JP2006316851A - Friction element slip preventing device for hybrid transmission - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To prevent the slip of a friction element during the transition of accelerating operation by controlling power to a hybrid transmission without depending on an increase in line pressure. <P>SOLUTION: To prevent the slip of a brake fastened in a low speed ratio fixing mode, the target motor torque of a motor/generator relating to a fixed rotating element is corrected in the following steps: a step S11 of finding required brake torque Tb, a step S12 of computing brake required hydraulic pressure to actualize the Tb and calculating required minimum target line pressure tP, a step S13 of computing actual brake fastening torque capacity obtained by actual line pressure P considering the response delay of line pressure control in accordance with the tP, a step S14 of finding a deviation ΔTb=Tb-Tbo, a step S15 of determining whether Tb>Tbo or Tb≤Tbo, a step S16 of setting ΔTb as a motor torque correcting amount opposite to input torque in the case of Tb>Tbo to prevent the slip of the brake, and a step S17 of correcting the torque correcting amount (=ΔTb only) of the motor/generator. <P>COPYRIGHT: (C)2007,JPO&INPIT

Description

  The present invention includes a motor / generator as a prime mover in addition to the main power source, and transmits the power from the main power source and the motor / generator by the transmission mechanism and outputs the transmission state of the friction element. The present invention relates to a technique for preventing slipping of a friction element in a hybrid transmission determined by fastening and releasing.

As a hybrid transmission, a series type in which one motor / generator driven by a main power source (usually an engine) drives the other motor / generator with electric power obtained by generating electric power and determines the output to the drive system. Hybrid transmissions,
A parallel hybrid transmission that determines the output to the drive system based on the power from the main power source and the power from at least one motor / generator,
There are hybrid transmissions that can selectively use both types.

  In any case, the hybrid transmission, as described in Patent Document 1, transmits the power from the main power source and the motor / generator constituting the prime mover by the speed change mechanism, and outputs the power from the speed change mechanism. Is often determined by selective engagement of friction elements such as clutches and brakes.

By the way, when engaging frictional elements such as clutches and brakes, the line pressure obtained by regulating the hydraulic oil from the oil pump is used as the original pressure, and this is reduced to the operating pressure of the frictional element. Execute the conclusion of.
For this reason, the fastening force of the friction element is uniquely determined by the operating pressure applied to the friction element.
Therefore, when determining the line pressure, which is the original pressure, it is necessary to set the line pressure so that the required operating pressure can be generated based on the required operating pressure corresponding to the required fastening force of the friction element.

If the line pressure does not meet the hydraulic pressure required to create the required operating pressure according to the required fastening force of the friction element, the fastening force of the friction element will be insufficient with respect to the required value, causing slipping. Not only can torque not be transmitted, but the durability of the friction element is reduced due to slipping.
On the other hand, if the line pressure greatly exceeds the hydraulic pressure required to create the required operating pressure according to the required fastening force of the friction element, the fastening force of the friction element becomes excessive with respect to the required value, Driving energy is consumed more than necessary, leading to deterioration of fuel consumption due to energy loss.

Therefore, as is done in automatic transmissions, it is common sense to control the line pressure so that it is the minimum hydraulic pressure necessary to produce the required operating pressure according to the required fastening force of the friction element. .
JP 2004-150627 A

However, if the line pressure is controlled to be the minimum required hydraulic pressure corresponding to the required operating pressure of the friction element,
Although it is not a problem in steady state when the driver is driving with the accelerator pedal depressed almost the same,
A transient in which the torque from the prime mover consisting of the main power source and the motor / generator changes suddenly, such as when the driver depresses the accelerator pedal and accelerates or when the engine brake travels (decelerates) with the accelerator pedal depressed. Sometimes the following problems occur.

At the time of acceleration / deceleration as described above, the required fastening force of the friction element changes accordingly, so that the line pressure is also changed in response to the change in the friction element required operating pressure.
However, since the line pressure control response speed is slower than the occurrence of torque change (change in required fastening force of the friction element) due to acceleration / deceleration, the line pressure is controlled during the delay time of the former line pressure control response to the latter torque change. However, there is a concern that the operating pressure of the friction element is made less than the required operating pressure, and the friction element temporarily slips to cause the above problem.

  As a countermeasure to solve this problem, it is conceivable to reduce the torque to the slipping friction element so that the slip does not occur. However, in this case, the output torque of the hybrid transmission is also lowered, and temporarily Another problem that the driving performance is affected by a decrease in the driving force of the vehicle is not a radical solution.

As another solution, the line pressure is always higher than the above-mentioned determined value so that the required operating pressure corresponding to the required fastening force of the friction element can be created, including during the above-mentioned transition, and the hydraulic margin is increased. It is possible to enlarge it.
However, in this case, the line pressure will always be higher than necessary during steady state, which is much longer than that at the time of transition, and the energy to drive the oil pump will be consumed more than necessary, resulting in deterioration of fuel consumption due to energy loss. This also cannot be a drastic solution.

The present invention follows a technique for reducing the torque applied to the slipping friction element so that the slip does not occur. In realizing this technique, in particular, the hybrid transmission is controlled by the torque control of the main power source and the motor / generator. By realizing this while keeping the torque from the machine unchanged,
It is an object of the present invention to propose a friction element slip prevention device in a hybrid transmission which can prevent slipping of the friction element without causing a problem that the running performance is affected by a decrease in vehicle driving force.

For this purpose, the friction element anti-slip device in the hybrid transmission of the present invention is constructed as described in claim 1.
First of all, the premise hybrid transmission is
In addition to the main power source as a prime mover, a motor / generator is provided, and the power from the main power source and the motor / generator is transmitted and output by a speed change mechanism, and the transmission state by the speed change mechanism is determined by fastening / release of a friction element. It is a decision.

The present invention is characterized in that the hybrid transmission is provided with a friction element slip detection means and a prime mover control means.
The friction element slip detection means detects that the friction element cannot maintain the engaged state due to a sudden change in torque from the prime mover.
In addition, when the friction element slip detection means detects that the friction element cannot be held, the prime mover control means prevents the change from being applied to the friction element that cannot maintain the engagement state while preventing the torque from the hybrid transmission from changing. The torque of the main power source and / or motor / generator is controlled to reduce the torque.

According to the slip prevention device for the friction element in the hybrid transmission of the present invention,
When the friction element cannot maintain the engaged state, the main power source or the motor is configured to reduce the torque to the friction element that cannot maintain the engaged state while preventing the torque from the hybrid transmission from changing. / To control the generator or both torque,
The frictional elements that can no longer maintain the engaged state can be prevented from slipping by reducing the torque, and at this time, the main power source or motor / generator, or these can be used so that the torque from the hybrid transmission does not change. Since the torque is reduced by controlling both torques, there is no problem that the running performance is affected.

Moreover, according to the anti-slip control of the friction element as described above, since it does not depend on measures to increase the hydraulic pressure margin of the line pressure, the required operating pressure corresponding to the required fastening force of the friction element in the steady state is created. Even if the line pressure is controlled to achieve the required minimum oil pressure, slipping of the friction element can be achieved.
The line pressure is not always set higher than necessary during a much longer period of time than during a transition, and the energy that drives the oil pump is consumed more than necessary, resulting in fuel consumption deterioration due to energy loss. It does not cause a problem.

Hereinafter, embodiments of the present invention will be described in detail based on examples shown in the drawings.
FIG. 1 illustrates a power train of a vehicle equipped with a hybrid transmission equipped with a friction element anti-slip device according to an embodiment of the present invention.
In this power train, the transmission mechanism 10 is coupled to the engine ENG, the motor / generator MG is attached to the transmission mechanism 10, the output shaft Out of the transmission mechanism 10 is coupled to the drive wheel 12, and the rotation from the transmission output shaft Out is performed. Is transmitted to the drive wheel 12 for driving the vehicle.

  The speed change mechanism 10 includes a three-element, two-degree-of-freedom differential device as represented by the alignment chart shown in FIG. The engine ENG is connected, the motor / generator MG is connected to the rotating element located on the opposite side, and the rotating element can be fixed by the brake B, and the transmission output shaft is connected to the rotating element located in the middle of the rotational speed forward direction. Join Out.

  In such a hybrid transmission, when the brake B is fastened and the corresponding rotating element is fixed, the gear ratio is fixed as shown by the alignment chart of FIG. 2, and the rotating element fixed by the brake B is used as a fulcrum. A lever with the rotating element coupled with ENG as the power point and the rotating element coupled with output shaft Out as the operating point increases the torque of engine ENG by the lever ratio (fixed gear ratio) and directs it toward output shaft Out. A transmission state in the low gear ratio fixed mode can be realized.

  On the other hand, when the brake B is released so that the corresponding rotating element can be rotated and fixed, the rotating element can be freely driven by the motor / generator MG (can move freely up and down on the collinear diagram of FIG. 2). Thus, it is possible to realize a transmission state in the free gear ratio mode in which torque from the engine ENG and / or torque from the motor / generator MG can be transmitted to the output shaft Out at an arbitrary gear ratio.

In the low transmission ratio fixed mode of the hybrid transmission, the slip occurrence state of the brake B (friction element) when the countermeasure of the present invention is not taken is as described below with reference to FIG. 5 (a).
Fig. 5 (a) is an operation time chart when the line pressure is controlled so that the line pressure P becomes the minimum hydraulic pressure necessary to create the required operating pressure according to the required engagement force of the brake B in the steady state. is there.
Note that because of the low gear ratio fixed mode, the torque Tmg and the rotational speed Nmg of the motor / generator MG are both zero.

At the instant t1, when the driver demands an increase in driving force and the accelerator pedal depression amount (accelerator opening) APO is increased stepwise as shown in the figure, the vehicle's target driving force tFd is As it changes, the target line pressure tP increases stepwise as indicated by the dashed line.
The actual line pressure P is controlled so as to follow the increase in the target line pressure tP, but coincides with the target line pressure tP with the predetermined response delay as shown by the solid line.

The target brake engagement torque capacity tTbo of the brake B corresponding to the target line pressure tP is as shown by a thick wavy line, and the actual brake engagement torque capacity Tbo of the brake B due to the actual line pressure P is as shown by a thick solid line.
Further, the target brake torque tTb of the brake B accompanying the change in the accelerator opening APO is as shown by a thin wavy line, whereas the required brake torque Tb of the brake B given with a predetermined characteristic is as shown by a thin solid line. Then,
At the instant t2 to t3, the actual brake engagement torque capacity Tbo is insufficient with respect to the required brake torque Tb, and the brake B slips.

  Such slip of brake B causes insufficient reaction force to make the vehicle driving force Fd unable to match the target driving force tFd as shown by the solid line. Invite.

In order to solve this problem, the target line pressure tP, as shown by the arrow in FIG. 5 (b), can be generated as shown by the arrow in FIG. It is conceivable that the oil pressure margin is increased by always increasing the oil pressure margin so that the actual line pressure P changes with time as indicated by a thick solid line.
In this case, the target brake engagement torque capacity tTbo corresponding to the target line pressure tP is raised in the direction of the arrow as indicated by a thick wavy line, and the actual brake engagement torque capacity Tbo due to the actual line pressure P is indicated by a thick solid line. Even during the transition between the instants t1 and t3, the actual brake engagement torque capacity Tbo does not become insufficient with respect to the required brake torque Tb, and the brake B can be prevented from slipping.

