₂ ) or at the time t ₄ immediately before that. It is determined whether or not there is. In the time charts of FIGS. 9 and 10, an example is shown in which torque vibration that occurs immediately after the shift is applied simultaneously with the application of reverse phase torque vibration simultaneously with the slip of the direct coupling clutch Ci. The application of the reverse phase torque vibration is executed when it is determined that slip of the direct clutch Ci is impossible as will be described later.

  The slip control availability determination means 116 determines whether or not the direct connection clutch control means 120 can control to slip or release the direct connection clutch Ci in order to suppress torque vibration generated in the power transmission path at the end of the shift. That is, the determination is based on the presence or absence of a failure of the direct clutch control valve 96, a failure of the direct clutch Ci, or overheating.

The direct coupling clutch control means 120 determines the magnitude of torque vibration generated in the power transmission path at the end of the shift when the shift end determination means 114 determines the end of the shift or immediately before the end of the shift. In order to suppress the slip control amount, the slip control amount is determined based on the actual gear ratio γ of the automatic transmission 10 from the pre-stored relationship shown in FIG. In FIG. 9 or FIG. 10, the direct clutch is used by using the direct clutch control valve 96 so that the determined slip control amount can be obtained in the section from the time t ₄ immediately before the synchronization to the time t ₆ that is the end of the antiphase torque vibration. Ci is controlled and slips temporarily to lower the engagement torque. The slip control amount is the slip amount of the direct coupling clutch Ci. Note that the torque vibration generated at the end of the shift is illustrated by broken line pulsation in FIGS.

  Further, the direct clutch control means 120 determines a slip ratio based on the torque control rate (%) from, for example, a previously stored relationship shown in FIG. 12, and sets the direct clutch control valve 96 so that the slip ratio can be obtained. Used to control the direct coupling clutch Ci and cause it to slip temporarily. The torque control rate is the ratio (ratio) of the torque control amount that actually acts by the anti-phase torque vibration output control means 122 and the engine output reduction means 124 with respect to the control amount necessary for suppressing the torque vibration. is there. The slip ratio is a ratio of the slip amount actually acting on the slip amount necessary for suppressing the torque vibration. In the relationship shown in FIG. 12, a torque control amount may be used instead of the torque control rate. If the torque control rate does not change so much, the same effect can be obtained. As described above, the direct coupling clutch control means 120 reduces the slip control amount when combined with the suppression of torque vibration by other means as compared with the case where it is carried out alone. The relationship shown in FIG. 12 is a relationship in which the slip rate is increased in order to compensate for the lower torque control rate, and the slip rate is decreased as the torque control rate is increased.

  For example, as shown in FIG. 9 and FIG. 10, the suppression torque control means 118 is used to suppress torque vibration immediately after the end of a shift in a state where the direct connection clutch Ci is used for slipping or releasing the direct connection clutch Ci. The torque vibration opposite in phase to the torque vibration is generated from the first motor generator MG1 or the second motor generator MG2 to cancel the torque vibration, the direct clutch Ci is released or slipped, and the output of the engine 8 is The torque vibration is suppressed by temporarily reducing the torque vibration.

  The suppression torque control means 118 first generates torque vibration having a phase opposite to that of the torque vibration generated in the power transmission path from the engine 8 to the drive wheels due to the stop of the change in the rotational speed of the engine 8 at the end of the shift, for example, immediately after the shift. The motor generator MG1 or the second motor generator MG2 generates the torque vibration, cancels or slips the direct coupling clutch Ci, temporarily reduces the output of the engine 8, and the automatic transmission 10 A suppression torque for suppressing torque vibration generated in the power transmission path of the vehicle at the end of the shift is applied to the power transmission path, thereby suppressing the torque vibration. The torque vibration generated at the end of the shift is indicated by broken-line pulsations in FIGS.

