Vehicle drive control system and method

A drive control system for an automotive vehicle includes an engine, an automatic transmission, a torque converter, a start clutch, an oil pump, a hydraulic pressure control unit, an engine control unit, an automatic engine stop/restart control unit and a torque transmission control unit configured to calculate target start clutch engagement time and torque based on an accelerator opening of the vehicle, calculate a target engine torque based on the target clutch engagement torque and cause the hydraulic control unit and the engine control unit to control the engagement state of the start clutch and the output torque of the engine according to the target clutch engagement time and torque and the target engine torque at a restart of the engine.

CROSS-REFERENCE TO RELATED PATENT APPLICATIONS

Japan Priority Application 2005-036410, filed Feb. 14, 2005 including the specification, drawings, claims and abstract, is incorporated herein by reference in its entirety.

BACKGROUND OF THE INVENTION

The present invention relates to a drive control system for an automotive vehicle having an automatic engine stop/restart function (also called an idle stop control function), particularly of the kind suitable for use with a belt-type continuously variable automatic transmission. The present invention also relates to a drive control method for an automotive vehicle.

Japanese Laid-Open Patent Publication No. 2000-266172 discloses a vehicle drive control system that has an automatic engine stop/restart function and performs clutch engagement control to shift a start clutch of an automatic transmission device from a precharge phase into an engagement phase when the engine speed reaches a predetermined level at an engine restart after an idle stop.

SUMMARY OF THE INVENTION

The amount of hydraulic oil supplied to the start clutch depends on various transmission hydraulic control parameters such as oil pump characteristics (e.g. the dependency of pump volumetric efficiency on pump operation speed) and valve leaks. Further, the oil pump and valve operation performance varies with time to exert a great effect on the amount of hydraulic oil supplied to the start clutch. It is thus difficult to judge whether the start clutch has secured a sufficient hydraulic pressure by predicting the completion of the clutch precharge phase based on the engine speed.

The engine speed is generally held at an idle speed until the start clutch comes into engagement. If the start clutch shifts into the engagement phase without securing a sufficient hydraulic pressure, it takes a long time to complete the engagement of the start clutch. In such a case, the engine speed cannot be easily increased even under full-throttle acceleration so that there arises a large vehicle starting time lag.

It is therefore an object of the present invention to provide a vehicle drive control system capable of enabling a vehicle to make a smooth start, even under insufficient transmission hydraulic pressure conditions, upon an engine restart after an idle stop.

It is also an object of the present invention to provide a drive control method for an automotive vehicle.

According to a first aspect of the present invention, there is provided a drive control system for an automotive vehicle, comprising: an engine; an automatic transmission; a torque converter disposed between the engine and the automatic transmission; an oil pump driven in synchronism with the engine; a start clutch engaged under a hydraulic oil pressure discharged from the oil pump to allow torque transmission from the engine to the automatic transmission via the torque converter at a start of the vehicle; an engine control unit that controls an output torque of the engine; a hydraulic pressure control unit that regulates the hydraulic pressure to control an engagement state of the start clutch; an automatic engine stop/restart control unit that stops the engine upon satisfaction of certain vehicle conditions and to restart the engine upon dissatisfaction of the certain vehicle conditions; and a torque transmission control unit configured to: calculate target engagement time and torque of the start clutch based on an accelerator opening of the vehicle; calculate a target engine torque based on the target clutch engagement torque; and cause the hydraulic control unit and the engine control unit to control the engagement state of the start clutch and the output torque of the engine according to the target clutch engagement time and torque and the target engine torque, at a restart of the engine.

According to a second aspect of the present invention, there is provided a drive control system for an automotive vehicle, the vehicle having an engine, an automatic transmission, a torque converter and a start clutch engaged to transmit an output torque of the engine to the automatic transmission via the torque converter at a start of the vehicle, the drive control system comprising: means for automatically stopping the engine upon satisfaction of certain vehicle conditions and restarting the engine upon dissatisfaction of the certain vehicle conditions; means for controlling an output torque of the engine; means for supplying a hydraulic pressure to engage the start clutch at a restart of the engine; means for detecting an accelerator opening of the vehicle; means for determining a target clutch engagement time based on the accelerator opening; means for determining target clutch engagement pressure and torque based on the accelerator opening and the target clutch engagement time; means for determining a target engine torque based on the target clutch engagement torque; means for adjusting the hydraulic pressure to the target clutch engagement pressure; and means for adjusting the output torque of the engine to the target engine torque.