  However, in this case, the line pressure will always be higher than necessary during steady state, which is much longer than that at the time of transition, and the energy to drive the oil pump will be consumed more than necessary, resulting in deterioration of fuel consumption due to energy loss. Invite.

Therefore, in this embodiment, the line pressure control is kept as shown in FIG. 5 (a), and the slip of the brake B at the time of transition is prevented by executing the control program shown in FIGS.
FIG. 4 is a main routine related to the control of the hybrid transmission, and FIG. 5 is a subroutine related to the calculation processing of the target motor / generator torque in the main routine.

First, the main routine of FIG. 4 will be described.
In step S1, the target driving force tFd of the vehicle requested by the driver is calculated from the vehicle speed VSP and the accelerator opening APO based on the planned target driving map.
In step S2, the optimum driving mode (low gear ratio fixed mode, free gear ratio mode) is determined from the above target driving force tFd, vehicle speed VSP, and battery storage state SOC (carryable power). To do.
In step S3, an optimal target engine speed tNe for realizing the target driving force tFd based on the optimal driving mode determined as described above is determined.

In step S4, the optimum driving mode determined as described above and the target engine speed tNe are taken into consideration, and the optimum target engine torque tTe for realizing the target driving force tFd is determined based on these.
In step S5, based on the accelerator opening APO, a steady target brake torque tTb necessary to fix the brake B so as not to rotate the rotating element related thereto is calculated.
In step S6, the minimum target line pressure tP required to produce the required operating pressure of the brake B for achieving the steady target brake torque tTb is calculated.

In step S7, when the optimum traveling mode determined in step S2 is the free gear ratio mode, the target of the motor / generator MG is set so as to obtain a predetermined gear ratio based on the target engine speed tNe and the target engine torque tTe. The motor torque tTmg is calculated, and if the optimum travel mode determined in step S2 is the low gear ratio fixed mode, the target motor torque tTmg for preventing the brake B from slipping is calculated by executing the subroutine of FIG. .
In step S8, the final command value obtained as described above is output to the corresponding points to achieve the target value.

The calculation process shown in FIG. 4, that is, the calculation process of the target motor torque tTmg for preventing the brake B from slipping in the low gear ratio fixed mode will be described below.
In step S11, the required brake torque Tb of the brake B to be given with a predetermined characteristic is obtained from the target brake torque tTb obtained in step S5 of FIG.
In step S12, the required hydraulic pressure of the brake B that realizes the required brake torque Tb is calculated, and the minimum target line pressure tP necessary to produce this is calculated.

In step S13, based on the target line pressure tP, an actual brake engagement torque capacity Tbo obtained from the actual line pressure P considering the response delay of the line pressure control is calculated.
In step S14, a deviation ΔTb = Tb−Tbo between the actual brake engagement torque capacity Tbo and the required brake torque Tb is obtained,
In step S15, whether or not Tb> Tbo (brake B slips) or Tb ≦ Tbo (brake B does not slip) is checked based on whether the deviation ΔTb is positive or negative.
Therefore, step S15 corresponds to the friction element slip detection means in the present invention.

When it is determined in step S15 that Tb> Tbo (brake B slips), in step S16, as a motor torque correction amount opposite to the torque input thereto, so as to prevent the brake B from slipping, The deviation ΔTb obtained in step S14 is set.
In step S17, the motor torque Tmg of the motor / generator MG is corrected by a motor torque correction amount = ΔTb.
Accordingly, step S17 corresponds to the prime mover control means in the present invention.

  However, when it is determined in step S15 that Tb ≦ Tbo (the brake B does not slip), the motor torque correction amount is 0 because step S16 is skipped, and the correction of the motor torque Tmg in step S17 is substantially also performed. Is not done.

The operation of the above embodiment will be described below with reference to FIG. 6 showing an operation time chart under the same conditions as in FIG.
At instant t1, when the driver requests an increase in driving force and the accelerator pedal depression amount (accelerator opening) APO is increased stepwise as shown in the figure, the target driving force tFd of the vehicle increases as shown by the wavy line At the same time, the target line pressure tP increases stepwise as indicated by the broken line.
The actual line pressure P is controlled so as to follow the increase in the target line pressure tP, but coincides with the target line pressure tP with a predetermined response delay as shown by the solid line.

As shown in FIG. 5 (a), during the instant t2 to t3 when the required brake torque Tb becomes larger than the actual brake engagement torque capacity Tbo of the brake B due to the actual line pressure P (the brake B slips), the motor / generator MG As shown in FIG. 6, a motor torque correction amount ΔTb opposite to the torque input to the brake B is applied, and the input torque to the brake B is reduced so that the slip is prevented.
As a result, the required brake torque Tb coincides with the actual brake engagement torque capacity Tbo due to the actual line pressure P, and the required brake torque Tb is larger than the actual brake engagement torque capacity Tbo, as seen during the instants t2 to t3 in FIG. It is possible to prevent the brake B from slipping even during a transition in which the accelerator opening APO is changed during traveling in the low gear ratio fixed mode that is selected by engaging the brake B. be able to.

For this reason, the vehicle driving force Fd can be made to coincide with the target driving force tFd as shown by the solid line as in the instants t2 to t3 in FIG. 6, and the driving force is insufficient even during a transient when the accelerator opening APO is changed. There is no deterioration in the durability of the brake B.
Moreover, this problem can be solved without relying on the increase in the hydraulic margin of the line pressure described above with reference to FIG. 5 (b), so that the line pressure is always higher than necessary during a much longer period of steady state than in the transient state. It is possible to solve the above-mentioned problem without having to make it, and there is no problem that the energy for driving the oil pump is consumed more than necessary and the fuel consumption is not deteriorated.

FIG. 7 shows another example of a hybrid transmission to which the friction element anti-slip device according to the present invention can be applied.
The speed change mechanism 20, which forms the main part of the hybrid transmission, is arranged in the middle of its axial direction (left-right direction in the figure) and is provided with a first simple planetary gear set G1, and on the right side in the figure (rear end far from the engine ENG) The second simple planetary gear set G2 is provided, and the third planetary gear set G3 is provided on the left side (front end close to the engine ENG) of the figure.

These planetary gear sets G1, G2, G3 are arranged coaxially with the engine ENG, respectively, and coaxially between the planetary gear sets G1, G2, G3 and the engine ENG, the first motor / generator MG1 and the second motor / A generator MG2 is provided.
The planetary gear set G1, G2, G3 has three elements, which are sun gears S1, S2, S3, ring gears R1, R2, R3, and carriers C1, C2, C3 as rotating members. It is assumed that a two-degree-of-freedom differential device is configured by correlating with.

The carrier C1 and the ring gear R2 are coupled to each other, and these combined bodies are connected to an input shaft 21 (indicated as an input In in the collinear diagram of FIG. 8) through which the engine ENG rotates, via an engine clutch E / C. Can be combined.
An output shaft 22 (shown as output Out in the collinear diagram of FIG. 5) is connected to the carrier C2 coaxially with the input shaft 21.

The sun gear S2 and the ring gear R3 are coupled to each other, and the ring gear R1 can be coupled to the first motor / generator MG1 by the motor / generator clutch MG1 / C and can be fixed by the high & low brake HL / B.
Further, the input shaft 21 and the first motor / generator MG1 can be coupled by a series clutch S / C.
The sun gears S1 and S2 are coupled to each other, and the combination is coupled to the second motor / generator MG2.
The carrier C3 can be fixed by the low brake L / B and can be coupled to the sun gear S3 by the high clutch H / C.

The hybrid transmission configured as described above is represented by a collinear diagram as shown in FIG.
The rotational speed order of the rotating members in the first planetary gear set G1 (the fast order or the slow order depends on the speed change state) is the ring gear R1, the carrier C1, and the sun gear S1,
The rotation speeds of the rotating members in the second planetary gear set G2 are the ring gear R2, the carrier C2, and the sun gear S2.
An intermediate carrier C1 in the order of the rotational speed in the first planetary gear set G1 and a ring gear R2 in the first position in the order of the rotational speed in the second planetary gear set G2 are coupled to each other, and the second planetary gear set G2 The ring gear R3 and the sun gear S3 in the third planetary gear set G3 are coupled to the third gear sun gear S2 in the order of rotation speed and the third gear sun gear S1 in the order of rotation speed in the first planetary gear set G1, respectively. To do.

Further, a low brake L / B for fixing the carrier C3 of the third planetary gear set G3 is provided, and the carrier C3 and the sun gear S3 of the third planetary gear set G3 are coupled to each other, resulting in the sun gears S1, S2 A high clutch H / C that rotates the unit integrally is provided.
The ring gear R1 of the first planetary gear set G1 can be coupled to the first motor / generator MG1 by the motor / generator clutch MG1 / C and can be fixed by the high & low brake HL / B.

The input EN from the engine ENG can be coupled to the carrier C1 of the first planetary gear set G1 and the ring gear R2 of the second planetary gear set G2 via the engine clutch E / C. The engine ENG and the first motor / generator MG1 can be connected to each other by series clutch S / C.
An output Out to the wheel drive system is coupled to the carrier C2 of the second planetary gear set G2, and the second motor / generator is connected to the sun gear S1 of the first planetary gear set G1 and the sun gear S3 of the third planetary gear set G3. Join MG2.
The horizontal axis in FIG. 8 represents the distance ratio between the rotating members determined by the gear ratio of the planetary gear sets G1, G2, G3, and the vertical axis represents the rotating speed of the rotating member (upward is the forward rotating speed, 0 Represents the reverse rotation speed).

The hybrid transmission represented by the collinear diagram of FIG. 8 described above releases the engine clutch E / C, disconnects the engine ENG from the carrier C1 and the ring gear R2, and engages the series clutch S / C to connect the motor / generator MG1. Is driven by the engine ENG, and the motor / generator clutch MG1 / C is released to disconnect the motor / generator MG1 from the ring gear R1.
In other words, the motor / generator MG1 is driven at a constant rotational speed by the torque from the engine ENG to generate electric power, and the generated electric power is used, and if necessary, the battery electric power is used to output the motor / generator MG2 on the output side. To drive the vehicle, and determine the output Out to the drive system only by the power from the motor / generator MG2.

  Further, the hybrid transmission represented by the collinear diagram of FIG. 8 engages the engine clutch E / C so that the engine rotation is input to the carrier C1 and the ring gear R2, and releases the series clutch S / C. In the state where motor / generator MG1 is disconnected from engine ENG and motor / generator clutch MG1 / C is engaged and motor / generator MG1 is coupled to ring gear R1, it operates as follows.