The suppression torque control means 118 includes an antiphase torque vibration output control means 122 and an engine output reduction means 124. The reverse phase torque vibration output control means 122, when it is determined by the slip control availability determination means 116 that the direct clutch control means 120 cannot control to slip or release the direct clutch Ci, the torque generated immediately after the shift. An antiphase torque vibration having a phase opposite to that of the vibration is applied to the power transmission system to suppress the torque vibration. The anti-phase torque vibration output control means 122 determines the magnitude of the anti-phase torque vibration based on the actual gear ratio γ of the automatic transmission 10 based on the relationship stored in advance as shown in FIG. The reverse phase torque vibration is output from the first motor generator MG1 or the second motor generator MG2 to the power transmission path. 9 or section from t ₅ the time in FIG. 10 t to ₆ time shows this state. The torque vibration generated in the power transmission path at the end of the shift is specific to the power transmission system of the vehicle, and its frequency and damping rate are obtained and stored in advance experimentally. Is a waveform inverted with respect to the above-described torque vibration, and has the same frequency and damping rate as the torque vibration. The magnitude of the anti-phase torque vibration is the torque control amount of the anti-phase torque vibration output control means 120 shown on the vertical axis in FIG. 13 and indicates the amplitude of the first waveform of the anti-phase torque vibration. The relationship shown in FIG. 13 is a relationship in which the torque control amount decreases as the speed ratio γ decreases.

  Further, the anti-phase torque vibration output control means 122 further determines the torque control amount based on the temporary reduction amount of the output torque of the engine 8 by the engine output reduction means 124 from the previously stored relationship shown in FIG. Then, the first motor generator MG1 or the second motor generator MG2 outputs the determined torque control amount, that is, the magnitude of the reverse phase torque vibration to the power transmission path.

  Further, the engine output lowering means 124 suppresses the magnitude of torque vibration generated in the power transmission path at the end of the shift, during the inertia phase in the shift transition period in the up shift, and in the shift transition period in the down shift. At the end of the inertia phase, for example, the output torque of the engine 8 is temporarily reduced by operating, for example, retarding the ignition timing, closing the throttle valve opening, and limiting the fuel injection amount.

  FIG. 15 is a flowchart for explaining a main part of the control operation of the electronic control unit 90, that is, the shock mitigation control operation at the end of the shift. This flowchart is repeatedly executed at a predetermined cycle.

  In FIG. 15, in step (hereinafter, step is omitted) S1 corresponding to the shift determining means 112, it is determined whether or not shift determination for shifting the automatic transmission 10 has been performed. If the determination in S1 is negative, the routine is terminated after other control is executed in S2. However, if the determination in S1 is affirmative, the output torque of the engine 8 is temporarily reduced in S3 corresponding to the engine output reduction means 124. This temporary decrease in engine output torque is executed over the entire inertia phase, which is a section where the engine speed NE changes as shown in FIG. 9 in the upshift, and in the downshift, the end of the inertia phase as shown in FIG. Executed in

  Next, in S4 corresponding to the shift end determination means 114, it is determined whether or not the shift is ended based on, for example, that the engine rotation speed NE matches the synchronous rotation speed determined by the gear stage after the shift. If the determination in S5 is negative, the routine is terminated after other control is executed in S2. However, if the determination in S4 is affirmative, in S5 corresponding to the slip control availability determination means 116, control can be performed to slip or release the direct clutch Ci in order to suppress torque vibration that occurs at the end of the shift. Is determined based on the abnormality of the direct clutch control valve 96, the abnormality of the direct clutch Ci, or the presence or absence of overheating.

  If the determination in S5 is affirmative, in S6 corresponding to the direct coupling clutch control means 120, the slip amount and the slip ratio are determined based on the speed ratio γ and the torque control rate from the relationship shown in FIGS. While the antiphase torque vibration is applied to the power transmission system, the direct coupling clutch Ci is slipped so as to achieve the determined slip amount and slip ratio, and the engagement torque is temporarily reduced. . In S8, a shift end process is executed, and this routine is ended.

