JP2008106871A - Control device for vehicular automatic transmission - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To perform a smooth gear shift by balancedly attempting compatibility between shortening of a speed-change time in upshifting and a reduction in the engagement shock of hydraulic engagement elements, in a control device 5 for a vehicular automatic transmission 1 including a transmission mechanism part 3 having at least two planetary mechanisms 31, 32 and plural hydraulic engagement elements C1-C4, B1, B2, and a hydraulic control unit 4 engaging or disengaging the hydraulic engagement elements. <P>SOLUTION: The control device 5, for controlling the hydraulic control unit 4, performs speed-change processing for engaging or disengaging the hydraulic engagement elements required for inertia phase control in speed-change on the basis of the previous hydraulic pressure command pattern; and correcting processing for measuring elapsed time (T1-T3) in plural zones from receiving of a speed-change command until finishing of inertia phase in the inertial phase control in upshifting so that based on this measured results, the previous hydraulic pressure command pattern (P) is deformed to form a new hydraulic pressure command pattern (as shown in Fig.17). <P>COPYRIGHT: (C)2008,JPO&INPIT

Description

  The present invention relates to a control device for an automatic transmission for a vehicle. In particular, the vehicle automatic transmission includes a plurality of hydraulic systems for establishing a power transmission path of each planetary mechanism corresponding to at least two planetary mechanisms installed from an input shaft to an output shaft and a required gear stage. It is comprised including the transmission mechanism part which has an engagement element, and the hydraulic control apparatus which engages or releases each hydraulic engagement element as needed.

  2. Description of the Related Art Conventionally, there is an automatic transmission for a vehicle in which a plurality of gear planetary mechanisms are combined to perform a multi-stage shift (see, for example, Patent Document 1).

  The speed change mechanism in the vehicle automatic transmission includes three gear type planetary mechanisms and a plurality of hydraulic engagement elements for establishing a power transmission path of each planetary mechanism corresponding to a required gear stage. It is configured to include.

  The hydraulic engagement element is a wet friction brake for bringing appropriate components in the three-gear planetary mechanism into a rotatable engagement state or a non-rotatable release state, and each element can be rotated integrally. Examples include a wet friction clutch for engaging and releasing a relatively rotatable state.

  A plurality of hydraulic engagement elements such as brakes and clutches are individually driven by a hydraulic control device having the same number of linear solenoid valves.

  Here, when the linear solenoid valve is of a normally closed type, if the solenoid of the linear solenoid valve is de-energized (off) and closed, the hydraulic engagement element is released, while the solenoid is energized. If an appropriate hydraulic pressure is applied to the hydraulic engagement element by controlling the valve opening, the engagement state is established.

  In this type of automatic transmission for a vehicle, a shift process called so-called upshift or downshift is performed while the vehicle is running. In either case, the shift stage before the shift is changed to the target shift stage. Therefore, it is necessary to disengage unnecessary hydraulic engagement elements or to engage necessary hydraulic engagement elements.

  When the hydraulic engagement element is engaged, so-called inertia phase control is performed to finely control the opening of the linear solenoid valve, that is, the hydraulic pressure applied to the hydraulic engagement element.

  The inertia phase is a section in which the input shaft rotation speed (turbine rotation speed) of the automatic transmission changes according to the deviation between the synchronous rotation speed of the gear stage before the shift and the required synchronous rotation speed of the gear stage. That is. The input shaft rotation speed in the inertia phase decreases during an upshift and increases during a downshift. The synchronous rotation speed is the input shaft rotation speed that matches the gear ratio of the shift speed.

  The engagement hydraulic pressure to be applied to the hydraulic engagement element in this inertia phase control is determined empirically so as to shorten the time required for the shift (shift time) and reduce the engagement shock of the hydraulic engagement element. This is performed based on the hydraulic command pattern.

By the way, if the shift time is shortened, the engagement shock increases, and the input shaft rotation speed and output shaft torque of the automatic transmission change abruptly. This makes it easier for the driver to recognize the shift shock. End up. However, if the shift time is lengthened, the engagement shock is reduced and the shift shock can be reduced, but the driver is given a feeling of extension during the shift.
JP 2003-287122 A

  When the hydraulic pressure command pattern is fixed as in the above-described conventional example, it is difficult to achieve a balance between shortening the shifting time and reducing the engagement shock of the hydraulic engagement element.

  Therefore, the inventors of the present application investigated the change in the input shaft rotation speed in the previous inertia phase control according to, for example, the driver's driving and traveling conditions without fixing the hydraulic pressure command pattern, and based on the result. It is considered to correct the hydraulic pressure command pattern.

  However, the current correction does not change the hydraulic pressure command pattern, but simply changes the level of the constant pressure standby pressure, and is insufficient. Here, it is considered that there is room to devise the form of correction, and the present invention has been filed.

  According to the present invention, in a control device for an automatic transmission for a vehicle, a reduction in shift time during an upshift and a reduction in engagement shock of a hydraulic engagement element are balanced and a smooth shift can be achieved. It is aimed.

  The present invention has at least two planetary mechanisms installed from the input shaft to the output shaft, and a plurality of hydraulic engagement elements for establishing a power transmission path of each planetary mechanism corresponding to the required gear stage. A control device for an automatic transmission for a vehicle including a speed change mechanism and a hydraulic control device that engages or releases each hydraulic engagement element as necessary, and is a hydraulic type that is necessary for inertia phase control at the time of shifting Shift processing that engages or disengages the engagement element based on the previous hydraulic pressure command pattern, and measures the elapsed time of multiple sections from when the shift command is received in the inertia phase control during upshift until the inertia phase ends. Then, based on this measurement result, a correction process is performed in which the previous hydraulic pressure command pattern is transformed into a new hydraulic pressure command pattern.

  In this configuration, the hydraulic command pattern in the inertia phase control is modified and updated according to the actual turbine speed (input shaft speed) for each upshift. It is possible to create a hydraulic command pattern adapted to the driving and driving conditions of the vehicle, and ideal inertia phase control becomes possible.

  As a result, it is possible to balance the reduction of the shift time and the reduction of the engagement shock of the hydraulic engagement element in a balanced manner, so that the driver does not need to feel extended during the upshift and the shift shock. Become.

  Preferably, the hydraulic engagement element integrally rotates the two components of each planetary mechanism and a wet friction brake that makes a suitable component of each planetary mechanism rotatable or non-rotatable. It is possible to have a configuration having a wet friction clutch that is in a possible state or a relatively rotatable state.

  In this configuration, the hydraulic engagement element is specified to be frictionally engaged via hydraulic oil.

  Preferably, in the correction process, a deviation between the measured value of the elapsed time of the first section from the reception of the shift command in the inertia phase control to the start of the inertia phase and the target value is calculated, and based on the calculation result, The constant pressure standby pressure in the hydraulic pressure command pattern can be corrected to increase or decrease.

  In this configuration, if the constant pressure standby pressure in the hydraulic command pattern is first corrected to be increased, the elapsed time of the first section is shortened and the increase in the input shaft rotational speed at the start of the inertia phase tends to be suppressed. Become.

  On the other hand, when the constant pressure standby pressure in the hydraulic pressure command pattern is corrected to be lowered, the elapsed time of the first section becomes longer and the input shaft rotational speed at the start of the inertia phase tends to increase.