When the carrier C3 and the sun gear S3 of the planetary gear set G3 are coupled by the engagement of the high clutch H / C, all the rotating members of the planetary gear set G3 are rotated together, so that the collinear line in FIG. In the figure, the sun gear S2 coincides with the sun gears S1 and S3.
In this case, the lever G2 in FIG. 8 rides on the lever G1, and the gear train constituted by the planetary gear sets G1 and G2 provides a speed change state represented by a straight line of four elements and two degrees of freedom, and the rotational speed of the rotating member In this order, the arrangement is motor / generator MG1, input In from engine ENG, output Out to wheel drive system, and motor / generator MG2.
Accordingly, the rotation of the output Out (carrier C2) becomes higher than that in the gear shifting state of FIG. 8, and the gear ratio can be freely changed in the high gear ratio region. (High-side free transmission ratio mode)

When the carrier C3 is fixed by engaging the low brake L / B, the sun gear S3 of the planetary gear set G3 rotates positively and the ring gear R3 rotates negatively as illustrated in the collinear diagram of FIG. A shift state in which the sun gear S1 rotates forward and the sun gear S2 rotates negatively is provided.
Accordingly, the rotation of the output Out (carrier C2) is lower than the rotation in the shift state when the high clutch H / C is engaged, and the gear ratio can be freely changed in the low gear ratio range. (Low side free gear ratio mode)

  In the above-described free gear ratio mode, the power from the engine ENG and the power from one or both of the motor / generators MG1 and MG2 are used both when shifting in the high gear ratio range and when shifting in the low gear ratio range. Thus, the output Out to the drive system is determined.

In addition, when the high & low brake HL / B is engaged in the low side free gear ratio mode, the ring gear R1 is fixed by this, and coupled with the fixing of the carrier C3 by the engagement of the low brake L / B, The hybrid transmission is represented by a collinear chart as shown in FIG.
In this case, the lever G1 on the nomographic chart of FIG. 9 transmits power while rotating with the ring gear R1 fixed by fastening the high and low brake HL / B as a fulcrum, so it is fixed at the low gear ratio. In this state, a low gear ratio fixed mode in which power transmission is performed is obtained.

In addition, when the low gear ratio fixed mode is selected, the high / low brake HL / B is released so that the motor / generator MG1 can freely drive the ring gear R1 and the high gear H / C is engaged, so that the sun gear S2 Is coupled to the sun gear S1 (S3), and the sun gear S2, (S3) is also fixed in cooperation with the fixing of the carrier C3 by fastening the low brake L / B. As shown in the diagram.
In this case, as shown in the collinear diagram of FIG. 10, the levers G1 and G2 are superposed to form one lever, and the lever rotates around the end near the output shaft Out as a fulcrum while the motor / generator Since the power of MG1 and engine ENG is transmitted to the output shaft Out, the power is transmitted with the gear ratio fixed.
In this case, since the rotational speed of the output shaft Out is made higher than in the case of FIG. 9, the transmission gear ratio is an intermediate transmission gear ratio, and an intermediate transmission gear ratio fixed mode that transmits power while being fixed to this intermediate transmission gear ratio. Is obtained.

Further, when the high & low brake HL / B is engaged in the high side free gear ratio mode, the ring gear R1 is fixed by this, and also the low brake L / B is released and the low brake L / B is engaged. Thus, in combination with the fact that the sun gears S2 and (S3) can be driven freely by the motor / generator MG2, the hybrid transmission is represented by a collinear diagram as shown in FIG.
In this case, levers G1 and G2 on the nomograph of FIG. 11 are united and transmit power while rotating with ring gear R1 fixed by fastening of high & low brake HL / B as a fulcrum, A high gear ratio fixed mode is obtained in which power transmission is performed in a state where the rotational speed of the output shaft Out is fixed to a high gear ratio that is further increased than in the case of FIG.

  The selection of the driving mode, the output control of the engine ENG and the motor / generators MG1 and MG2 constituting the prime mover, and the anti-slip control of the friction element targeted by the present invention are controlled by the control system shown in FIG. Do.

The control system shown in FIG. 12 includes a hybrid controller 31. The hybrid controller 31 obtains a target engine torque and commands the engine controller 32. The engine controller 32 controls the opening degree of the throttle valve 33, and the fuel injection by the fuel injection device 34. The engine ENG is controlled by the amount control and the ignition timing control by the ignition device 35 so that the target engine torque is achieved.
The hybrid controller 31 further calculates the target torque of the motor / generators MG1 and MG2 and instructs the corresponding motor controllers 36 and 37 from the battery 40 via the corresponding inverters 38 and 39. By controlling the power supply amount, the motor / generators MG1 and MG2 are controlled to achieve their target torques.

The hybrid controller 31 obtains the above various target values based on the vehicle speed VSP, the accelerator opening APO, and the battery storage state SOC (carryable power) from the battery controller 41. The battery controller 41 determines the battery voltage, current consumption, Based on information such as the battery temperature, the state of the battery 40 is detected to obtain the chargeable / dischargeable power of the battery, and this is input to the hybrid controller 31 as the battery storage state SOC (capable power to be taken out).
The hybrid controller 31 gives predetermined clutch fastening torque capacity commands to these clutches when switching the states of the engine clutch E / C and the series clutch S / C, and the clutch responds by current control of a solenoid valve (not shown). The necessary oil pressure is supplied to the clutch.

The calculation process of the hybrid controller 31 will be described below with reference to FIGS.
FIG. 13 shows a main routine. In step S21, the vehicle target driving force tFd requested by the driver is calculated from the vehicle speed VSP and the accelerator opening APO based on the planned target driving map.
In step S22, from the target driving force tFd, the vehicle speed VSP, and the battery storage state SOC (carryable power), the driving mode (low side free gear ratio mode, high side free gear ratio mode, Low gear ratio fixed mode, intermediate gear ratio fixed mode, high gear ratio fixed mode) are determined, and the engagement hydraulic pressure of the friction elements (clutch and brake) necessary to realize the determined mode is determined.

In step S23, the optimum driving mode tNe optimum for realizing the target driving force tFd is determined based on the optimum traveling mode determined as described above, the engagement state of the friction element, and the like.
In step S24, the optimum target engine torque tTe is determined in consideration of the optimum travel mode, the friction element engagement state, and the target engine speed tNe.
In step S25, when the optimum traveling mode obtained in step S22 is different from the currently selected traveling mode, the mode transition for the control that favorably makes the transition from the currently selected traveling mode to the optimum traveling mode. Perform the operation.

In step S26, based on the accelerator opening APO, a target engagement torque of a steady friction element necessary to prevent the friction element (clutch or brake) from slipping is calculated.
In step S27, a transient correction of the target driving force tFd when there is a change in the accelerator opening APO is performed.
In steps S28 and S29, the target motor torque tTmg1, tTmg2 of the motor / generators MG1, MG2 necessary to realize the target driving force tFd in consideration of the target engine speed tNe and the target engine torque tTe. Is calculated.
In step S30, the minimum target line pressure tP required to produce the required operating pressure of the friction element for achieving the target engagement torque of the friction element obtained in step S26 is calculated.
In step S31, the final command value obtained as described above is output to the corresponding points to achieve the target value.

FIG. 14 is a subroutine showing a calculation process of the target line pressure tP performed in step S30 of FIG.
First, in step S41 to step S44, which of the low-side free gear ratio mode, the high-side free gear ratio mode, the low gear ratio fixed mode, the intermediate gear ratio fixed mode, and the high gear ratio fixed mode is selected. Determine.
When it is determined in step S41 that the current travel mode is the low-side free gear ratio mode, line pressure control for the mode is executed in step S45, and in step S42 the current travel mode is the high-side free gear ratio mode. Is determined in step S46, the line pressure control for the mode is executed in step S46, and in step S43, the line pressure control for the mode is executed in step S47. If it is determined in step S44 that the intermediate gear ratio fixed mode is set, line pressure control for the mode is executed in step S48, and if it is determined in step S44 that the high gear ratio fixed mode is selected, step S49 is performed. The line pressure control for the mode is executed at.

Here, in the case where the hybrid transmission is in the low gear ratio fixed mode represented by the collinear chart of FIG. 9, the slip occurrence state of the high & low brake HL / B when the countermeasure of the present invention is not taken is As described below based on FIG. 16 (a), the slip occurrence state of the engine clutch E / C (see FIGS. 7 and 8) is as described below based on FIG. 18 (a). .
These Fig. 16 (a) and Fig. 18 (a) are necessary to create the line pressure P and the required operating pressure according to the required engagement force of the high and low brakes HL / B and engine clutch E / C at steady state. It is an operation | movement time chart at the time of controlling a line pressure so that it may become a minimum hydraulic pressure.

First, the slip occurrence state of the high & low brake HL / B will be described with reference to FIG.
When the driver requests an increase in driving force at the instant t1 and increases the accelerator opening APO stepwise as shown in the figure, the target driving force tFd of the vehicle changes with time as indicated by the wavy line and the target line pressure tP rises stepwise as shown by the wavy line.
The actual line pressure P is controlled so as to follow the increase in the target line pressure tP, but coincides with the target line pressure tP with the predetermined response delay as shown by the solid line.

The target brake engagement torque capacity tTbo of the high & low brake HL / B corresponding to the target line pressure tP is as shown by the thick wavy line, and the actual brake engagement torque capacity Tbo of the high & low brake HL / B by the actual line pressure P Is as shown by a thick solid line.
The target brake torque tTb of the high & low brake HL / B that accompanies the change in the accelerator opening APO is as shown by a thin wavy line, and the required brake torque of the high & low brake HL / B given to it with a predetermined characteristic. If Tb is as shown by a thin solid line,
At the instant t2 to t3, the actual brake engagement torque capacity Tbo is insufficient with respect to the required brake torque Tb, and the high & low brake HL / B slips.

  The slip of the high & low brake HL / B generates the rotational speed Nmg1 of the motor / generator MG1 and causes insufficient reaction force so that the vehicle driving force Fd cannot be matched with the target driving force tFd as shown by the solid line. In the transition, the driving performance is reduced due to insufficient driving force, and the durability of the high & low brake HL / B is reduced.

To solve this problem, the target line pressure tP is indicated by the arrow in Fig. 16 (b) so that the required operating pressure corresponding to the required fastening force of the high and low brakes HL / B can be created even during transition. It is conceivable that the oil pressure margin is increased by always increasing the oil pressure margin as compared with FIG. 16 (a) so that the actual line pressure P changes with time as indicated by a thick solid line.
In this case, the target brake engagement torque capacity tTbo corresponding to the target line pressure tP is raised in the direction of the arrow as indicated by a thick wavy line, and the actual brake engagement torque capacity Tbo due to the actual line pressure P is indicated by a thick solid line. Even during the transition between the instants t1 and t3, the actual brake engagement torque capacity Tbo does not become insufficient with respect to the required brake torque Tb, and slipping of the high & low brake HL / B can be prevented.

  However, in this case, the line pressure will always be higher than necessary during steady state, which is much longer than that at the time of transition, and the energy to drive the oil pump will be consumed more than necessary, resulting in deterioration of fuel consumption due to energy loss. Invite.

Next, referring to FIG. 18 (a), the state of occurrence of slipping of the engine clutch E / C will be described.
When the driver requests an increase in driving force at the instant t1 and increases the accelerator opening APO stepwise as shown in the figure, the target driving force tFd of the vehicle changes with time as indicated by the wavy line and the target line pressure tP rises stepwise as shown by the wavy line.
The actual line pressure P is controlled so as to follow the increase in the target line pressure tP, but coincides with the target line pressure tP with the predetermined response delay as shown by the solid line.