However, if the determination at S5 is affirmative, it occurs when the rotational speed NE of the engine 8 matches the synchronous rotational speed and is maintained at S8 corresponding to the antiphase torque vibration output control means 122. The magnitude (torque control amount) of the torque vibration having the opposite phase to the torque vibration is determined based on the actual gear ratio γ of the automatic transmission 10 from the relationship stored in advance as shown in FIG. 13, for example, as shown in FIG. The torque control amount is determined based on the torque reduction amount by the engine output reduction means 122 from the relationship stored in advance, and the reverse phase torque vibration corresponding to the determined torque control amount is the first motor generator MG1 or the second motor. The power is output from the generator MG2 to the power transmission path. 9 or section from t ₅ the time in FIG. 10 t to ₆ time shows this state. In step S8, a shift end process is executed.

  As described above, the present embodiment includes the direct clutch control means 120 for slipping or releasing the direct clutch Ci at the end of the shift of the automatic transmission 10, and therefore occurs due to a sudden stop of the engine speed change at the end of the shift. Since the torque vibration to be absorbed is absorbed by the slip or release of the direct coupling clutch Ci, the shift shock due to the torque vibration at the end of the shift is suitably suppressed.

  In this embodiment, the direct connection clutch operation enable / disable determination means 116 for determining whether or not slip control or release control of the direct connection clutch Ci is possible at the end of the shift, and the direct connection clutch operation enable / disable determination means 116 are used. Suppression that imparts suppression torque to the power transmission path for suppressing torque vibration generated in the power transmission path of the vehicle at the end of shifting of the automatic transmission 10 when it is determined that slip or release control is impossible. The torque control means 118 is further included. Therefore, when the torque vibration generated due to the sudden stop of the engine speed change at the end of the shift is not absorbed by the slip or release of the direct coupling clutch Ci, the power transmission path is replaced instead. Torque vibration by the above-mentioned suppression torque for suppressing torque vibration applied to Since is suppressed, shift shock due to torque fluctuation in the gear shifting completion is suitably suppressed.

  Further, in this embodiment, the first motor generator MG1 and / or the second motor generator MG2 (electric motor) operatively connected to the input shaft 16 of the automatic transmission 10 is provided, and the suppression torque control means 118 includes Since the suppression torque for suppressing the torque vibration is output from the electric motor, the torque vibration is suppressed by the suppression torque applied to the power transmission path from the electric motor. Shock is suitably suppressed.

  Further, in this embodiment, the suppression torque control means 118 outputs an antiphase torque vibration having a phase opposite to that of the torque vibration from the electric motor as the suppression torque for suppressing the torque vibration. 122, the torque vibration is canceled by the antiphase torque vibration applied to the power transmission path from the electric motor, so that the shift shock due to the torque vibration at the end of the shift is suitably suppressed.

  In this embodiment, the engine output reduction means 124 for temporarily reducing the output torque of the engine 8 at the end of the shift of the automatic transmission 10 is further provided. Since the output torque is reduced by a temporary decrease in the output torque, the slip control amount of the direct coupling clutch for absorbing it and the antiphase torque vibration for canceling it are reduced.

  Further, in the present embodiment, the antiphase torque output control means 122 changes the magnitude of the antiphase torque vibration depending on the amount of inertia torque reduction that occurs at the end of the shift, so that the appropriate magnitude is obtained. By applying the reverse phase torque vibration to the power transmission system, the torque vibration generated at the end of the shift is preferably canceled out.

  Further, in the present embodiment, the anti-phase torque output control means 122 changes the magnitude of the anti-phase torque vibration depending on the type of shift, so that the anti-phase torque vibration of an appropriate magnitude is obtained. Is applied to the power transmission system, so that the torque vibration generated at the end of the shift is preferably canceled out.

  Further, in the present embodiment, the direct clutch control means 120 changes the slip amount of the direct clutch Ci depending on the amount of inertia reduction that occurs at the end of the shift. Torque vibration generated at the end of the shift is suitably absorbed by the slip being generated in the direct coupling clutch Ci inserted in series in the power transmission system.

  Further, in the present embodiment, the direct clutch control means 120 changes the slip amount of the direct clutch Ci depending on the type of the shift, so that slip of an appropriate amount of slip is caused by the power transmission system. Torque vibrations generated at the end of the shift by being generated in the direct coupling clutch Ci inserted in series are preferably absorbed.