  Preferably, in the correction process, a deviation between the measured value of the elapsed time of the second section from the start of the inertia phase until the input shaft rotation speed reaches the predetermined rotation speed and the target value in the inertia phase control is calculated. Based on the calculation result, the constant pressure standby pressure in the hydraulic pressure command pattern can be gradually increased from the start point of the inertia phase, and the pressure increase gradient can be corrected to be increased or decreased.

  When the constant pressure standby pressure in the hydraulic pressure command pattern is corrected to gradually increase from the start of the inertia phase as in this configuration, the elapsed time of the second section is shortened.

  When the pressure increase gradient is increased, the elapsed time of the second section is shortened. On the other hand, when the pressure increase gradient is decreased, the elapsed time of the second section tends to be longer. In any case, the correction makes it possible to promote a decrease in the rotational speed of the input shaft after the shift command, and is advantageous in shortening the shift time.

  Preferably, in the correction process, a deviation between the measured value of the elapsed time of the third section from the start of the inertia phase to the establishment of the requested gear stage and the target value in the inertia phase control is calculated, and the calculated result is On the basis of this, it is possible to gradually reduce the constant pressure standby pressure in the hydraulic pressure command pattern at the end stage and correct the pressure reducing gradient in a direction to increase or decrease.

  If the correction is made to gradually decrease at the end of the constant pressure standby pressure in the hydraulic pressure command pattern as in this configuration, the output shaft torque at the end of the inertia phase is gradually decreased.

  When the depressurization gradient is increased, the elapsed time of the third section becomes longer, whereas when the depressurization gradient is decreased, the elapsed time of the third section tends to be shortened. In any case, the correction reduces the hydraulic pressure when the required hydraulic engagement element is engaged, which is advantageous in reducing the engagement shock.

  The present invention has at least two planetary mechanisms installed from the input shaft to the output shaft, and a plurality of hydraulic engagement elements for establishing a power transmission path of each planetary mechanism corresponding to the required gear stage. A control device for an automatic transmission for a vehicle, including a speed change mechanism and a hydraulic control device that engages or releases each hydraulic engagement element as necessary, and includes a hydraulic command pattern for each shift stage and an appropriate target Storage means for storing values, input shaft rotation speed detection means for detecting the rotation speed of the input shaft, measurement means for measuring the elapsed time of a plurality of sections from when receiving a shift command until the inertia phase ends, The deviation between the measured value and the target value is calculated for each section, and based on the calculation result, a correcting means for deforming a predetermined portion of the previous hydraulic pressure command pattern, and the deformed new hydraulic pressure command pattern is changed to the next hydraulic pressure command pattern. And and a learning means for updating the hydraulic pressure command pattern in said storage means, and characterized in that.

  Here, the configuration of the control device according to the present invention is specified in more detail than the problem solving configuration described above.

  Preferably, the measuring means includes a first measuring means for measuring an elapsed time of the first section from when a shift command is received in the inertia phase control until the start of the inertia phase, and from the start of the inertia phase in the inertia phase control. A second measuring means for measuring an elapsed time of the second section until the input shaft rotational speed reaches a predetermined rotational speed, and a third section from the start of the inertia phase to the establishment of the requested shift stage in the inertia phase control. And a third measuring means for measuring the elapsed time of, and the correcting means calculates a deviation between the measured value of the first measuring means and the first target value corresponding to the calculated value. First correction means for correcting the level of the constant pressure standby pressure in the previous hydraulic pressure command pattern based on the result obtained by multiplying an appropriate gain, the measurement value by the second measurement means, and the corresponding value And a constant pressure standby pressure in the previous hydraulic pressure command pattern is gradually increased from the start point of the inertia phase based on a result obtained by multiplying the calculated deviation by an appropriate gain, and the pressure increase gradient Based on the result of multiplying the calculated deviation by an appropriate gain and calculating the deviation between the second correction means for correcting the magnitude of the error and the measurement value obtained by the third measurement means and the corresponding third target value. And a third correction means for gradually decreasing the constant pressure standby pressure at the end stage of the hydraulic pressure command pattern and correcting the magnitude of the pressure reduction gradient.

  In this configuration, the form in which the hydraulic pressure command pattern is deformed is clarified by finely specifying the measurement unit and the correction unit that constitute the control device according to the present invention.

  According to the present invention, it is possible to perform inertia phase control adapted to each driver's driving and driving situation for each upshift accompanying the progress of driving, shortening the shift time and engaging shock of the hydraulic engaging element. It becomes possible to achieve both reduction and balance. As a result, smooth shifting can be achieved, such as eliminating the need for the driver to feel extended during the upshift and shifting shock.

  Hereinafter, embodiments of the present invention will be described with reference to the drawings. 1 to 17 show an embodiment of the present invention.

  First, prior to the description of the features of the present invention, the configuration of the vehicle automatic transmission 1 to which the features of the present invention are applied will be described.

  FIG. 1 is a diagram showing a schematic configuration of a vehicle power train using a vehicular automatic transmission 1 to which the present invention is applied, and FIG. 2 is a skeleton diagram showing an example of the vehicular automatic transmission 1 of FIG. FIG. 3 is a perspective view schematically showing the transmission mechanism unit 3 of FIGS. 1 and 2.

  An automatic transmission 1 for a vehicle shown in FIG. 1 mainly includes a torque converter 2 as a fluid transmission device, a transmission mechanism unit 3, a hydraulic control device 4, a transmission control device 5, and an oil pump 6, and includes eight forward speeds and two reverse gears. It is configured to be able to change gears.

  In FIG. 1, 7 is an engine (internal combustion engine), and 8 is an engine control device that controls the operation of the engine 7. The engine control device 8 is connected to the transmission control device 5 so as to be able to transmit and receive each other. A powertrain is configured including the vehicular automatic transmission 1 and the engine 7.

  The torque converter 2 is rotationally connected to the engine 7 and includes a pump impeller 21, a turbine runner 22, a stator 23, a one-way clutch 24, a stator shaft 25, and a lockup clutch 26.

  The one-way clutch 24 supports the stator 23 while allowing the stator 23 to rotate only in one direction on the case 1a of the transmission mechanism unit 3. The stator shaft 25 fixes the inner race of the one-way clutch 24 to the case 1a. The lockup clutch 26 directly connects the pump impeller 21 and the turbine runner 22.

  The speed change mechanism 3 shifts the rotational power input from the torque converter 2 to the input shaft 9 and outputs it to the output shaft 10. As shown in FIGS. 2 and 3, the front planetary 31 and the rear planetary 32 are provided. And an intermediate drum 33, first to fourth clutches C1 to C4, and first and second brakes B1 and B2.

  The front planetary 31 is a gear type planetary mechanism called a double pinion type, and includes a first sun gear S1, a first ring gear R1, a plurality of inner pinion gears P1, and a plurality of outer pinion gears P2. The configuration includes the first carrier CA1.

  The first sun gear S1 is fixed to the case 1a and cannot be rotated, and the first ring gear R1 can be rotated integrally with the intermediate drum 33 via the third clutch C3 or in a relatively rotatable state. The first sun gear S1 is supported concentrically on the inner diameter side of the first ring gear R1.