According to a third aspect of the present invention, there is provided a drive control method for an automotive vehicle, the vehicle having an engine, an automatic transmission, a torque converter and a start clutch engaged to transmit an output torque of the engine to the automatic transmission via the torque converter at a start of the vehicle, the drive control method comprising: stopping the engine upon satisfaction of certain vehicle conditions; after said stopping, restarting the engine upon dissatisfaction of the certain vehicle conditions; upon said restarting, supplying a hydraulic pressure to engage the start clutch; detecting an accelerator opening of the vehicle; determining a target clutch engagement time based on the accelerator opening; determining target clutch engagement pressure and torque based on the accelerator opening and the target clutch engagement time; determining a target engine torque based on the target clutch engagement torque; adjusting the hydraulic pressure to the target clutch engagement pressure; and adjusting the output torque of the engine to the target engine torque.

The other objects and features of the present invention will also become understood from the following description.

DESCRIPTION OF THE EMBODIMENTS

The present invention will be described below by way of the following first and second embodiments, in which like parts and portions are designated by like reference numerals to thereby omit repeated explanations thereof.

A drive control system for an automotive vehicle according to the first embodiment of the present invention will be now explained below with reference toFIGS. 1 to 16.

As shown inFIG. 1, the vehicle drive control system of the first embodiment includes an engine11, an automatic transmission device with a torque converter1, a forward/reverse changeover unit20, a belt-type continuously variable transmission (CVT)3, a hydraulic control valve unit7and an oil pump8, a starter motor19a, a CVT control unit10with a torque transmission controller10a, an engine control unit (ECU)18, an idle stop control unit12and an idle stop switch13.

The torque converter1is connected to an output shaft19bof the engine11and has a lock-up clutch2to establish a connection between the engine11and the CVT3via the forward/reverse changeover unit20.

The oil pump8is also connected to the engine output shaft19bso as to operate in synchronism with the engine11to supply a hydraulic oil to the hydraulic control valve unit7.

The forward/reverse changeover unit20has a planetary gear train consisting of a ring gear21connected to a turbine shaft (output shaft)1aof the torque converter1, a sun gear23connected to an input shaft1bof the CVT3and a pinion carrier22disposed between the ring gear21and the sun gear23, a reverse brake24for fixing the pinion carrier22to a transmission case and a forward clutch25(as a start clutch) for connecting the pinion carrier22to the transmission input shaft1b. The forward clutch25is engaged at a start of the vehicle to change its engagement state in response to the supply of a clutch engagement pressure Pc from the oil pump8through the hydraulic control valve unit7and thereby allow torque transmission from the engine11to the CVT3via the torque converter1.

The CVT3has a primary pulley30amounted on an end of the transmission input shaft1b, a secondary pulley30bmounted on a driven shaft38of the CVT3and a power transmission belt34drivingly connecting the primary pulley30ato the secondary pulley30bas shown inFIG. 1.

The primary pulley30aincludes a fixed conical pulley disc31rotated together with the transmission input shaft1b, a movable conical pulley disc32opposed to the pulley disc31to define a V-shaped pulley groove between the pulley discs31and32, and a cylinder chamber33. The pulley disc32is moved in the axial direction of the transmission input shaft1baccording to an oil pressure supplied to the cylinder chamber33(referred to as a “primary pulley clamping pressure Ppri”), thereby adjusting the width of the primary pulley groove and changing the contact radius between the belt34and the primary pulley30a.

The secondary pulley30bhas a fixed conical pulley disc35rotated together with the transmission driven shaft38, a movable conical pulley disc36opposed to the pulley disc35to define a V-shaped pulley groove between the pulley discs35and36, and a cylinder chamber37. The pulley disc36is moved in the axial direction of the transmission driven shaft38according to an oil pressure supplied to the cylinder chamber37(referred to as a “secondary pulley clamping pressure Psec”), thereby adjusting the width of the secondary pulley groove and changing the contact radius between the belt34and the secondary pulley30b.