  FIG. 16 is a flowchart for explaining the operation of the main part of another embodiment of the control operation of the electronic control unit 90, that is, the operation of another embodiment of the shock relaxation control operation at the end of shifting. This flowchart is also repeatedly executed at a predetermined cycle.

  In FIG. 16, step S11 corresponding to the shift determining means, S12 in which other control is executed, S13 corresponding to the engine output lowering means, and S14 corresponding to the shift end determining means are S1, S2, It is executed in the same manner as S3 and S4. That is, if the determination in S11 is negative, the routine is terminated after other controls in S12 are executed. If the determination in S11 is affirmative, in S13 corresponding to the engine output reduction means 124, The output torque of the engine 8 is temporarily reduced. This temporary decrease in engine output torque is executed over the entire inertia phase, which is a section where the engine speed NE changes as shown in FIG. 9 in the upshift, and in the downshift, the end of the inertia phase as shown in FIG. Executed in Next, in S14, whether or not the shift is completed is determined based on, for example, that the engine rotation speed NE matches the synchronous rotation speed determined by the gear stage after the shift, as in S4 of FIG.

  If the determination in S14 is negative, the routine is terminated after other control is executed in S2. However, if the determination in S14 is affirmative, the engine output torque, that is, the input torque of the automatic transmission 10 can be controlled to be temporarily reduced in S15 corresponding to the input torque down control availability determination means. Is determined based on the failure of the ignition timing adjusting device, the presence or absence of overheating of the catalyst, and the like.

  If the determination in S15 is affirmative, in S16 corresponding to the antiphase torque vibration output control means, the reverse of the torque vibration generated when the rotational speed NE of the engine 8 matches and is maintained at the synchronous rotational speed. The magnitude of the phase torque vibration is determined based on the actual gear ratio γ of the automatic transmission 10 from the relationship stored in advance as shown in FIG. 11, for example, and the determined reverse phase torque vibration is the first motor generator MG1. Alternatively, it is output from the second motor generator MG2 to the power transmission path. At the same time, in S17 corresponding to the direct clutch control means, in order to suppress the magnitude of torque vibration generated in the power transmission path at the end of the shift, for example, the actual torque control rate is changed from the previously stored relationship shown in FIG. Based on this, the slip ratio is determined, and while the antiphase torque vibration is applied to the power transmission system, the engagement torque of the direct clutch Ci is temporarily reduced so as to achieve the determined slip ratio. In step S8, a shift end process is executed.

  However, if the determination in S15 is negative, the slip ratio is determined to be larger in S19 corresponding to the direct clutch control means than in S17, and the antiphase torque vibration is applied to the power transmission system. During this time, the engagement torque of the direct clutch Ci is temporarily reduced so as to achieve the determined slip ratio. In step S8, a shift end process is executed.

  As described above, also in the present embodiment, similarly to the above-described embodiment, the shock of the shift cycle is suitably suppressed. In this embodiment, when the input torque down control is impossible, the engagement torque of the direct clutch Ci is temporarily reduced so that the slip ratio is determined to be larger than in the case of S17. Therefore, even when the input torque down control cannot be executed, the shift shock is preferably mitigated. In the present embodiment, S17 executed when the determination of S15 is affirmed may not necessarily be provided, and the step corresponding to the antiphase torque vibration output control means may be performed before S19 or It may be provided adjacent to the rear.

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

  For example, in the embodiment described above, the anti-torque torque vibration output control means 122, the suppression torque control means 118 including the engine output reduction means 124, and the slip control availability determination means 116 are not necessarily provided.

  In the above-described embodiment, the second motor generator MG2 is provided on the input shaft 16, but is provided downstream of the input shaft 16 of the automatic transmission 16, for example, between the input shaft 16 and the output shaft 28. It may be.

  In the above-described embodiment, the direct clutch control means 122 temporarily slips the direct clutch Ci within the shift period in order to reduce torque vibration. Instead, the direct clutch Ci is temporarily switched. May be released.