  The plurality of inner pinion gears P1 and the plurality of outer pinion gears P2 are interposed at several circumferential positions in the opposed annular space between the first sun gear S1 and the first ring gear R1. P1 is meshed with the first sun gear S1, and the plurality of outer pinion gears P2 are meshed with the inner pinion gear P1 and the first ring gear R1.

  The first carrier CA1 rotatably supports both pinion gears P1, P2, and the central shaft portion of the first carrier CA1 is integrally connected to the input shaft 9, and both the pinion gears P1 in the first carrier CA1. , P2 are supported by the intermediate drum 33 via the fourth clutch C4 so as to be integrally rotatable or relatively rotatable.

  The intermediate drum 33 is rotatably disposed on the outer diameter side of the first ring gear R1, and is supported by the case 1a via the first brake B1 so as not to rotate or to be relatively rotatable. Yes.

  The rear planetary 32 is a gear-type planetary mechanism called a Ravinio type, and includes a large-diameter second sun gear S2, a small-diameter third sun gear S3, a second ring gear R2, a plurality of short pinion gears P3, The configuration includes a plurality of long pinion gears P4 and a second carrier CA2.

  The second sun gear S2 is connected to the intermediate drum 33, and the third sun gear S3 is connected to the first ring gear R1 of the front planetary 31 via the first clutch C1 so as to be integrally rotatable or relatively rotatable. The two ring gear R2 is integrally connected to the output shaft 10.

  The plurality of short pinion gears P3 are meshed with the third sun gear S3, and the plurality of long pinion gears P4 are meshed with the second sun gear S2 and the second ring gear R2 and are connected via the short pinion gear P3. 3 meshed with sun gear S3.

  Further, the second carrier CA2 rotatably supports a plurality of short pinion gears P3 and a plurality of long pinion gears P4, and a central shaft portion thereof is coupled to the input shaft 9 via the second clutch C2. The support shafts that support the pinion gears P3 and P4 in the second carrier CA2 are supported by the case 1a via the second brake B2 and the one-way clutch F1.

  The first to fourth clutches C1 to C4 and the first and second brakes B1 and B2 correspond to the hydraulic engagement elements recited in the claims, and here, wet using oil viscosity It is a multi-plate friction engagement device.

  The first clutch C1 is configured to bring the third sun gear S3 of the rear planetary 32 into an engaged state in which the third sun gear S3 can rotate integrally with the first ring gear R1 of the front planetary 31 or a released state in which relative rotation is possible.

  The second clutch C <b> 2 sets the second carrier CA <b> 2 of the rear planetary 32 in an engaged state in which the second carrier CA <b> 2 can rotate integrally with the input shaft 9 or in a released state in which relative rotation is possible.

  The third clutch C <b> 3 sets the first ring gear R <b> 1 of the front planetary 31 to an engaged state in which the first ring gear R <b> 1 can rotate integrally with the intermediate drum 33 or a released state in which relative rotation is possible.

  The fourth clutch C <b> 4 sets the first carrier CA <b> 1 of the front planetary 31 to an engaged state where the first carrier CA <b> 1 can rotate integrally with the intermediate drum 33 or a released state which allows relative rotation.

  The first brake B1 integrates the intermediate drum 33 with respect to the case 1a of the vehicle automatic transmission 1 so as to be in a non-rotatable engaged state or a relatively rotatable disengaged state.

  The second brake B <b> 2 integrates the second carrier CA <b> 2 of the rear planetary 32 with respect to the case 1 a so as to be in a non-rotatable engaged state or a relatively rotatable disengaged state.

  The one-way clutch F1 allows rotation of the rear planetary 32 in only one direction of the second carrier CA2.

  The hydraulic control device 4 controls the speed change operation of the speed change mechanism unit 3, and mainly includes a pressure control valve 41, a manual valve 42, and a plurality of linear solenoid valves SLC1, SLC2, SLC3, SLC4 as shown in FIG. The configuration includes SLB1, B2 control valve 44, cutoff valves 45, 46, 47 as fail-safe valves, switching valves 48, 49, and the like.

  The pressure control valve 41 controls the hydraulic pressure from the oil pump 6 to a predetermined line pressure and supplies it to the port PL of the manual valve 42.

  The manual valve 42 is appropriately connected to the linear solenoid valves SLC1, SLC2, SLC3, SLC4 from the port D in order to secure a neutral range N, forward travel range D or reverse travel range R corresponding to the operation of the shift lever by the driver. The hydraulic pressure is supplied to the SLB 1 and from the port R to the B2 control valve 44, respectively.

  The plurality of linear solenoid valves SLC1, SLC2, SLC3, SLC4, and SLB1 individually drive the first to fourth clutches C1 to C4 and the first brake B1 in the transmission mechanism unit 3, and the basic configuration thereof is a known configuration Therefore, detailed illustration and explanation are omitted here.

  In addition, the meaning of the code | symbol of linear solenoid valve SLC1, SLC2, SLC3, SLC4, SLB1 is a reference code which shows each hydraulic-type engagement element (1st-4th clutch C1-C4 and 1st brake B1) respectively corresponding. It is displayed after SL.

  A solenoid (not shown) of each of the linear solenoid valves SLC1, SLC2, SLC3, SLC4, and SLB1 operates in response to a control signal (duty control pulse) supplied from the transmission control device 5, and a valve body (not shown) Is moved to a position that balances with the spring force of the compression spring, and necessary ports are opened or closed or the opening degree is adjusted.

  The B2 control valve 44 drives the second brake B2.

  The first cut-off valve 45 is interposed between the first clutch C1 and the linear solenoid valve SLC1, and when the hydraulic pressure is supplied to the two input ports, the first cut-off valve 45 passes through the output port from the linear solenoid valve SLC1. Thus, the hydraulic pressure supplied to the first clutch C1 is cut off, and the fail-safe valve is discharged from the drain port into the case 1a.

  The second cutoff valve 46 is interposed between the fourth clutch C4 and the linear solenoid valve SLC4, and when the hydraulic pressure is supplied from the linear solenoid valve SLC3 to a single input port, the linear solenoid valve SLC4. Is configured as a fail-safe valve that shuts off the hydraulic pressure supplied to the fourth clutch C4 from the drain port and discharges it from the drain port into the case 1a.

  The third cutoff valve 47 is interposed between the first brake B1 and the linear solenoid valve SLB1, and when hydraulic pressure is supplied from the linear solenoid valve SLC3 or SLC4 to one of the two input ports. The hydraulic pressure supplied from the linear solenoid valve SLB1 to the first brake B1 via the output port is shut off, and the failsafe valve is discharged from the drain port into the case 1a.

  The switching valves 48 and 49 are disposed in series between the linear solenoid valve SLB1 and one input port of the first cut-off valve 45.

  The two input ports of the first switching valve 48 are connected in parallel with the hydraulic piping of the linear solenoid valve SLB1 and the hydraulic piping of the linear solenoid valve SLC4. The two input ports of the second switching valve 49 are connected in parallel with the output piping of the first switching valve 48 and the hydraulic piping of the linear solenoid valve SLC3. The first and second switching valves 48 and 49 are configured to output the supplied hydraulic pressure from the output port when hydraulic pressure is supplied to one of the input ports.

  The transmission control device 5 controls the hydraulic control device 4 to establish an appropriate shift stage, that is, a power transmission path in the transmission mechanism unit 3, and is generally a known ECU (Electronic Control Unit).