The target clutch engagement torque capacity tTco of the engine clutch E / C corresponding to the target line pressure tP is as shown by a thick wavy line, and the actual clutch engagement torque capacity Tco of the engine clutch E / C by the actual line pressure P is a thick solid line. It is the time to show.
Further, the target clutch torque tTc of the engine clutch E / C accompanying the change in the accelerator opening APO is as shown by a thin wavy line, whereas the required clutch torque Tc of the engine clutch E / C given with a predetermined characteristic is a thin solid line If it is as shown in
At the instant t2 to t3, the actual clutch engagement torque capacity Tco becomes insufficient with respect to the required clutch torque Tc, and the engine clutch E / C slips.

  The slip of the engine clutch E / C increases the engine clutch E / C rotation speed (engine rotation speed Ne) (i.e., causes the engine to blow idle) and causes a shortage of torque from the engine, resulting in a decrease in vehicle driving force Fd. As indicated by the solid line, it cannot be made to coincide with the target driving force tFd, resulting in a decrease in running performance due to insufficient driving force and a decrease in durability of the engine clutch E / C during the transition.

In order to solve this problem, the target line pressure tP, as indicated by the arrow in FIG. 18 (b), can be created so that the required operating pressure can be created according to the required engagement force of the engine clutch E / C, including during transition. It is conceivable that the oil pressure margin is increased by always increasing the oil pressure margin from (a) so that the actual line pressure P changes over time as indicated by a thick solid line.
In this case, the target clutch engagement torque capacity tTco corresponding to the target line pressure tP is raised in the direction of the arrow as indicated by a thick wavy line, and the actual clutch engagement torque capacity Tco due to the actual line pressure P is increased as indicated by a thick solid line. Even during the transition between the instants t1 and t3, the actual clutch engagement torque capacity Tco does not become insufficient with respect to the required clutch torque Tc, and the slip of the engine clutch E / C can be prevented.

  However, in this case, the line pressure will always be higher than necessary during steady state, which is much longer than that at the time of transition, and the energy to drive the oil pump will be consumed more than necessary, resulting in deterioration of fuel consumption due to energy loss. Invite.

  Therefore, in this embodiment, the slip at the time of transition of the high & low brake HL / B described above with reference to FIG. 16 (a) and the slip at the time of transition of the engine clutch E / C described above with reference to FIG. FIG. 15 shows the line pressure control in step S47 of FIG. 14 to be executed in the low gear ratio fixed mode instead of preventing the increase of the line pressure described above with reference to FIGS. 16 (b) and 18 (b). At the same time, in the control program of FIG. 15, the anti-slip control at the transient time of the high & low brake HL / B and the engine clutch E / C is executed as follows.

In step S51, the required brake torque Tb of the high & low brake HL / B to be given with a predetermined characteristic is obtained from the target brake torque tTb of the high & low brake HL / B obtained in step S26 of FIG.
In step S52, the required hydraulic pressure of the high & low brake HL / B that realizes the required brake torque Tb is calculated.
In calculating the required oil pressure, the required oil pressure can be obtained by referring to a predetermined torque / oil pressure map.

  In determining the relationship between the torque and the hydraulic pressure, the brake torque Tb is calculated by multiplying the number of brake plates × 2, the effective brake radius, the friction coefficient, the hydraulic pressure × the working piston area−the piston friction−the return spring force. Since it can be regarded as a constant except that the friction coefficient varies depending on the temperature, a torque / hydraulic map for each temperature is prepared in advance, and based on this, the high & low brake HL corresponding to the required brake torque Tb Necessary oil pressure of / B can be obtained.

When obtaining the relationship between torque and hydraulic pressure, it can also be obtained by learning the characteristics when the high & low brake HL / B is released.
In other words, the steady engagement torque of the high & low brake HL / B is calculated from the torque of the engine ENG and the motor / generators MG1, MG2 and the output torque of the hybrid transmission. From the relationship with the hydraulic pressure at the time of release of the brake HL / B, the relationship between the hydraulic pressure and the tightenable torque of the high & low brake HL / B is memorized, and interpolation is performed using this memorized relationship and the past relationship. & The constant required hydraulic pressure for low brake HL / B can be estimated.

The relationship between torque and hydraulic pressure can be obtained by learning the characteristics during the slip control of the high & low brake HL / B.
In other words, when the high & low brake HL / B is slipping, it is possible to estimate the torque received by the high & low brake HL / B based on the axial acceleration, and obtain it from the relationship between this torque and hydraulic pressure. .

In step S53, the required brake torque of the low brake L / B to be given with a predetermined characteristic is obtained from the target brake torque of the low brake L / B obtained in step S26 of FIG.
In step S54, the required hydraulic pressure of the low brake L / B that realizes the required brake torque is calculated.
In calculating the required oil pressure, the required oil pressure can be obtained by referring to a planned torque / oil pressure map obtained in the same manner as described above for the case of the high & low brake HL / B.

In step S55, the required clutch torque of the engine clutch E / C to be given with a predetermined characteristic is obtained from the target clutch torque of the engine clutch E / C obtained in step S26 of FIG.
In step S56, the required hydraulic pressure of the engine clutch E / C that realizes the required clutch torque is calculated.
Even when the required hydraulic pressure is calculated, the required hydraulic pressure can be obtained by referring to the planned torque / hydraulic map obtained in the same manner as described above for the case of the high & low brake HL / B.

In step S57, the frictional element is required depending on whether the required hydraulic pressure for the high and low brake HL / B is higher than the required hydraulic pressure for the low brake L / B and higher than the required hydraulic pressure for the engine clutch E / C. Judgment is made as to whether the required hydraulic pressure of the high & low brake HL / B is the highest among the hydraulic pressures.
Otherwise, depending on whether or not the required oil pressure of the low brake L / B is higher than the required oil pressure of the engine clutch E / C in step S58, the required oil pressure of the low brake L / B among the required oil pressures of the friction elements is determined. It is determined whether or not it is the highest. Otherwise, it is determined that the required hydraulic pressure of the engine clutch E / C is the highest among the required hydraulic pressures of the friction elements.

If it is determined in step S57 that the required hydraulic pressure for the high & low brake HL / B is the highest, in step S59, the required hydraulic pressure for the high & low brake HL / B is set to the target line pressure tP. In addition to the low brake HL / B, the minimum target line pressure tP required to prevent all friction elements from slipping in a steady state is determined.
If it is determined in step S58 that the required hydraulic pressure for the low brake L / B is the highest, in step S60, the required hydraulic pressure for the low brake L / B is set to the target line pressure tP. Of course, the minimum target line pressure tP required to prevent all the friction elements from slipping in the steady state is determined.
If it is determined in step S58 that the required hydraulic pressure of the engine clutch E / C is the highest, in step S61, the required hydraulic pressure of the engine clutch E / C is set to the target line pressure tP. Of course, the minimum target line pressure tP required to prevent all the friction elements from slipping in the steady state is determined.
Therefore, step S57 to step S61 correspond to the line pressure control means in the present invention.

In step S62, the actual line pressure P is estimated by applying a second order delay to the target line pressure tP determined as described above, and the actual brake engagement torque of the high & low brake HL / B obtained by this actual line pressure P. Calculate the capacity Tbo.
In step S63, a deviation ΔTb = Tb−Tbo between the required brake torque Tb (step S51) of the high and low brake HL / B and the actual brake engagement torque capacity Tbo is calculated, and the positive / negative judgment of the deviation ΔTb is performed. Check whether Tb> Tbo (high & low brake HL / B slips) or Tb ≦ Tbo (high & low brake HL / B does not slip).
Accordingly, step S63 corresponds to the friction element slip detection means in the present invention.

When it is determined in step S63 that Tb> Tbo (high & low brake HL / B slips), in step S64, the deviation ΔTb is added to the target motor torque tTmg1 of the motor / generator MG1 obtained in step S28 of FIG. Is added.
The addition of the deviation ΔTb leads to giving the motor / generator MG1 a motor torque in the opposite direction to the torque inputted to the high / low brake HL / B to prevent slippage.
Therefore, step S64 corresponds to the prime mover control means in the present invention.

  However, when it is determined in step S63 that Tb ≦ Tbo (high & low brake HL / B does not slip), the motor torque correction amount of the motor / generator MG1 is 0 because step S64 is skipped, and in step S64. The target motor torque tTmg1 is not substantially corrected.

The above-described operation will be described below with reference to FIG. 17 showing an operation time chart under the same conditions as in FIG.
At instant t1, when the driver requests an increase in driving force and the accelerator pedal depression amount (accelerator opening) APO is increased stepwise as shown in the figure, the target driving force tFd of the vehicle increases as shown by the wavy line At the same time, the target line pressure tP increases stepwise as indicated by the broken line.
The actual line pressure P is controlled so as to follow the increase in the target line pressure tP, but coincides with the target line pressure tP with a predetermined response delay as shown by the solid line.

As shown in FIG. 16 (a), the required brake torque Tb becomes larger than the actual brake engagement torque capacity Tbo of the high & low brake HL / B due to the actual line pressure P (the high & low brake HL / B slips). During t2 to t3, the motor torque Tmg1 of the motor / generator MG1 is given a motor torque correction amount ΔTb opposite to the torque input to the high & low brake HL / B as shown in FIG. Reduce the input torque to the brake HL / B to prevent the above slip.
As a result, the required brake torque Tb coincides with the actual brake engagement torque capacity Tbo due to the actual line pressure P, and the required brake torque Tb is larger than the actual brake engagement torque capacity Tbo, as seen during the instants t2 to t3 in FIG. High & low brakes even during transitions when the accelerator opening APO is changed while driving in the low gear ratio fixed mode that is selected by engaging the high & low brakes HL / B. HL / B can be prevented from slipping.

For this reason, the vehicle driving force Fd can be made to coincide with the target driving force tFd as shown by the solid line as in the instants t2 to t3 in FIG. 17, and the driving force is insufficient even when the accelerator opening APO is changed. There is no loss of durability of the high and low brake HL / B.
Moreover, since this problem can be solved without relying on the increase in the hydraulic margin of the line pressure described above with reference to FIG. 16 (b), the line pressure is always higher than necessary during a much longer period of steady state than in the transient state. It is possible to solve the above-mentioned problem without having to make it, and there is no problem that the energy for driving the oil pump is consumed more than necessary and the fuel consumption is not deteriorated.

In step S65 of FIG. 15, the actual clutch engagement torque capacity Tco of the engine clutch E / C obtained from the actual line pressure P estimated in step S62 is calculated.
In step S66, a deviation ΔTc = Tc−Tco between the clutch required torque Tc of the engine clutch E / C (step S55) and the actual clutch engagement torque capacity Tco is calculated, and Tc> It is checked whether Tco (engine clutch E / C slips) or Tc ≦ Tco (engine clutch E / C does not slip).
Therefore, step S66 corresponds to the friction element slip detection means in the present invention.

When it is determined in step S66 that Tc> Tco (engine clutch E / C slips), in step S67, the above deviation ΔTc is added to the target motor torques tTmg1 and tTmg2 obtained in steps S28 and S29 of FIG. Add proportionally according to the gear ratio.
The addition by the proportional distribution of the deviation ΔTc is to maintain the output torque unchanged by compensating for the reduction in transmission output torque caused by the engine torque limit for preventing slipping of the engine clutch E / C performed in the next step S68. Leads to
In step S68, in order to reduce the engine torque input to the engine clutch E / C so as to prevent slippage, the target engine torque tTe obtained in step S24 in FIG. It is limited to the actual clutch engagement torque capacity Tco of the clutch E / C.
Therefore, step S67 and step S68 correspond to the prime mover control means in the present invention.