  In the direct clutch control means 120 of the above-described embodiment, the slip amount and slip ratio of the direct clutch Ci are determined depending on the actual gear ratio γ of the automatic transmission 10 from the relationship shown in FIG. 12 is determined depending on the torque control rate related to the engine output lowering means 124 that is temporarily executed within the speed change period, the speed ratio γ and the torque of the automatic transmission 10 are determined. It may be determined depending on one of the control rates, or may not necessarily be determined depending on the gear ratio γ and torque reduction amount of the automatic transmission 10. For example, even if the slip amount is a constant value, a corresponding effect can be obtained.

  Further, in the above-described embodiment, the horizontal axis of FIGS. 11 and 13 indicates the speed ratio γ before or after the automatic transmission 10 is shifted, but 1 → 2 shift, 2 → 3 shift, 3 → It may indicate the type of shift such as four shifts, or may indicate the size of the inertia torque. In short, torque vibration is generated corresponding to the magnitude of the inertia torque generated when the rotational speed of the engine 8 is matched with the synchronous rotational speed immediately after the end of the shift and the rotational speed change of the engine 8 is stopped. The horizontal axis may be a parameter related to the size of the inertia torque. In the case of the jumping speed change, the gear ratio γ of the automatic transmission 10 is greatly changed at a stroke, so that the torque control amount is also greatly changed from the relationship of FIG.

  Further, in the anti-phase torque vibration output control means 122 of the above-described embodiment, the magnitude of the anti-phase torque vibration (torque control amount) depends on the actual gear ratio γ of the automatic transmission 10 from the relationship shown in FIG. 14 or depending on the amount of torque reduction by the engine output reduction means 124 temporarily executed within the shift period from the relationship shown in FIG. It may be determined depending on one of the ratio γ and the torque decrease amount, or may not necessarily be determined depending on the gear ratio γ and the torque decrease amount of the automatic transmission 10. For example, a corresponding effect can be obtained even if the torque control amount is a constant value.

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

BRIEF DESCRIPTION OF THE DRAWINGS FIG. 1 is a skeleton diagram illustrating a main configuration of a stepped automatic transmission to which a control device according to an embodiment of the present invention is applied. FIG. 3 is an alignment chart for explaining the operation of the stepped automatic transmission of FIG. 2. 2 is an operation table showing a relationship between a gear position of the stepped automatic transmission of FIG. 1 and a combination of operations of a hydraulic friction engagement device necessary to establish it. It is a figure explaining the control system of the control apparatus for controlling the stepped automatic transmission of FIG. It is a figure which shows the relationship used in the electronic throttle valve opening control of the electronic control apparatus of FIG. It is a figure which shows the shift diagram used in the shift control of the electronic controller of FIG. It is a figure explaining the operation position of the shift lever of FIG. It is a functional block diagram explaining the principal part of the control function of the electronic control apparatus of FIG. FIG. 5 is a time chart for explaining a control operation of the electronic control device of FIG. 4, showing an operation at the time of a 2 → 3 upshift. FIG. FIG. 5 is a time chart for explaining a control operation of the electronic control device of FIG. 4, showing an operation at a 3 → 2 down shift. FIG. 9 is a diagram showing a relationship used when determining the slip amount based on the gear ratio of the automatic transmission in the direct clutch control means of FIG. 8. FIG. 9 is a diagram illustrating a relationship used when determining a torque control amount based on a torque control rate in the direct clutch control unit of FIG. 8. FIG. 9 is a diagram showing a relationship used when determining the torque control amount based on the gear ratio of the automatic transmission in the antiphase torque vibration output control means of FIG. 8. FIG. 9 is a diagram illustrating a relationship used when determining a torque control amount based on a torque reduction amount by an engine output torque reduction unit in the antiphase torque vibration output control unit of FIG. 8. It is a flowchart explaining the principal part of the control action of the electronic controller of FIG. It is a flowchart explaining the principal part of the control action of the electronic control apparatus in the other Example of this invention.