  That is, the transmission control device 5 includes a central processing unit (CPU) 51, a read only memory (ROM) 52, a random access memory (RAM) 53, a backup RAM 54, an input interface 55, as shown in FIG. The output interface 56 is connected to each other by a bidirectional bus 57.

  The engine control device 8 has the same hardware configuration as the transmission control device 5.

  The CPU 51 executes arithmetic processing based on various control programs and control maps stored in the ROM 52. The ROM 52 stores various control programs for controlling the speed change operation of the speed change mechanism unit 3 and the fail safe operation to which the features of the present invention are applied. The fail safe operation will be described in detail later. The RAM 53 is a memory that temporarily stores calculation results in the CPU 51, data input from each sensor, and the like. The backup RAM 54 is a non-volatile memory that stores various data to be saved.

  Connected to the input interface 55 are at least an engine speed sensor 11, an input shaft speed sensor 12, an output shaft speed sensor 13, a range position sensor 14, a throttle opening sensor 15, an output shaft torque sensor 16, and the like. . Further, at least the components of the hydraulic control device 4 (pressure control valve 41, manual valve 42, linear solenoid valves SLC1, SLC2, SLC3, SLC4, SLB1, B2 control valve 44) are connected to the output interface 56. .

  The engine speed sensor 11 detects the engine speed NE of the torque converter 2 to which the engine speed is transmitted. The input shaft rotational speed sensor 12 detects the rotational speed NT of the input shaft 9. The output shaft rotational speed sensor 13 detects the rotational speed NO of the output shaft 10. The range position sensor 14 sends out a detection signal when the manual valve 42 is shifted to the forward travel range D and the neutral range N. The throttle opening sensor 15 detects the amount of accelerator depression. The output shaft torque sensor 16 detects the torque of the output shaft 10.

  Here, the conditions for establishing each gear position in the above-described transmission mechanism unit 3 will be described with reference to FIGS.

  FIG. 6 is an engagement table showing the relationship between the engagement state or the disengagement state in the first to fourth clutches C1 to C4, the first and second brakes B1 and B2, and the one-way clutch F1, and the respective shift speeds. In this engagement table, ◯ indicates an “engaged state”, x indicates a “released state”, ◎ indicates an “engaged state during engine braking”, and Δ indicates an “engaged state only during driving”.

  FIGS. 7 to 14 show hydraulic paths for establishing the first speed to the eighth speed in the hydraulic control device 4, respectively. In these figures, a dot pattern is given to the hydraulic pressure supplied from the linear solenoid valve to the clutch or brake.

(First gear: 1st)
The first speed 1st is established by engagement of the first clutch C1 and automatic engagement of the one-way clutch F1.

  That is, as shown in FIG. 7, only the hydraulic path from the linear solenoid valve SLC1 to the first clutch C1 is secured and the first clutch C1 is engaged.

  In this case, the engagement of the first clutch C1 allows the first ring gear R1 of the front planetary 31 and the third sun gear S3 of the rear planetary 32 to be rotated together, and the one-way clutch F1 automatically engages the rear. The rotation of the second carrier CA2 of the planetary 32 is stopped.

  Accordingly, the third sun gear S3 rotated from the first carrier CA1 directly connected to the input shaft 9 via the first ring gear R1, the second carrier CA2 prevented from reverse rotation by the one-way clutch F1, and free rotation is possible. With the engagement with the second sun gear S2, the second ring gear R2 and the output shaft 10 are rotated at the gear ratio of the first gear.

(2nd speed: 2nd)
The second speed stage 2nd is established by engagement of the first clutch C1 and the first brake B1.

  That is, as shown in FIG. 8, a hydraulic path from the linear solenoid valve SLC1 to the first clutch C1 is secured to engage the first clutch C1, and a hydraulic path from the linear solenoid valve SLB1 to the first brake B1 is set. Ensuring and engaging the first brake B1.

  In this case, first, the engagement of the first clutch C1 allows the first ring gear R1 of the front planetary 31 and the third sun gear S3 of the rear planetary 32 to rotate together, and the engagement of the first brake B1. As a result, the intermediate drum 33 and the second sun gear S2 of the rear planetary 32 are fixed to the case 1a so that they cannot rotate.

  Accordingly, the third sun gear S3 rotated from the first carrier CA1 directly connected to the input shaft 9 via the first ring gear R1, the second sun gear S2 made non-rotatable, and the second carrier capable of free rotation. Along with the meshing with CA2, the second ring gear R2 and the output shaft 10 are rotated at the gear ratio of the second speed stage.

(3rd speed: 3rd)
The third speed stage 3rd is established by engagement of the first clutch C1 and the third clutch C3.

  That is, as shown in FIG. 9, a hydraulic path from the linear solenoid valve SLC1 to the first clutch C1 is secured to engage the first clutch C1, and a hydraulic path from the linear solenoid valve SLC3 to the third clutch C3 is set. Ensuring and engaging the third clutch C3.

  In this case, first, the engagement of the first clutch C1 allows the first ring gear R1 of the front planetary 31 and the third sun gear S3 of the rear planetary 32 to rotate together, and the engagement of the third clutch C3. As a result, the first ring gear R1 of the front planetary 31 and the intermediate drum 33 and the second sun gear S2 of the rear planetary 32 can be rotated together.

  Accordingly, the second sun gear S2 and the third sun gear S3 rotated from the first carrier CA1 directly connected to the input shaft 9 via the first ring gear R1 and the intermediate drum 33, and the second carrier CA2 capable of free rotation. Along with the meshing, the second ring gear R2 and the output shaft 10 are rotated at the gear ratio of the third speed stage.

(4th speed: 4th)
The fourth speed stage 4th is established by engagement of the first clutch C1 and the fourth clutch C4.

  That is, as shown in FIG. 10, the hydraulic path from the linear solenoid valve SLC1 to the first clutch C1 is secured to engage the first clutch C1, and the hydraulic path from the linear solenoid valve SLC4 to the fourth clutch C4 is set. Ensuring and engaging the fourth clutch C4.

  In this case, first, the engagement of the first clutch C1 allows the first ring gear R1 of the front planetary 31 and the third sun gear S3 of the rear planetary 32 to rotate together, and the engagement of the fourth clutch C4. Thus, the first carrier CA1 of the front planetary 31 and the intermediate drum 33 and the second sun gear S2 of the rear planetary 32 can be rotated together.

  Accordingly, the second sun gear S2 rotated through the first carrier CA1 and the intermediate drum 33 directly connected to the input shaft 9 and the first carrier CA1 directly connected to the input shaft 9 are rotated through the first ring gear R1. The second ring gear R2 and the output shaft 10 are rotated at a gear ratio of the fourth speed stage in accordance with the meshing of the third sun gear S3 and the second carrier CA2 that can freely rotate.

(5th gear: 5th)
The fifth speed stage 5th is established by engagement of the first clutch C1 and the second clutch C2.

  That is, as shown in FIG. 11, a hydraulic path from the linear solenoid valve SLC1 to the first clutch C1 is secured to engage the first clutch C1, and a hydraulic path from the linear solenoid valve SLC2 to the second clutch C2 is set. Ensuring and engaging the second clutch C2.