  However, when it is determined in step S66 that Tc ≦ Tco (the engine clutch E / C does not slip), step S67 and step S68 are skipped, so that the motor torque correction amount of the motor / generators MG1 and MG2 and the engine ENG The torque correction amount is 0, and the correction of the target motor torques tTmg1 and tTmg2 and the correction of the target engine torque tTe in steps S67 and S68 are not substantially performed.

The above operation will be described below with reference to FIG. 19 showing an operation time chart under the same conditions as in FIG.
When the driver requests an increase in driving force at the instant t1 and increases the accelerator opening APO stepwise as shown in the figure, the target driving force tFd of the vehicle rises as indicated by the wavy line and the target line pressure tP Rises stepwise as indicated by the wavy line.
The actual line pressure P is controlled so as to follow the increase in the target line pressure tP, but coincides with the target line pressure tP with a predetermined response delay as shown by the solid line.

As shown in FIG. 18 (a), the required clutch torque Tc becomes larger than the actual clutch engagement torque capacity Tco of the engine clutch E / C due to the actual line pressure P (the engine clutch E / C slips) during instants t2 to t3. As a result of the engine torque Te being limited to the actual clutch engagement torque capacity Tco of the engine clutch E / C, the wavy line level corresponding to the level shown in FIG. 18 (a) is lowered to the level shown by the solid line.
As a result, the input torque to the engine clutch E / C is reduced, so that the required clutch torque Tc of the engine clutch E / C matches the actual clutch engagement torque capacity Tco as shown in the figure, and therefore Tc> Tco. The engine clutch E / C slips even when the accelerator opening APO is changed during driving in the low gear ratio fixed mode that is selected by engaging the engine clutch E / C. Can be prevented.

For this reason, the vehicle driving force Fd can be made to coincide with the target driving force tFd as shown by the solid line as in the instant t2 to t3 in FIG. 19, and the driving force is insufficient even during a transient when the accelerator opening APO is changed. There is no reduction in the durability of the engine clutch E / C.
Moreover, since this problem can be solved without relying on the increase in the hydraulic pressure margin of the line pressure described above with reference to FIG. 18 (b), the line pressure is always higher than necessary during a much longer steady state than during the transient state. It is possible to solve the above-mentioned problem without having to make it, and there is no problem that the energy for driving the oil pump is consumed more than necessary and the fuel consumption is not deteriorated.

By the way, the reduction of the engine torque performed to prevent the engine clutch E / C from slipping causes a reduction in transmission output torque, which adversely affects the running performance of the vehicle.
Therefore, in this embodiment, the deviation ΔTc between the required clutch torque Tc of the engine clutch E / C and the actual clutch engagement torque capacity Tco is added to the target motor torque tTmg1, tTmg2 by proportional distribution according to the gear ratio.
The motor torques Tmg1 and Tmg2 of the motor / generators MG1 and MG2 are increased from the level indicated by the wavy line, which is the same level as shown in FIG. 18 (a), to the level indicated by the solid line. It is possible to avoid the problem that the transmission performance of the vehicle is adversely affected without the reduction of the transmission output torque accompanying the reduction of the engine torque performed for slip prevention.

20 to 24 show the low brake L / B and the engine clutch E / that should be engaged when the mode is selected when the hybrid transmission is in the intermediate gear ratio fixed mode represented by the collinear diagram of FIG. An embodiment in which slippage of C is prevented will be described.
First, when the hybrid transmission is in the intermediate speed ratio fixed mode, the slip occurrence state of the low brake L / B when the countermeasure of the present invention is not taken is as described below based on FIG. In addition, the slip occurrence state of the engine clutch E / C is as described below based on FIG. 23 (a).
These Fig. 21 (a) and Fig. 23 (a) show the line pressure P, which is the minimum required to produce the required operating pressure according to the required engagement force of the low brake L / B and engine clutch E / C in the steady state. It is an operation | movement time chart at the time of controlling a line pressure so that it may become a limit hydraulic pressure.

First, the slip occurrence state of the low brake L / B will be described with reference to FIG.
When the driver requests an increase in driving force at the instant t1 and increases the accelerator opening APO stepwise as shown in the figure, the target driving force tFd of the vehicle changes with time as indicated by the wavy line and the target line pressure tP rises stepwise as shown by the wavy line.
The actual line pressure P is controlled so as to follow the increase in the target line pressure tP, but coincides with the target line pressure tP with the predetermined response delay as shown by the solid line.

The target brake engagement torque capacity tTbo of the low brake L / B corresponding to the target line pressure tP is as shown by a thick wavy line, and the actual brake engagement torque capacity Tbo of the low brake L / B by the actual line pressure P is a thick solid line. It is the time to show.
The target brake torque tTb of the low brake L / B accompanying the change in the accelerator opening APO is as shown by a thin wavy line, whereas the required brake torque Tb of the low brake L / B given with a predetermined characteristic is a thin solid line If it is as shown in
At the instant t2 to t3, the actual brake engagement torque capacity Tbo becomes insufficient with respect to the required brake torque Tb, and the low brake L / B slips.

  This low brake L / B slip generates the rotation speed Nmg2 of the motor / generator MG2 and causes insufficient reaction force to make the vehicle driving force Fd unable to match the target driving force tFd as shown by the solid line. Occasionally, driving performance is reduced due to insufficient driving force, and low brake L / B durability is reduced.

In order to solve this problem, the target line pressure tP, as indicated by the arrow in FIG. 21 (b), can be created so that the required operating pressure corresponding to the required engagement force of the low brake L / B can be created even during transition. It is conceivable that the oil pressure margin is increased by always increasing the oil pressure margin from (a) so that the actual line pressure P changes over time as indicated by a thick solid line.
In this case, the target brake engagement torque capacity tTbo corresponding to the target line pressure tP is raised in the direction of the arrow as indicated by a thick wavy line, and the actual brake engagement torque capacity Tbo due to the actual line pressure P is indicated by a thick solid line. Even during the transition between the instants t1 and t3, the actual brake engagement torque capacity Tbo does not become insufficient with respect to the required brake torque Tb, and the slip of the low brake L / B can be prevented.

  However, in this case, the line pressure will always be higher than necessary during steady state, which is much longer than that at the time of transition, and the energy to drive the oil pump will be consumed more than necessary, resulting in deterioration of fuel consumption due to energy loss. Invite.

Next, referring to FIG. 23 (a), the state of occurrence of slip of the engine clutch E / C will be described.
When the driver requests an increase in driving force at the instant t1 and increases the accelerator opening APO stepwise as shown in the figure, the target driving force tFd of the vehicle changes with time as indicated by the wavy line and the target line pressure tP rises stepwise as shown by the wavy line.
The actual line pressure P is controlled so as to follow the increase in the target line pressure tP, but coincides with the target line pressure tP with the predetermined response delay as shown by the solid line.

The target clutch engagement torque capacity tTco of the engine clutch E / C corresponding to the target line pressure tP is as shown by a thick wavy line, and the actual clutch engagement torque capacity Tco of the engine clutch E / C by the actual line pressure P is a thick solid line. It is the time to show.
Further, the target clutch torque tTc of the engine clutch E / C accompanying the change in the accelerator opening APO is as shown by a thin wavy line, whereas the required clutch torque Tc of the engine clutch E / C given with a predetermined characteristic is a thin solid line If it is as shown in
At the instant t2 to t3, the actual clutch engagement torque capacity Tco becomes insufficient with respect to the required clutch torque Tc, and the engine clutch E / C slips.

  The slip of the engine clutch E / C increases the engine clutch E / C rotation speed (engine rotation speed Ne) (i.e., causes the engine to blow idle) and causes a shortage of torque from the engine, resulting in a decrease in vehicle driving force Fd. As indicated by the solid line, it cannot be made to coincide with the target driving force tFd, resulting in a decrease in running performance due to insufficient driving force and a decrease in durability of the engine clutch E / C during the transition.

In order to solve this problem, the target line pressure tP, as indicated by the arrow in FIG. 23 (b), can be created so that the required operating pressure corresponding to the required engagement force of the engine clutch E / C can be created even during transition. It is conceivable that the oil pressure margin is increased by always increasing the oil pressure margin from (a) so that the actual line pressure P changes over time as indicated by a thick solid line.
In this case, the target clutch engagement torque capacity tTco corresponding to the target line pressure tP is raised in the direction of the arrow as indicated by a thick wavy line, and the actual clutch engagement torque capacity Tco due to the actual line pressure P is increased as indicated by a thick solid line. Even during the transition between the instants t1 and t3, the actual clutch engagement torque capacity Tco does not become insufficient with respect to the required clutch torque Tc, and the slip of the engine clutch E / C can be prevented.

  However, in this case, the line pressure will always be higher than necessary during steady state, which is much longer than that at the time of transition, and the energy to drive the oil pump will be consumed more than necessary, resulting in deterioration of fuel consumption due to energy loss. Invite.

  Therefore, in this embodiment, the slip at the time of transition of the low brake L / B described above with reference to FIG. 21 (a) and the slip at the time of transition of the engine clutch E / C described above with reference to FIG. The line pressure control in step S48 in FIG. 14 to be executed in the intermediate gear ratio fixed mode is performed as shown in FIG. 20 instead of preventing the line pressure from increasing as described above with respect to 21 (b) and FIG. 23 (b). At the same time, in the control program of FIG. 20, the slip prevention control at the transient time of the low brake L / B and the engine clutch E / C is executed as follows.

In step S71, the required brake torque Tb of the low brake L / B to be given with a predetermined characteristic is obtained from the target brake torque tTb of the low brake L / B obtained in step S26 of FIG.
In step S72, the required hydraulic pressure of the low brake L / B that realizes the required brake torque Tb is calculated.
In calculating the required oil pressure, the required oil pressure can be calculated with reference to a torque / oil pressure map that can be obtained in the same manner as described above.

In step S73, the required clutch torque of the engine clutch E / C to be given with a predetermined characteristic is obtained from the target clutch torque of the engine clutch E / C obtained in step S26 of FIG.
In step S74, the required hydraulic pressure of the engine clutch E / C that realizes the required clutch torque is calculated.
In calculating the required oil pressure, the required oil pressure can be calculated with reference to a torque / oil pressure map that can be obtained in the same manner as described above.

In step S75, the required clutch torque of the high clutch H / C to be given with a predetermined characteristic is obtained from the target clutch torque of the high clutch H / C obtained in step S26 of FIG.
In step S76, the required hydraulic pressure of the high clutch H / C that realizes the required clutch torque is calculated.
In calculating the required oil pressure, it can be calculated with reference to a torque / oil pressure map that can be obtained in the same manner as described above.

In step S77, the required hydraulic pressure of the friction element is determined depending on whether the required hydraulic pressure of the low brake L / B is higher than the required hydraulic pressure of the engine clutch E / C and higher than the required hydraulic pressure of the high clutch H / C. Of these, it is determined whether or not the required hydraulic pressure for the low brake L / B is the highest.
Otherwise, in step S78, the engine clutch E / C is required among the required hydraulic pressures of the friction elements depending on whether the required oil pressure of the engine clutch E / C is higher than the required oil pressure of the high clutch H / C. It is determined whether or not the hydraulic pressure is the highest. Otherwise, it is determined that the required hydraulic pressure of the high clutch H / C is the highest among the required hydraulic pressures of the friction elements.