Explanation of symbols

8: Engine 10: Automatic transmission (Stepped automatic transmission)
16: Input shaft 90: Electronic control device (control device)
116: Reverse phase torque application availability determination means 118: Suppression torque control means 120: Direct coupling clutch control means 122: Reverse phase torque vibration output control means 124: Engine output reduction means Ci: Direct coupling clutch MR1: First motor generator (electric motor)
MR2: Second motor generator (electric motor)

Claims (11)

A stepped automatic transmission, and a direct coupling clutch interposed between the engine and the input shaft of the stepped automatic transmission, the input shaft of the stepped automatic transmission being interposed via the direct coupling clutch A control device for a stepped automatic transmission for a vehicle that performs a shift while being mechanically connected to an engine,
A control device for a stepped automatic transmission for a vehicle, comprising: a direct coupling clutch control means for slipping or releasing the direct coupling clutch at the end of shifting of the stepped automatic transmission. Direct coupling clutch operation availability determination means for determining whether or not the direct coupling clutch can be slipped or released at the end of the shift;
Torque vibrations generated in the power transmission path of the vehicle at the end of shifting of the stepped automatic transmission are suppressed when the direct clutch is determined not to be able to slip or disengage by the direct clutch activatability determining means. 2. The control device for a stepped automatic transmission for a vehicle according to claim 1, further comprising: suppression torque control means for applying a suppression torque to the power transmission path. An electric motor operatively connected to the input shaft of the stepped automatic transmission;
The control device for a stepped automatic transmission for a vehicle according to claim 2, wherein the suppression torque control means outputs a suppression torque for suppressing the torque vibration from the electric motor.   4. The vehicle according to claim 3, wherein the suppression torque control means includes an anti-phase torque output control means for outputting an anti-phase torque vibration having a phase opposite to the torque vibration from the electric motor as the suppression torque for suppressing the torque vibration. Control device for stepped automatic transmission.   5. The stepped automatic for vehicles according to any one of claims 1 to 4, further comprising an engine output lowering means for temporarily reducing an output torque of the engine at the end of shifting of the stepped automatic transmission. Transmission control device.   5. The stepped automatic transmission for a vehicle according to claim 4, wherein the reverse phase torque output control means changes the magnitude of the reverse phase torque vibration depending on an amount of decrease in inertia torque generated at the end of the shift. Control device.   5. The control device for a stepped automatic transmission for a vehicle according to claim 4, wherein the reverse phase torque output control means changes the magnitude of the reverse phase torque vibration depending on the type of the shift.   The stepped automatic transmission for a vehicle according to any one of claims 1 to 7, wherein the direct coupling clutch control means changes a slip amount of the direct coupling clutch depending on an amount of decrease in inertia torque generated at the end of the shift. Control device.   The control device for a stepped automatic transmission for a vehicle according to any one of claims 1 to 7, wherein the direct clutch control means changes a slip amount of the direct clutch depending on a type of the shift. A stepped automatic transmission, and a direct coupling clutch interposed between the engine and the input shaft of the stepped automatic transmission, the input shaft of the stepped automatic transmission being interposed via the direct coupling clutch A control device for a stepped automatic transmission for a vehicle that performs a shift while being mechanically connected to an engine,
Engine output lowering means for temporarily lowering the output torque of the engine at the end of shifting of the stepped automatic transmission;
And a direct-coupled clutch control means for slipping or releasing the direct-coupled clutch at the end of shifting of the step-variable automatic transmission. A stepped automatic transmission, and a direct coupling clutch interposed between the engine and the input shaft of the stepped automatic transmission, the input shaft of the stepped automatic transmission being interposed via the direct coupling clutch A control device for a stepped automatic transmission for a vehicle that performs a shift while being mechanically connected to an engine,
Engine output lowering means for temporarily reducing the output torque of the engine at the end of shifting of the stepped automatic transmission;
Direct coupling clutch control means for slipping or releasing the direct coupling clutch at the end of shifting of the stepped automatic transmission when the engine output torque cannot be temporarily reduced by the engine output reduction means. A control device for a stepped automatic transmission for vehicles.

Images