  In this case, first, the engagement of the first clutch C1 allows the first ring gear R1 of the front planetary 31 and the third sun gear S3 of the rear planetary 32 to rotate together, and the engagement of the second clutch C2. Thus, the input shaft 9 and the second carrier CA2 of the rear planetary 32 can be rotated together.

  As a result, the third sun gear S3 rotated from the first carrier CA1 directly connected to the input shaft 9 via the first ring gear R1, the second sun gear S2 that can rotate freely, and the second sun gear S2 that rotates integrally with the input shaft 9. Along with the meshing with the carrier CA2, the second ring gear R2 and the output shaft 10 are rotated at the gear ratio of the fifth speed stage.

(6th speed: 6th)
The sixth speed stage 6th is established by engagement of the second clutch C2 and the fourth clutch C4.

  That is, as shown in FIG. 12, a hydraulic path from the linear solenoid valve SLC2 to the second clutch C2 is ensured to engage the second clutch C2, and a hydraulic path from the linear solenoid valve SLC4 to the fourth clutch C4 is set. Ensuring and engaging the fourth clutch C4.

  In this case, first, the input shaft 9 and the second carrier CA2 of the rear planetary 32 can be rotated together by the engagement of the second clutch C2, and the front planetary 31 is engaged by the engagement of the fourth clutch C4. The first carrier CA 1, the intermediate drum 33, and the second sun gear S 2 of the rear planetary 32 can be rotated together.

  Accordingly, the second sun gear S2 rotated through the first carrier CA1 and the intermediate drum 33 directly connected to the input shaft 9, the second carrier CA2 rotated integrally with the input shaft 9, and the third sun gear S3 capable of free rotation. , The second ring gear R2 and the output shaft 10 are rotated at the sixth gear ratio.

(7th gear: 7th)
The seventh speed stage 7th is established by engagement of the second clutch C2 and the third clutch C3.

  That is, as shown in FIG. 13, a hydraulic path from the linear solenoid valve SLC2 to the second clutch C2 is secured to engage the second clutch C2, and a hydraulic path from the linear solenoid valve SLC3 to the third clutch C3 is set. Ensuring and engaging the third clutch C3.

  In this case, first, the input shaft 9 and the second carrier CA2 of the rear planetary 32 can be integrally rotated by the engagement of the second clutch C2, and the front planetary 31 is engaged by the engagement of the third clutch C3. The first ring gear R 1, the intermediate drum 33, and the second sun gear S 2 of the rear planetary 32 can be rotated together.

  Thus, the second sun gear S2 rotated from the first carrier CA1 directly connected to the input shaft 9 via the first ring gear R1 and the intermediate drum 33, the second carrier CA2 rotated integrally with the input shaft 9, and the free Along with the meshing with the rotating third sun gear S3, the second ring gear R2 and the output shaft 10 are rotated at the gear ratio of the seventh speed stage.

(8th gear: 8th)
The eighth speed stage 8th is established by engagement of the second clutch C2 and the first brake B1.

  That is, as shown in FIG. 14, the hydraulic path from the linear solenoid valve SLC2 to the second clutch C2 is secured and the second clutch C2 is engaged, and the hydraulic path from the linear solenoid valve SLB1 to the first brake B1 is set. Ensuring and engaging the first brake B1.

  In this case, first, the input shaft 9 and the second carrier CA2 of the rear planetary 32 can be rotated together by the engagement of the second clutch C2, and the intermediate drum 33 and the second brake CA are engaged by the engagement of the first brake B1. The second sun gear S2 of the rear planetary 32 is fixed to the case 1a and cannot rotate.

  As a result, the second carrier CA2 that rotates integrally with the input shaft 9, the intermediate drum 33 and the second sun gear S2 that are made non-rotatable, and the third sun gear S3 that is free to rotate are engaged with the second ring. The gear R2 and the output shaft 10 are rotated at the gear ratio of the eighth speed stage.

(Reverse first speed: R1)
The reverse first speed R1 is established by engagement of the third clutch C3 and the second brake B2.

  That is, although not shown, a hydraulic path from the linear solenoid valve SLC3 to the third clutch C3 is secured to engage the third clutch C3, and a hydraulic path from the B2 control valve 44 to the second brake B2 is secured. Then, the second brake B2 is engaged.

  In this case, first, the engagement of the third clutch C3 allows the first ring gear R1 of the front planetary 31 and the intermediate drum 33 and the second sun gear S2 of the rear planetary 32 to be integrally rotated. Due to the engagement of the brake B2, the second carrier CA2 of the rear planetary 32 is fixed to the case 1a and cannot rotate.

  Thus, the second sun gear S2 rotated from the first carrier CA1 directly connected to the input shaft 9 via the first ring gear R1 and the intermediate drum 33, the second carrier CA2 made non-rotatable, and free rotation With the engagement with the third sun gear S3, the second ring gear R2 and the output shaft 10 are rotated in reverse at the gear ratio of the reverse first speed.

(Reverse second speed: R2)
The second reverse speed R2 is established by engagement of the fourth clutch C4 and the second brake B2.

  That is, although not shown, a hydraulic path from the linear solenoid valve SLC4 to the fourth clutch C4 is secured to engage the fourth clutch C4, and a hydraulic path from the B2 control valve 44 to the second brake B2 is secured. Then, the second brake B2 is engaged.

  In this case, first, engagement of the fourth clutch C4 allows the first carrier CA1 of the front planetary 31 and the intermediate drum 33 and the second sun gear S2 of the rear planetary 32 to be integrally rotated, and the second brake. Due to the engagement of B2, the second carrier CA2 of the rear planetary 32 is fixed to the case 1a so that it cannot rotate.

  Thus, the second sun gear S2 rotated through the first carrier CA1 and the intermediate drum 33 directly connected to the input shaft 9, the second carrier CA2 made non-rotatable, and the third sun gear S3 that is free to rotate. As a result, the second ring gear R2 and the output shaft 10 are rotated in reverse at the gear ratio of the second reverse speed.

  Next, portions to which the features of the present invention are applied will be described in detail with reference to FIGS. 15 and 16.

  In the first place, in general, when engaging or disengaging the clutches C1 to C4 and the brake B1 that are necessary at the time of shifting such as upshifting and downshifting, each clutch C1 to C1 By changing the opening of linear solenoid valves SLC1 to SLC4 and SLB1 for driving C4 and brake B1, that is, the hydraulic pressure applied to each clutch C1 to C4 and brake B1 so as to gradually increase or decrease in an appropriate pattern, The inertia phase is controlled.

  Note that the inertia phase means, for example, as shown in FIG. 16A, the rotational speed NT of the input shaft 9 (turbine shaft) is the same as the synchronous rotational speed of the gear stage before the shift (first speed stage). It is the section BL that changes according to the deviation from the synchronous rotation speed of the stage (second speed stage). The rotational speed of the input shaft 9 in the inertia phase decreases during an upshift and increases during a downshift. The synchronous rotation speed is the rotation speed NT of the input shaft 9 corresponding to the gear ratio of the gear position. Further, the rotational speed NT of the input shaft 9 is directly detected by the input shaft rotational speed sensor 12, and this detected value is input to the transmission control device 5.