If it is determined in step S77 that the required hydraulic pressure for the low brake L / B is the highest, in step S79, the required hydraulic pressure for the low brake L / B is set to the target line pressure tP, etc. Of course, the minimum target line pressure tP required to prevent all the friction elements from slipping in the steady state is determined.
When it is determined in step S78 that the required hydraulic pressure of the engine clutch E / C is the highest, in step S80, the required hydraulic pressure of the engine clutch E / C is set to the target line pressure tP. Of course, the minimum target line pressure tP required to prevent all the friction elements from slipping in the steady state is determined.
When it is determined in step S78 that the required hydraulic pressure of the high clutch H / C is the highest, in step S81, the required hydraulic pressure of the high clutch H / C is set to the target line pressure tP. Of course, the minimum target line pressure tP required to prevent all the friction elements from slipping in the steady state is determined.
Therefore, step S77 to step S81 correspond to the line pressure control means in the present invention.

In step S82, the actual line pressure P is estimated by applying a second order delay to the target line pressure tP determined as described above, and the actual brake engagement torque capacity Tbo of the low brake L / B obtained from the actual line pressure P is obtained. Is calculated.
In step S83, a deviation ΔTb = Tb−Tbo between the required brake torque Tb of the low brake L / B (step S71) and the actual brake engagement torque capacity Tbo is calculated, and Tb> Check whether Tbo (low brake L / B slips) or Tb ≦ Tbo (low brake L / B does not slip).
Therefore, step S83 corresponds to the friction element slip detection means in the present invention.

When it is determined in step S83 that Tb> Tbo (low brake L / B slips), in step S84, the deviation ΔTb is added to the target motor torque tTmg2 of the motor / generator MG2 obtained in step S29 of FIG. To do.
The addition of the deviation ΔTb leads to giving the motor / generator MG2 a motor torque in the opposite direction to the torque input thereto so as to prevent the slip of the low brake L / B.
Therefore, step S84 corresponds to the prime mover control means in the present invention.

  However, when it is determined in step S83 that Tb ≦ Tbo (low brake L / B does not slip), the motor torque correction amount of the motor / generator MG2 is 0 because step S84 is skipped, and the target in step S84 The motor torque tTmg2 is not substantially corrected.

The above operation will be described below with reference to FIG. 22 showing an operation time chart under the same conditions as in FIG.
When the driver requests an increase in driving force at the instant t1 and increases the accelerator opening APO stepwise as shown in the figure, the target driving force tFd of the vehicle rises as indicated by the wavy line and the target line pressure tP Rises stepwise as indicated by the wavy line.
The actual line pressure P is controlled so as to follow the increase in the target line pressure tP, but coincides with the target line pressure tP with a predetermined response delay as shown by the solid line.

As shown in FIG. 21 (a), the required brake torque Tb becomes larger than the actual brake engagement torque capacity Tbo of the low brake L / B due to the actual line pressure P (the low brake L / B slips) during the instant t2 to t3. As shown in FIG. 22, the motor torque Tmg2 of the motor / generator MG2 is given a motor torque correction amount ΔTb opposite to the torque input to the low brake L / B, and the input torque to the low brake L / B. Is reduced to prevent the above slip.
As a result, the required brake torque Tb coincides with the actual brake engagement torque capacity Tbo due to the actual line pressure P, and the required brake torque Tb is larger than the actual brake engagement torque capacity Tbo, as seen in the instant t2 to t3 in FIG. The low brake L / B is maintained even when the accelerator pedal position APO is changed during driving in the intermediate gear ratio fixed mode that is selected by engaging the low brake L / B. It is possible to prevent slipping.

For this reason, the vehicle driving force Fd can be matched with the target driving force tFd as shown by the solid line as in the instants t2 to t3 in FIG. 22, and the driving force is insufficient even during a transient when the accelerator opening APO is changed. There is no loss of durability of the low brake L / B.
Moreover, since this problem can be solved without relying on the increase in the hydraulic margin of the line pressure described above with reference to FIG. 21 (b), the line pressure is constantly higher than necessary during a much longer period of steady operation than during the transition. It is possible to solve the above-mentioned problem without having to make it, and there is no problem that the energy for driving the oil pump is consumed more than necessary and the fuel consumption is not deteriorated.

In step S85 of FIG. 20, the actual clutch engagement torque capacity Tco of the engine clutch E / C obtained from the actual line pressure P estimated in step S82 is calculated.
In step S86, a difference ΔTc = Tc−Tco between the clutch required torque Tc of the engine clutch E / C (step S73) and the actual clutch engagement torque capacity Tco is calculated, and Tc> It is checked whether Tco (engine clutch E / C slips) or Tc ≦ Tco (engine clutch E / C does not slip).
Therefore, step S86 corresponds to the friction element slip detection means in the present invention.

When it is determined in step S86 that Tc> Tco (engine clutch E / C slips), in step S87, the gear ratio of the deviation ΔTc is added to the target motor torque tTmg1 of the motor / generator MG1 obtained in step S28 of FIG. Add proportional part A.
The addition of the gear ratio proportional portion A of the deviation ΔTc makes the output torque unchanged by compensating for the reduction in transmission output torque that occurs in the next step S88 due to the engine torque limit for preventing slippage of the engine clutch E / C. Leads to maintenance.
In step S88, in order to reduce the engine torque input to the engine clutch E / C so as not to slip, the target engine torque tTe obtained in step S24 in FIG. It is limited to the actual clutch engagement torque capacity Tco of the clutch E / C.
Therefore, step S87 and step S88 correspond to the prime mover control means in the present invention.

  However, when it is determined in step S86 that Tc ≦ Tco (the engine clutch E / C does not slip), step S77 and step S88 are skipped, so the motor torque correction amount of motor / generator MG1 and the torque correction of engine ENG The amount is 0, and the correction of the target motor torque tTmg1 and the correction of the target engine torque tTe in step S87 and step S88 are substantially not performed.

The above operation will be described below with reference to FIG. 24 which shows an operation time chart under the same conditions as in FIG.
When the driver requests an increase in driving force at the instant t1 and increases the accelerator opening APO stepwise as shown in the figure, the target driving force tFd of the vehicle rises as indicated by the wavy line and the target line pressure tP Rises stepwise as indicated by the wavy line.
The actual line pressure P is controlled so as to follow the increase in the target line pressure tP, but coincides with the target line pressure tP with a predetermined response delay as shown by the solid line.

As shown in FIG. 23 (a), the required clutch torque Tc becomes larger than the actual clutch engagement torque capacity Tco of the engine clutch E / C due to the actual line pressure P (the engine clutch E / C slips) during instants t2 to t3. As a result of the engine torque Te being limited to the actual clutch engagement torque capacity Tco of the engine clutch E / C, the wavy line level corresponding to the level shown in FIG. 23 (a) is lowered to the level shown by the solid line.
As a result, the input torque to the engine clutch E / C is reduced, so that the required clutch torque Tc of the engine clutch E / C matches the actual clutch engagement torque capacity Tco as shown in the figure, and therefore Tc> Tco. The engine clutch E / C slips even when the accelerator pedal position APO is changed during traveling in the fixed intermediate gear ratio mode that is selected by engaging the engine clutch E / C. Can be prevented.

Therefore, the vehicle driving force Fd can be matched with the target driving force tFd as shown by the solid line as in the instants t2 to t3 in FIG. 24, and the driving force is insufficient even when the accelerator opening APO is changed. There is no reduction in the durability of the engine clutch E / C.
Moreover, since this problem can be solved without relying on the increase in the hydraulic margin of the line pressure described above with reference to FIG. 23 (b), the line pressure is always higher than necessary during a much longer period of steady state than in the transient state. It is possible to solve the above-mentioned problem without having to make it, and there is no problem that the energy for driving the oil pump is consumed more than necessary and the fuel consumption is not deteriorated.

By the way, the reduction of the engine torque performed to prevent the engine clutch E / C from slipping causes a reduction in transmission output torque, which adversely affects the running performance of the vehicle.
Therefore, in this embodiment, in order to add the gear ratio proportional portion A of the deviation ΔTc between the required clutch torque Tc of the engine clutch E / C and the actual clutch engagement torque capacity Tco to the target motor torque tTmg1 of the motor / generator MG1.
The motor torque Tmg1 of the motor / generator MG1 is increased from the level indicated by the wavy line, which is the same level as shown in FIG. 23 (a), to the level indicated by the solid line, which is performed to prevent the engine clutch E / C from slipping. Further, the transmission output torque is not reduced due to the reduction of the engine torque, and the problem that the running performance of the vehicle is adversely affected can be avoided.

25 to 29 show the high and low brake HL / B and the engine clutch that should be engaged when the hybrid transmission is in the high gear ratio fixed mode represented by the collinear diagram of FIG. 11 when the mode is selected. An embodiment in which E / C slip is prevented will be described.
First, when the hybrid transmission is in the high gear ratio fixed mode, the slip occurrence state of the high & low brake HL / B when the measures of the present invention are not taken will be described below based on FIG. 26 (a). Further, the slip occurrence state of the engine clutch E / C is as described below with reference to FIG.

  These Fig. 26 (a) and Fig. 28 (a) are necessary to produce the line pressure P and the required operating pressure according to the required engagement force of the high and low brakes HL / B and engine clutch E / C at steady state. It is an operation | movement time chart at the time of controlling a line pressure so that it may become a minimum hydraulic pressure.

First, referring to FIG. 26 (a), the slip occurrence state of the high & low brake HL / B will be described.
When the driver requests an increase in driving force at the instant t1 and increases the accelerator opening APO stepwise as shown in the figure, the target driving force tFd of the vehicle changes with time as indicated by the wavy line and the target line pressure tP rises stepwise as shown by the wavy line.
The actual line pressure P is controlled so as to follow the increase in the target line pressure tP, but coincides with the target line pressure tP with the predetermined response delay as shown by the solid line.

The target brake engagement torque capacity tTbo of the high and low brake HL / B corresponding to the target line pressure tP is as shown by the thick wavy line, and the actual brake engagement torque capacity Tbo of the high & low brake HL / B with the actual line pressure P Is as shown by a thick solid line.
The target brake torque tTb of the high & low brake HL / B that accompanies the change in the accelerator opening APO is as shown by a thin wavy line, and the required brake torque of the high & low brake HL / B given to it with a predetermined characteristic. If Tb is as shown by a thin solid line,
At the instant t2 to t3, the actual brake engagement torque capacity Tbo is insufficient with respect to the required brake torque Tb, and the high & low brake HL / B slips.

  The slip of the high & low brake HL / B generates the rotational speed Nmg1 of the motor / generator MG1 and causes insufficient reaction force so that the vehicle driving force Fd cannot be matched with the target driving force tFd as shown by the solid line. In the transition, the driving performance is reduced due to insufficient driving force, and the durability of the high & low brake HL / B is reduced.

To solve this problem, the target line pressure tP is indicated by the arrow in Fig. 26 (b) so that the required operating pressure corresponding to the required fastening force of the high & low brake HL / B can be created even during transition. It is conceivable that the oil pressure margin is increased by always increasing the oil pressure margin as compared with FIG. 26 (a) so that the actual line pressure P changes with time as indicated by a thick solid line.
In this case, the target brake engagement torque capacity tTbo corresponding to the target line pressure tP is raised in the direction of the arrow as indicated by a thick wavy line, and the actual brake engagement torque capacity Tbo due to the actual line pressure P is indicated by a thick solid line. Even during the transition between the instants t1 and t3, the actual brake engagement torque capacity Tbo does not become insufficient with respect to the required brake torque Tb, and slipping of the high & low brake HL / B can be prevented.

  However, in this case, the line pressure will always be higher than necessary during steady state, which is much longer than that at the time of transition, and the energy to drive the oil pump will be consumed more than necessary, resulting in deterioration of fuel consumption due to energy loss. Invite.

Next, with reference to FIG. 28 (a), the slip occurrence state of the engine clutch E / C will be described.
When the driver requests an increase in driving force at the instant t1 and increases the accelerator opening APO stepwise as shown in the figure, the target driving force tFd of the vehicle changes with time as indicated by the wavy line and the target line pressure tP rises stepwise as shown by the wavy line.
The actual line pressure P is controlled so as to follow the increase in the target line pressure tP, but coincides with the target line pressure tP with the predetermined response delay as shown by the solid line.

The target clutch engagement torque capacity tTco of the engine clutch E / C corresponding to the target line pressure tP is as shown by a thick wavy line, and the actual clutch engagement torque capacity Tco of the engine clutch E / C by the actual line pressure P is a thick solid line. It is the time to show.
Further, the target clutch torque tTc of the engine clutch E / C accompanying the change in the accelerator opening APO is as shown by a thin wavy line, whereas the required clutch torque Tc of the engine clutch E / C given with a predetermined characteristic is a thin solid line If it is as shown in
At the instant t2 to t3, the actual clutch engagement torque capacity Tco becomes insufficient with respect to the required clutch torque Tc, and the engine clutch E / C slips.

  The slip of the engine clutch E / C increases the engine clutch E / C rotation speed (engine rotation speed Ne) (i.e., causes the engine to blow idle) and causes a shortage of torque from the engine, resulting in a decrease in vehicle driving force Fd. As indicated by the solid line, it cannot be made to coincide with the target driving force tFd, resulting in a decrease in running performance due to insufficient driving force and a decrease in durability of the engine clutch E / C during the transition.

In order to solve this problem, the target line pressure tP is generated as shown by the arrow in FIG. 28 (b) so that the required operating pressure corresponding to the required engagement force of the engine clutch E / C can be created even during transition. It is conceivable that the oil pressure margin is increased by always increasing the oil pressure margin from (a) so that the actual line pressure P changes over time as indicated by a thick solid line.
In this case, the target clutch engagement torque capacity tTco corresponding to the target line pressure tP is raised in the direction of the arrow as indicated by a thick wavy line, and the actual clutch engagement torque capacity Tco due to the actual line pressure P is increased as indicated by a thick solid line. Even during the transition between the instants t1 and t3, the actual clutch engagement torque capacity Tco does not become insufficient with respect to the required clutch torque Tc, and the slip of the engine clutch E / C can be prevented.

  However, in this case, the line pressure will always be higher than necessary during steady state, which is much longer than that at the time of transition, and the energy to drive the oil pump will be consumed more than necessary, resulting in deterioration of fuel consumption due to energy loss. Invite.

  Therefore, in this embodiment, the slip at the time of transition of the high & low brake HL / B described above with reference to FIG. 26 (a) and the slip at the time of transition of the engine clutch E / C described above with reference to FIG. FIG. 25 shows the line pressure control in step S49 in FIG. 14 to be executed in the high gear ratio fixed mode instead of preventing the increase in the line pressure described above with reference to FIGS. 26 (b) and 28 (b). At the same time, in the control program shown in FIG. 25, the anti-slip control during transient of the high & low brake HL / B and engine clutch E / C is executed as follows.

In step S91, the required brake torque Tb of the high & low brake HL / B to be given with a predetermined characteristic is obtained from the target brake torque tTb of the high & low brake HL / B obtained in step S26 of FIG.
In step S92, the required hydraulic pressure of the high & low brake HL / B that realizes the required brake torque Tb is calculated.
In calculating the required oil pressure, the required oil pressure can be calculated with reference to a torque / oil pressure map that can be obtained in the same manner as described above.

In step S93, the required clutch torque of the engine clutch E / C to be given with a predetermined characteristic is obtained from the target clutch torque of the engine clutch E / C obtained in step S26 of FIG.
In step S94, the required hydraulic pressure of the engine clutch E / C that realizes the required clutch torque is calculated.
In calculating the required oil pressure, the required oil pressure can be calculated with reference to a torque / oil pressure map that can be obtained in the same manner as described above.

In step S95, the required clutch torque of the high clutch H / C to be given with a predetermined characteristic is obtained from the target clutch torque of the high clutch H / C obtained in step S26 of FIG.
In step S96, the required hydraulic pressure of the high clutch H / C that realizes the required clutch torque is calculated.
In calculating the required oil pressure, the required oil pressure can be calculated with reference to a torque / oil pressure map that can be obtained in the same manner as described above.

In step S97, the required hydraulic pressure for the high and low brakes HL / B is higher than the required hydraulic pressure for the engine clutch E / C and whether it is higher than the required hydraulic pressure for the high clutch H / C. Judgment is made as to whether or not the required hydraulic pressure of the high & low brake HL / B is the highest.
Otherwise, in step S98, the required hydraulic pressure of the engine clutch E / C among the required hydraulic pressures of the friction elements depends on whether the required hydraulic pressure of the engine clutch E / C is higher than the required hydraulic pressure of the high clutch H / C. It is determined whether or not it is the highest. Otherwise, it is determined that the required hydraulic pressure of the high clutch H / C is the highest among the required hydraulic pressures of the friction elements.

If it is determined in step S97 that the required hydraulic pressure of the high & low brake HL / B is the highest, in step S99, the required hydraulic pressure of the high & low brake HL / B is set to the target line pressure tP. In addition to the low brake HL / B, the minimum target line pressure tP required to prevent all friction elements from slipping in a steady state is determined.
If it is determined in step S98 that the required hydraulic pressure of the engine clutch E / C is the highest, in step S100, the required hydraulic pressure of the engine clutch E / C is set to the target line pressure tP. Of course, the minimum target line pressure tP required to prevent all the friction elements from slipping in the steady state is determined.
When it is determined in step S98 that the required hydraulic pressure of the high clutch H / C is the highest, in step S101, the required hydraulic pressure of the high clutch H / C is set to the target line pressure tP. Of course, the minimum target line pressure tP required to prevent all the friction elements from slipping in the steady state is determined.
Therefore, step S97 to step S101 correspond to the line pressure control means in the present invention.

In step S102, the actual line pressure P is estimated by applying a second order delay to the target line pressure tP determined as described above, and the actual brake engagement torque of the high & low brake HL / B obtained by this actual line pressure P. Calculate the capacity Tbo.
In step S103, a deviation ΔTb = Tb−Tbo between the required brake torque Tb (step S91) of the high and low brake HL / B and the actual brake engagement torque capacity Tbo is calculated, and the positive / negative judgment of the deviation ΔTb is performed. Check whether Tb> Tbo (high & low brake HL / B slips) or Tb ≦ Tbo (high & low brake HL / B does not slip).
Accordingly, step S103 corresponds to the friction element slip detection means in the present invention.

When it is determined in step S103 that Tb> Tbo (high & low brake HL / B slips), in step S104, the deviation ΔTb is added to the target motor torque tTmg1 of the motor / generator MG1 obtained in step S28 of FIG. Is added.
The addition of the deviation ΔTb leads to giving the motor / generator MG1 a motor torque in the opposite direction to the torque inputted to the high / low brake HL / B to prevent slippage.
Therefore, step S104 corresponds to the prime mover control means in the present invention.

  However, when it is determined in step S103 that Tb ≦ Tbo (high & low brake HL / B does not slip), the motor torque correction amount of the motor / generator MG1 is 0 because step S104 is skipped, and in step S104. The target motor torque tTmg1 is not substantially corrected.

The above operation will be described below with reference to FIG. 27 showing an operation time chart under the same conditions as in FIG.
When the driver requests an increase in driving force at the instant t1 and increases the accelerator opening APO stepwise as shown in the figure, the target driving force tFd of the vehicle rises as indicated by the wavy line and the target line pressure tP Rises stepwise as indicated by the wavy line.
The actual line pressure P is controlled so as to follow the increase in the target line pressure tP, but coincides with the target line pressure tP with a predetermined response delay as shown by the solid line.

As shown in FIG. 26 (a), the required brake torque Tb becomes larger than the actual brake engagement torque capacity Tbo of the high & low brake HL / B due to the actual line pressure P (the high & low brake HL / B slips). During t2 to t3, the motor torque Tmg1 of the motor / generator MG1 is given a motor torque correction amount ΔTb opposite to the torque input to the high & low brake HL / B as shown in FIG. Reduce the input torque to the brake HL / B to prevent the above slip.
As a result, the required brake torque Tb coincides with the actual brake engagement torque capacity Tbo due to the actual line pressure P, and the required brake torque Tb is larger than the actual brake engagement torque capacity Tbo, as seen during the instants t2 to t3 in FIG. High and low brakes even during transitions when the accelerator opening APO is changed while driving in the fixed high gear ratio mode, which is selected by engaging the high and low brakes HL / B. HL / B can be prevented from slipping.

Therefore, the vehicle driving force Fd can be matched with the target driving force tFd as shown by the solid line as in the instants t2 to t3 in FIG. 27, and the driving force is insufficient even at the transient time when the accelerator opening APO is changed. There is no loss of durability of the high and low brake HL / B.
Moreover, since this problem can be solved without relying on the increase in the hydraulic margin of the line pressure described above with reference to FIG. 26 (b), the line pressure is constantly higher than necessary during a much longer period of steady state than in the transient state. It is possible to solve the above-mentioned problem without having to make it, and there is no problem that the energy for driving the oil pump is consumed more than necessary and the fuel consumption is not deteriorated.

In step S105 in FIG. 25, the actual clutch engagement torque capacity Tco of the engine clutch E / C obtained from the actual line pressure P estimated in step S102 is calculated.
In step S106, a deviation ΔTc = Tc−Tco between the clutch required torque Tc of the engine clutch E / C (step S93) and the actual clutch engagement torque capacity Tco is calculated, and Tc> It is checked whether Tco (engine clutch E / C slips) or Tc ≦ Tco (engine clutch E / C does not slip).
Accordingly, step S106 corresponds to the friction element slip detection means in the present invention.

When it is determined in step S106 that Tc> Tco (engine clutch E / C slips), in step S107, the gear ratio of the deviation ΔTc is added to the target motor torque tTmg2 of the motor / generator MG2 obtained in step S29 of FIG. Add proportional part A.
The addition of the gear ratio proportional portion A of the deviation ΔTc compensates for the reduction in transmission output torque that occurs in the next step S108 due to the engine torque limit for preventing slipping of the engine clutch E / C, and makes the output torque unchanged. Leads to maintenance.
In step S108, in order to reduce the engine torque input to the engine clutch E / C so as to prevent the engine clutch E / C from slipping, the target engine torque tTe obtained in step S24 in FIG. It is limited to the actual clutch engagement torque capacity Tco of the clutch E / C.
Therefore, step S107 and step S108 correspond to the prime mover control means in the present invention.