  Here, in this embodiment, the change pattern of the hydraulic pressure command value, that is, the initial value of the hydraulic pressure command pattern is empirically defined in advance, and the shift command is received every upshift associated with vehicle travel. Measures the elapsed time of the appropriate section from the start of inertia phase to the end of the inertia phase, and based on this measurement result, it is devised to execute a correction process to deform the previous hydraulic pressure command pattern to a new hydraulic pressure command pattern .

  In this embodiment, the hydraulic pressure command pattern correction process is performed in three stages.

  As a first stage, the constant pressure standby pressure in the hydraulic pressure command pattern is corrected to be high or low. As a second stage, the constant pressure standby pressure in the hydraulic pressure command pattern is gradually increased from the start point of the inertia phase, and the pressure increase gradient is corrected in magnitude. As a third stage, the pressure reduction gradient is gradually reduced at the end of the constant pressure standby pressure in the hydraulic pressure command pattern, and the pressure reduction gradient is corrected in magnitude.

  Hereinafter, the learning operation of the hydraulic pressure command pattern during the upshift by the transmission control device 5 to which the features of the present invention are applied will be specifically described with reference to the flowchart shown in FIG. 15 and the time charts shown in FIG. 16 and FIG. Explained.

  Here, as an example of the upshift, a case where a shift process from the first speed 1st to the second speed 2nd is performed is taken as an example.

  In the first place, as shown in FIG. 7, the condition for establishing the first speed 1st is to automatically engage the one-way clutch F1 by engaging the first clutch C1, and the condition for establishing the second speed 2nd. Is engaging the first clutch C1 and the first brake B1, as shown in FIG.

  Therefore, as a procedure for upshifting from the first speed 1st to the second speed 2nd, first, the first clutch C1 is kept engaged and the first brake B1 is engaged. That's fine.

  The initial value P of the hydraulic pressure command pattern for the linear solenoid valve SLB1 when the first brake B1 is engaged is, for example, a pattern as shown by the solid line in FIG. The initial value P of the hydraulic pressure command pattern is empirically defined in advance, and is stored in the backup RAM 54 (corresponding to storage means described in claims).

  When an upshift command is received at time t1 in FIG. 16, the process enters the flowchart shown in FIG.

  In the flowchart of FIG. 15, steps S11 to S15 are the first stage, steps S21 to S25 are the second stage, and steps S31 to S35 are the third stage.

  First, when a shift command is received, quick apply control is performed at a time t2 after the elapse of a predetermined time from the time t1 when the shift command is received, and a predetermined level of hydraulic pressure is applied to the first brake B1 with the linear solenoid valve SLB1 as a predetermined opening. And improve responsiveness. Thereafter, by controlling the inertia phase, the hydraulic pressure is appropriately supplied from the fully closed linear solenoid valve SLB1 to the first brake B1, and the first brake B1 is engaged. As a result, as shown in FIG. 8, the first clutch C1 and the first brake B1 are engaged and the second speed 2nd is established, as indicated by the dot pattern and the hatched lines in FIG. become.

  Here, regarding the correction of the hydraulic pressure command pattern, first, in step S11, the first timer in the transmission control device 5 performs the first operation from the time point t1 when the gear shift command is received in FIG. 16 to the time point t4 when the inertia phase starts. The elapsed time T1 of one section is measured.

  Note that the inertia phase start time t4 is not the position t3 at which the rotational speed NT of the input shaft 9 actually starts to decrease. This is the time when the number Δr drops.

  In subsequent step S12, a deviation Δta between the measured value T1 measured in step S11 and the first target value A is calculated, and it is determined whether or not the calculated result Δta is within a predetermined error ± α. . Here, when an affirmative determination is made, the first stage is canceled and the process proceeds to the second stage of the following steps S21 to S25. When a negative determination is made, the first stage is related at steps S13 to S15. The process shifts to the process of deforming the reference (or previous) hydraulic command pattern indicated by the solid line in FIG.

Here, first, in step S13, a substantial deviation is calculated by subtracting the absolute value α of the error ± α from the deviation Δta calculated in step S12, and in the subsequent step S14, the substantial deviation calculated in step S13 is calculated. The deviation is multiplied by an appropriate gain G1, and the result calculated in step S14 in the subsequent step S15 is added to or subtracted from the constant pressure standby pressure P _L in the reference (or previous) hydraulic pressure command pattern P, as shown in FIG. The height is corrected as indicated by a two-dot chain line, and then the hydraulic pressure command pattern in the backup RAM 54 is updated with the deformed new hydraulic pressure command pattern as the next hydraulic pressure command pattern.

For reference, for example, as shown in FIG. 17A, only the level of the constant pressure standby pressure P _L in the reference hydraulic pressure command pattern P shown by the solid line in FIG. 16C can be increased. Command pattern P1 is secured.

Here, if correction is made to increase the constant pressure standby pressure P _L in the reference (or previous) hydraulic pressure command pattern P, the first time from the time point t1 when the shift command is received to the time point t4 when the inertia phase starts. As the elapsed time T1 of the section becomes shorter, the increase in the rotational speed NT of the input shaft 9 at the inertia phase start time t4 tends to be suppressed. If the increase in the rotational speed NT of the input shaft 9 at the start point t4 of the inertia phase can be suppressed in this way, the start position where the torque of the output shaft 10 drops as shown in FIG. Can be shortened.

On the other hand, when the constant pressure standby pressure P _L in the reference (or previous) hydraulic pressure command pattern P is corrected to decrease, the elapsed time T1 of the first section becomes longer and the input at the start time t4 of the inertia phase. The rotational speed NT of the shaft 9 tends to increase.

  Thereafter, the process proceeds to the second stage. In this second stage, first, in step S21, when the rotational speed NT of the input shaft 9 reaches the predetermined rotational speed NT1 from the inertia phase start time t4 in FIG. 16 by the second timer inside the transmission control device 5. The elapsed time T2 of the second section up to t5 is measured. The predetermined rotational speed NT1 is arbitrarily set.

  In subsequent step S22, a deviation Δtb between the measured value T2 measured in step S21 and the second target value B is calculated, and it is determined whether or not the calculated result Δtb is within a predetermined error ± α. . Here, if an affirmative determination is made, the second stage is canceled and the process proceeds to the third stage of steps S31 to S35 described below. If a negative determination is made, the second stage is related to steps S23 to S25. The process proceeds to a process for deforming the hydraulic command pattern.

Here, first, in step S23, a substantial deviation is calculated by subtracting the absolute value α of the error ± α from the deviation Δtb calculated in step S22, and in the subsequent step S24, the substantial deviation calculated in step S23 is calculated. Based on the result calculated in step S24 in the subsequent step S25 by multiplying the deviation by an appropriate gain G2, the constant pressure in the reference (or previous) hydraulic pressure command pattern P as shown by the two-dot chain line in FIG. The standby pressure P _L is gradually increased from the position P _L1 corresponding to the start point t4 of the inertia phase, and the pressure increase gradient β is corrected in magnitude. Thereafter, the deformed new hydraulic command pattern is used as the next hydraulic command pattern as a backup RAM 54. The hydraulic command pattern in the inside is updated.

For reference, for example, as shown in (b) of FIG. 17, the hydraulic pressure command pattern P1 shown in FIG. 17 (a) corrected by the first step, the start time t4 of the inertia phase of the pressure standby pressure P _L It can be gradually increased from the corresponding position P _L1 , and a new hydraulic pressure command pattern P2 can be secured. Note that the reduced pressure gradient γ2 of the new hydraulic pressure command pattern P2 remains the same as the reduced pressure gradient γ1 of the hydraulic pressure command pattern P1 shown in FIG.