  However, when it is determined in step S106 that Tc ≦ Tco (the engine clutch E / C does not slip), step S107 and step S108 are skipped, so the motor torque correction amount of the motor / generator MG2 and the torque correction of the engine ENG The amount is 0, and the correction of the target motor torque tTmg2 and the correction of the target engine torque tTe in steps S107 and S108 are not substantially performed.

The above operation will be described below with reference to FIG. 29 showing an operation time chart under the same conditions as in FIG.
When the driver requests an increase in driving force at the instant t1 and increases the accelerator opening APO stepwise as shown in the figure, the target driving force tFd of the vehicle rises as indicated by the wavy line and the target line pressure tP Rises stepwise as indicated by the wavy line.
The actual line pressure P is controlled so as to follow the increase in the target line pressure tP, but coincides with the target line pressure tP with a predetermined response delay as shown by the solid line.

As shown in FIG. 28 (a), the required clutch torque Tc becomes larger than the actual clutch engagement torque capacity Tco of the engine clutch E / C due to the actual line pressure P (the engine clutch E / C slips) during instants t2 to t3. As a result of the engine torque Te being limited to the actual clutch engagement torque capacity Tco of the engine clutch E / C, the wavy line level corresponding to the level shown in FIG. 28 (a) is lowered to the level shown by the solid line.
As a result, the input torque to the engine clutch E / C is reduced, so that the required clutch torque Tc of the engine clutch E / C matches the actual clutch engagement torque capacity Tco as shown in the figure, and therefore Tc> Tco. The engine clutch E / C slips even when the accelerator opening APO is changed during driving in the high gear ratio fixed mode that is selected by engaging the engine clutch E / C. Can be prevented.

Therefore, the vehicle driving force Fd can be matched with the target driving force tFd as shown by the solid line as in the instants t2 to t3 in FIG. 29, and the driving force is insufficient even when the accelerator opening APO is changed. There is no reduction in the durability of the engine clutch E / C.
Moreover, since this problem can be solved without relying on the increase in the hydraulic margin of the line pressure described above with reference to FIG. 28 (b), the line pressure is always higher than necessary during a much longer period of steady operation than in the transient state. It is possible to solve the above-mentioned problem without having to make it, and there is no problem that the energy for driving the oil pump is consumed more than necessary and the fuel consumption is not deteriorated.

By the way, the reduction of the engine torque performed to prevent the engine clutch E / C from slipping causes a reduction in transmission output torque, which adversely affects the running performance of the vehicle.
Therefore, in the present embodiment, in order to add the gear ratio proportional portion A of the deviation ΔTc between the required clutch torque Tc of the engine clutch E / C and the actual clutch engagement torque capacity Tco to the target motor torque tTmg2,
The motor torque Tmg2 of the motor / generator MG2 is increased from the level indicated by the wavy line, which is the same level as shown in FIG. 28 (a), to the level indicated by the solid line. Further, the transmission output torque is not reduced due to the reduction of the engine torque, and the problem that the running performance of the vehicle is adversely affected can be avoided.

It is a basic diagram which shows the power train of the vehicle carrying the hybrid transmission which can apply the slip prevention apparatus of the friction element by this invention. FIG. 2 is a collinear diagram of the hybrid transmission shown in FIG. It is a flowchart which shows the main routine of the shift control which the controller of the hybrid transmission performs. 7 is a flowchart showing motor torque control of a motor / generator for preventing a friction element from slipping in a low gear ratio fixed mode in a subroutine in the main routine. FIG. 1 is an operation time chart of the hybrid transmission when the measures of the present invention are not taken for the hybrid transmission shown in FIG. 1 and FIG. 2; These are operation | movement time charts when the oil pressure margin of a line pressure is enlarged so that the slip of a friction element may be prevented. 4 is an operation time chart showing a slip prevention action of a friction element when the countermeasure of the present invention is applied to the hybrid transmission. It is a skeleton figure which shows the other example of the hybrid transmission which can apply the slip prevention apparatus of the friction element by this invention. It is an alignment chart of the hybrid transmission. FIG. 6 is an alignment chart when the hybrid transmission is in a low gear ratio fixed mode. FIG. 6 is an alignment chart when the hybrid transmission is in an intermediate gear ratio fixed mode. FIG. 6 is an alignment chart when the hybrid transmission is in a high gear ratio fixed mode. It is a block diagram according to function which shows the control system of the hybrid transmission. It is a flowchart which shows the main routine of the control program which the hybrid controller in the same control system performs. It is a flowchart which shows the subroutine regarding the line pressure control in the main routine. 3 is a flowchart showing line pressure control and friction element slip prevention control when the hybrid transmission is in a low gear ratio fixed mode. FIG. 6 is an operation time chart of the hybrid transmission in the low gear ratio fixed mode when the countermeasure of the present invention is not taken for the hybrid transmission, (a) is an operation time chart showing a slip occurrence state of the high & low brake, (B) is an operation time chart when the oil pressure margin of the line pressure is increased so as to prevent the slip of the high & low brake. 6 is an operation time chart showing the anti-slip action of the high & low brake in the low gear ratio fixed mode when the countermeasure of the present invention is applied to the hybrid transmission. FIG. 6 is an operation time chart of the hybrid transmission in the low gear ratio fixed mode when the countermeasure of the present invention is not taken for the hybrid transmission, (a) is an operation time chart showing a state of occurrence of engine clutch slip, (b) ) Is an operation time chart when the line pressure hydraulic margin is increased to prevent engine clutch slip. 3 is an operation time chart showing an anti-slip action of an engine clutch in a low gear ratio fixed mode when the hybrid transmission is subjected to the measures of the present invention. 4 is a flowchart showing line pressure control and friction element slip prevention control when the hybrid transmission is in an intermediate gear ratio fixed mode. FIG. 6 is an operation time chart of the hybrid transmission in the intermediate transmission ratio fixed mode when the countermeasure of the present invention is not taken for the hybrid transmission, (a) is an operation time chart showing a low brake slip occurrence state, (b) ) Is an operation time chart when the hydraulic margin of the line pressure is increased to prevent low brake slip. 6 is an operation time chart showing an anti-slip action of a low brake in an intermediate gear ratio fixed mode when the hybrid transmission is subjected to the measures of the present invention. FIG. 7 is an operation time chart of the hybrid transmission in the intermediate transmission ratio fixed mode when the countermeasure of the present invention is not taken for the hybrid transmission, (a) is an operation time chart showing a situation of occurrence of engine clutch slip, (b) ) Is an operation time chart when the line pressure hydraulic margin is increased to prevent engine clutch slip. 6 is an operation time chart showing an anti-slip action of an engine clutch in an intermediate gear ratio fixed mode when the hybrid transmission is subjected to the measures of the present invention. 4 is a flowchart showing line pressure control and friction element slip prevention control when the hybrid transmission is in a high gear ratio fixed mode. FIG. 6 is an operation time chart of the hybrid transmission in the high gear ratio fixed mode when the countermeasure of the present invention is not taken for the hybrid transmission, (a) is an operation time chart showing a slip occurrence state of the high & low brake, (B) is an operation time chart when the oil pressure margin of the line pressure is increased so as to prevent the slip of the high & low brake. 6 is an operation time chart showing the anti-slip action of the high & low brake in the high gear ratio fixed mode when the hybrid transmission is subjected to the measures of the present invention. FIG. 6 is an operation time chart of the hybrid transmission in the high gear ratio fixed mode when the countermeasure of the present invention is not taken for the hybrid transmission, (a) is an operation time chart showing a situation of occurrence of engine clutch slip, (b) ) Is an operation time chart when the line pressure hydraulic margin is increased to prevent engine clutch slip. 4 is an operation time chart showing an anti-slip action of an engine clutch in a high gear ratio fixed mode when the hybrid transmission is subjected to the measures of the present invention.

Explanation of symbols

ENG engine (main power source)
10,20 Transmission mechanism
12 driving wheels
MG1 1st motor / generator
MG2 Second motor / generator
21 Input shaft
22 Output shaft
G1 1st planetary gear set
G2 2nd planetary gear set
G3 3rd planetary gear set
S1, S2, S3 Sun gear
R1, R2, R3 ring gear
C1, C2, C3 carrier
HL / B High & Low brake (friction element)
H / C high clutch (friction element)
L / B low brake (friction element)
E / C engine clutch (friction element)
S / C series clutch
MG1 / C motor / generator clutch
31 Hybrid controller
32 Engine controller
33 Throttle valve
34 Fuel injector
35 Ignition system
36 Motor controller
37 Motor controller
38 Inverter
39 Inverter
40 battery
41 Battery controller

Claims (5)

In addition to the main power source as a prime mover, a motor / generator is provided, and the power from the main power source and the motor / generator is transmitted and output by a speed change mechanism, and the transmission state by the speed change mechanism is determined by fastening / release of a friction element. In the hybrid transmission designed to be determined,
Friction element slip detection means for detecting that the friction element cannot maintain the fastening state due to a sudden change in torque from the prime mover;
When the means detects the inability to hold the friction element, the main power source or the main power source, or the like, reduces the torque to the friction element that is unable to maintain the engaged state while preventing the torque from the hybrid transmission from changing. An anti-slip device for a friction element in a hybrid transmission, comprising a motor / generator, or a prime mover control means for controlling the torque of both. In the anti-slip device for the friction element in the hybrid transmission according to claim 1,
When the friction element slip detection means detects that the brake cannot be held as a friction element, the prime mover control means uses the torque of the motor / generator or the main power source fixed by the brake as the torque to the brake. An anti-slip device for a friction element in a hybrid transmission that is controlled to be reduced. The slip prevention device for a friction element in the hybrid transmission according to claim 1 or 2,
When the friction element slip detection means detects that the clutch cannot be held as a friction element, the prime mover control means uses the torque of the main power source or motor / generator coupled to the components of the transmission mechanism by the clutch. A motor / generator or main power source that is not involved in the clutch so as to control the torque to the clutch to be reduced and to compensate for a change in torque from the hybrid transmission that changes due to a reduction in the torque to the clutch. An anti-slip device for a friction element in a hybrid transmission that controls the torque of the motor. In the hybrid transmission of any one of Claims 1-3, the slip prevention apparatus of the friction element in a hybrid transmission,
Hybrid transmission is performed so as to estimate the steady engagement torque of each of the friction elements by estimating the steady engagement torque of each of the friction elements, and to prevent all the friction elements from slipping in a steady state based on the highest hydraulic pressure value of these engagement necessary oil pressures. An anti-slip device for a friction element in a hybrid transmission, characterized in that line pressure control means for determining a line pressure that is a control source pressure of the machine is provided. The anti-slip device for a friction element in the hybrid transmission according to claim 4,
The line pressure control means calculates the steady engagement torque of each friction element from the torque of the main power source and the motor / generator and the output torque of the hybrid transmission, and when the steady engagement torque and the friction element are released. The relationship between the hydraulic pressure and the frictional element fastening possible torque is stored from the relationship with the hydraulic pressure, and the individual steady required fastening hydraulic pressure of the friction element is estimated by interpolation using the stored relationship and the past relationship. An anti-slip device for a friction element in a hybrid transmission.

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