Here, if the constant pressure standby pressure P _L in the reference (or previous) hydraulic pressure command pattern P is corrected so as to gradually increase from the position P _L1 corresponding to the inertia phase start time t4, the input is made from the inertia phase start time t4. The elapsed time T2 of the second section until the rotational speed NT of the shaft 9 reaches the predetermined rotational speed NT1 is shortened. Incidentally, when the pressure increase gradient β is increased, the elapsed time T2 of the second section tends to be shortened, while when the pressure increase gradient β is decreased, the elapsed time T2 of the second section tends to be long.

  In any case, the correction in the second stage makes it possible to promote the decrease in the rotational speed NT of the input shaft 9 from the inertia phase start time t4 after the shift command, and to shorten the shift time. This is advantageous.

  Thereafter, the process proceeds to the third stage. In this third stage, first, in step S31, the third timer in the transmission control device 5 causes the gear shift stage where the inertia phase start time t4 to end time t6 in FIG. The elapsed time T3 of the third section until the synchronous rotational speed is reached is measured.

  Note that the end point t6 of the inertia phase is not the position t6 at which the rotational speed NT of the input shaft 9 actually reaches the synchronous rotational speed of the requested shift stage, and the rotational speed NT of the input shaft 9 is required for control purposes. The time point is higher by a predetermined speed Δr than the time point t7 when the synchronous speed of the gear stage is actually reached.

  In subsequent step S32, a deviation Δtc between the measured value T3 measured in step S31 and the third target value C is calculated, and it is determined whether or not the calculated result Δtc is within a predetermined error ± α. . Here, if an affirmative determination is made, the third stage is canceled and the process returns to the main flow chart of the speed change process, but if a negative determination is made, the hydraulic pressure command pattern related to the third stage is modified in steps S33 to S35. Transition to processing.

Here, first, in step S33, a substantial deviation is calculated by subtracting the absolute value α of the error ± α from the deviation Δtc calculated in step S32. Then, in step S34, the substantial deviation calculated in step S33 is calculated. Based on the result calculated in step S34 in the subsequent step S35 by multiplying the deviation by an appropriate gain G3, the constant pressure in the reference (or previous) hydraulic command pattern P as shown by the two-dot chain line in FIG. The standby pressure P _L is gradually decreased from the end time t6 and the pressure reduction gradient γ is corrected in magnitude, and then the hydraulic command pattern in the backup RAM 54 is updated with the deformed new hydraulic command pattern as the next hydraulic command pattern.

For reference, for example, as shown in FIG. 17 (c), the position corresponding to the end point t6 of the constant pressure standby pressure P _L for the hydraulic pressure command pattern P2 shown in FIG. 17 (b) corrected in the second stage. Thus, a new hydraulic pressure command pattern P3 can be secured. The reduced pressure gradient γ3 of the new hydraulic pressure command pattern P3 is made smaller than the reduced pressure gradient γ2 of the hydraulic pressure command pattern P2 shown in FIG.

Here, if it is corrected so as to gradually decrease from the end time t6 of the constant pressure standby pressure P _L in the reference (or previous) oil pressure command pattern P, as shown in FIG. 16B, at the end time t6 of the inertia phase. The output shaft torque is gradually reduced. This output shaft torque can be recognized based on the detection output from the output shaft torque sensor 16.

  Incidentally, when the decompression gradient γ is increased, the elapsed time T3 of the third section from the start time t4 of the inertia phase to the end time t6 tends to be longer, whereas when the decompression gradient γ is decreased, the passage of the third section. Time T3 tends to be shorter.

  In any case, the correction in the third stage reduces the hydraulic pressure when the required brake B1 (or clutch) is engaged, which is advantageous in reducing the engagement shock of the brake B1.

  By the way, the gains G1 to G3 used in each of the above steps are specific values necessary for converging the shift time T as quickly as possible, and are specified as values obtained through experiments.

  The steps S11, S21, and S31 are the measuring means according to claim 6. Specifically, step S11 is the first measuring means according to claim 7, and step S21 is the second measuring means according to claim 7. Further, step S31 corresponds to the third measuring means described in claim 7, respectively. The steps S13 to S15 correspond to the first correcting means described in claim 7, the steps S23 to S25 correspond to the second correcting means described in claim 7, and the steps S33 to S35 are claimed. This corresponds to the third correction unit according to Item 7. Steps S15, S25, and S35 also serve as learning means according to claim 6.

  Such corrections in the first to third stages are repeatedly performed every time the same upshift is performed.

  As described above, according to the embodiment to which the features of the present invention are applied, learning is performed by learning the change pattern of the actual rotational speed NT of the input shaft 9 for the hydraulic command pattern in the inertia phase control for each upshift. Since the result is fed back and the previous hydraulic command pattern is modified and updated, it becomes possible to create a hydraulic command pattern adapted to each driver's driving and driving situation as the driving progresses. Ideal inertia phase control becomes possible.

  As a result, it is possible to achieve both the shortening of the shift time T and the reduction of the engagement shock of the hydraulic engagement element in a well-balanced manner, so that the driver does not need to feel extended during the upshift or the shift shock. become.

  By the way, in the above-described embodiment, the case of shifting from the first speed 1st to the second speed 2nd is described as an example, but detailed description will be given even when upshifting to other speeds. Although omitted, it is possible to deal with the same as above.

  In addition, this invention is not limited only to the said embodiment, Various application and deformation | transformation can be considered.

  In the above embodiment, an example is given in which the present invention is applied to the vehicle automatic transmission 1 set to 8 forward speeds and 2 reverse speeds, but the present invention is also applied to other vehicle automatic transmissions. Can do.

  For example, for the speed change mechanism 3, at least two planetary members 31 and 32, an intermediate drum 33 (intermediate rotating element) that connects the components of the planetary members 31 and 32 so that power can be transmitted, and a hydraulic engagement element Although the first to fourth clutches C1 to C4 and the first and second brakes B1 and B2 are provided, the configuration of the transmission mechanism unit 3 can be changed as appropriate.

  The configuration of the hydraulic control device 4 that hydraulically controls the operation of each hydraulic engagement element can be changed as appropriate in accordance with the configuration change of the speed change mechanism portion 3.

It is a figure which shows schematic structure of the powertrain of the vehicle using the automatic transmission for vehicles used as this invention. It is a skeleton figure of the automatic transmission for vehicles of FIG. FIG. 3 is a perspective view schematically showing a transmission mechanism unit in FIG. 2. It is a figure which shows the structure of the hydraulic control apparatus of FIG. It is a figure which shows the structure of the transmission control apparatus of FIG. FIG. 3 is an engagement table for each gear position of each clutch and each brake in the speed change mechanism portion of FIG. FIG. 5 is a diagram corresponding to the hydraulic control device of FIG. 4 and shows a hydraulic path for establishing the first speed stage. It is a figure corresponding to the hydraulic control apparatus of FIG. 4, and shows the hydraulic path for establishing the second gear. FIG. 5 is a diagram corresponding to the hydraulic control device of FIG. 4 and shows a hydraulic path for establishing the third speed stage. It is a figure corresponding to the hydraulic control apparatus of FIG. 4, and shows the hydraulic path for establishing the fourth speed. It is a figure corresponding to the hydraulic control apparatus of FIG. 4, and shows the hydraulic path for establishing the fifth gear. It is a figure corresponding to the hydraulic control apparatus of FIG. 4, and shows the hydraulic path for establishing the sixth gear. FIG. 9 is a diagram corresponding to the hydraulic control device of FIG. 4 and shows a hydraulic path for establishing the seventh speed stage. It is a figure corresponding to the hydraulic control apparatus of FIG. 4, and shows the hydraulic path for establishing the eighth gear. 6 is a flowchart for explaining operations related to detection and handling of a shift abnormality by the transmission control device of FIG. 5. FIG. 3 is a time chart related to the operation explanation when upshifting from the first gear to the second gear in the speed change mechanism shown in FIG. 2, (a) is the input shaft rotation speed, and (b) is the output shaft. (C) shows the opening of the linear solenoid valve, that is, the hydraulic pressure applied to the corresponding brake or clutch. 16A is a time chart corresponding to FIG. 16C, where FIG. 16A shows the first stage of correction of the hydraulic pressure command pattern, FIG. 16B shows the second stage of correction of the hydraulic pressure command pattern, and FIG. Each of the third stages of correction is shown.

Explanation of symbols

1 Automatic transmission for vehicles
3 Transmission mechanism
31 Front planetary
32 Rear planetary
33 Intermediate drum
C1 first clutch (hydraulic engagement element)
C2 Second clutch (hydraulic engagement element)
C3 3rd clutch (hydraulic engagement element)
C4 4th clutch (hydraulic engagement element)
B1 First brake (hydraulic engagement element)
B2 Second brake (hydraulic engagement element)
F1 one-way clutch
4 Hydraulic control device
41 Pressure control valve
42 Manual valve
SLC1 ~ SLC4 Linear solenoid valves for the 1st to 4th clutches
SLB1 Linear solenoid valve for the first brake
44 B2 control valve
5 Transmission control device
6 Oil pump
7 engine
9 Input shaft
10 Output shaft
11 Engine speed sensor
12 Input shaft speed sensor
13 Output shaft speed sensor
14 Range position sensor
15 Throttle opening sensor
16 Output shaft torque sensor

Claims (7)

A transmission mechanism having a plurality of hydraulic engagement elements for establishing a power transmission path of each of the planetary mechanisms corresponding to a required gear stage, and at least two planetary mechanisms installed from the input shaft to the output shaft; A control device for an automatic transmission for a vehicle, including a hydraulic control device for engaging or releasing each hydraulic engagement element as required,
A shift process for engaging or releasing a hydraulic engagement element necessary for inertia phase control at the time of shift based on the previous hydraulic command pattern;
In the inertia phase control at the time of upshifting, the elapsed time of a plurality of sections from when a shift command is received until the inertia phase is completed is measured, and based on this measurement result, the previous hydraulic command pattern is transformed into a new hydraulic command pattern. A control device for an automatic transmission for a vehicle, wherein the correction processing is executed. The control device for an automatic transmission for a vehicle according to claim 1,
The hydraulic engagement element includes a wet friction brake that makes a suitable component in each planetary mechanism rotatable or non-rotatable, and a state in which two components of each planetary mechanism can rotate integrally. Or it has a wet friction clutch made into the state which can be rotated relatively, The control apparatus of the automatic transmission for vehicles characterized by the above-mentioned. The control device for an automatic transmission for a vehicle according to claim 1 or 2,
In the correction process, a deviation between the measured value of the elapsed time of the first section from when the shift command is received in the inertia phase control to the start of the inertia phase and the target value is calculated, and the hydraulic command pattern is calculated based on the calculation result. A control device for an automatic transmission for a vehicle, wherein the constant pressure standby pressure is corrected in a direction to increase or decrease. The control device for an automatic transmission for a vehicle according to claim 1 or 2,
In the correction process, in the inertia phase control, the deviation between the measured value and the target value of the elapsed time from the start of the inertia phase until the input shaft rotation speed reaches the predetermined rotation speed is calculated. A control device for an automatic transmission for a vehicle, wherein the constant pressure standby pressure in the hydraulic pressure command pattern is gradually increased from the start point of the inertia phase and the pressure increasing gradient is corrected to increase or decrease. The control device for an automatic transmission for a vehicle according to claim 1 or 2,
In the correction process, the deviation between the measured value of the elapsed time of the third section from the start of the inertia phase to the establishment of the requested gear stage and the target value in the inertia phase control is calculated, and the hydraulic pressure is calculated based on the calculation result. A control device for an automatic transmission for a vehicle, wherein the controller gradually reduces the constant pressure standby pressure in the command pattern at the end stage and corrects the pressure reduction gradient in a direction of increasing or decreasing. A transmission mechanism having a plurality of hydraulic engagement elements for establishing a power transmission path of each of the planetary mechanisms corresponding to a required gear stage, and at least two planetary mechanisms installed from the input shaft to the output shaft; A control device for an automatic transmission for a vehicle, including a hydraulic control device for engaging or releasing each hydraulic engagement element as required,
Storage means for storing a hydraulic command pattern for each shift stage and an appropriate target value;
An input shaft rotational speed detection means for detecting the rotational speed of the input shaft;
Measuring means for measuring the elapsed time of a plurality of sections from the receipt of the shift command to the end of the inertia phase;
A correction means for calculating a deviation between the measured value and the target value for each section, and deforming a predetermined portion of the previous hydraulic pressure command pattern based on the calculation result;
A control device for an automatic transmission for a vehicle, comprising: learning means for updating the hydraulic pressure command pattern in the storage means by using the deformed new hydraulic pressure command pattern as a next hydraulic pressure command pattern. The control device for an automatic transmission for a vehicle according to claim 6,
As the measuring means, a first measuring means for measuring an elapsed time of the first section from the start of the inertia phase after receiving the shift command in the inertia phase control;
A second measuring means for measuring an elapsed time of the second section from the start of the inertia phase until the input shaft rotational speed reaches a predetermined rotational speed in the inertia phase control;
A third measuring means for measuring an elapsed time of the third section from the start of the inertia phase to the establishment of the requested shift stage in the inertia phase control;
The correction means calculates a deviation between the measurement value obtained by the first measurement means and the corresponding first target value, and based on the result obtained by multiplying the calculated deviation by an appropriate gain, the previous hydraulic pressure command pattern First correction means for correcting the level of the constant pressure standby pressure at
The deviation between the measured value of the second measuring means and the corresponding second target value is calculated, and the constant pressure standby pressure in the previous hydraulic pressure command pattern is calculated based on the result obtained by multiplying the calculated deviation by an appropriate gain. A second correction means that gradually increases from the start of the phase and corrects the magnitude of the pressure increase gradient;
The deviation between the measured value of the third measuring means and the corresponding third target value is calculated, and the end of the constant pressure standby pressure in the previous hydraulic pressure command pattern is based on the result of multiplying the calculated deviation by an appropriate gain. A control device for an automatic transmission for a vehicle, comprising: third correction means for gradually reducing the pressure reduction gradient in stages and correcting the magnitude of the pressure reduction gradient.

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