Although not shown in the drawings, the transmission driven shaft38is connected to driving wheels through a driving gear, an idler gear, a final reduction gear and a differential gear.

With such a power train arrangement, a driving torque of the engine11is inputted into the CVT3through the torque converter1and the forward/reverse changeover unit20, transmitted from the primary pulley30ato the secondary pulley30bvia the belt34, and then, outputted to the driving wheels through the driving gear, the idler gear, the final reduction gear and the differential gear. The rotation speed ratio between the primary pulley30aand the secondary pulley30b, i.e., the power transmission ratio of the CVT3is varied continuously by moving the pulley discs32and36of the primary and secondary pulley30aand30band changing the radius of the belt34coming into contact with the pulleys30aand30b. In other words, the power transmission ratio of the CVT3varies by adjusting the primary and secondary pulley clamping pressures Ppri and Psec.

As shown inFIGS. 1 and 2, the vehicle drive control system further includes a primary pulley rotation speed sensor4to detect a rotation speed Npri of the primary pulley30awhich corresponds to an output shaft rotation speed of the forward clutch25, a turbine rotation speed sensor5to detect a rotation speed Ntb of the turbine shaft1a, a secondary pulley rotation speed sensor6to detect a rotation speed Nsec of the secondary pulley30b, an accelerator opening sensor14ato detect the stroke θ of an accelerator pedal of the vehicle (hereinafter referred to as an “accelerator opening”), an oil temperature sensor14bto detect a transmission oil temperature Toil of the vehicle, a steering angle sensor14cto detect an steering wheel angle (hereinafter just referred to as a “steering angle”) of the vehicle, a vehicle speed sensor14dto detect a traveling speed of the vehicle, a line pressure sensor14eto detect a line pressure of the hydraulic control valve unit6(i.e. a source pressure of the pulley clamping pressures Ppri and Psec), a brake switch15to detect the ON-OFF state of a brake pedal of the vehicle and an engine speed sensor to detect an operating speed Ne of the engine11. The detection signals from the accelerator opening sensor14a, the oil temperature sensor14b, the pulley rotation speed sensors4,6, the turbine speed sensor5and the engine speed sensor16are inputted into the CVT control unit10, whereas the detection signals from the steering angle sensor14c, the idle stop switch13, the vehicle speed sensor14dand the brake switch15are inputted into the idle stop control unit12. These sensor signals are mutually communicated between the CVT control unit10and the idle stop control unit12.

The CVT control unit10controls the operations of the hydraulic control valve unit7based on the sensor detection signals so as to adjust the forward clutch engagement pressure Pc and the pulley clamping pressure Ppri and Psec and thereby change the engagement state of the start clutch25and the power transmission ratio of the CVT3according to vehicle driving conditions.

In the first embodiment, the hydraulic control valve unit7includes a pressure regulator valve40, a clutch regulator valve45, a pilot valve50, a lock-up solenoid71, a selector switching solenoid70, a selector switching valve75, a selector control valve80with a return spring81and a valve spool82, a line pressure solenoid72, a manual valve60, a pressure modifier valve73, oil passages90ato90e,91to95,96a,96b,97and98and an orifice90fas shown inFIG. 2.

The pressure regulator valve40receives a hydraulic oil pressure discharged from the oil pump8via the oil passage90aand regulates the pump discharge pressure into the line pressure (the source pressure of the pulley clamping pressures Ppri and Psec). The line pressure is supplied to the pulley cylinder chambers33and37and the pilot valve50via the oil passages90a,90band90e. The line pressure is also supplied to the pressure modifier valve73via the oil passage90b.

The clutch regulator valve45receives a relief pressure from the pressure regulator valve40via the oil passage91and regulates the relief pressure into the forward clutch engagement pressure Pc. Herein, the oil passage91is communicated with the oil passage90bthrough the oil passages90cand90dand the orifice90f. The clutch engagement pressure Pc is supplied to the selector switching valve75and the selector control valve80via the oil passage92.

The pilot valve50regulates the line pressure to supply a constant pilot pressure to the selector switching solenoid70, the lock-up solenoid71and the line pressure solenoid92via the oil passage93.

The selector switching solenoid70is selectively switched between the ON state and the OFF state to supply according to its ON/OFF state a signal pressure to the selector switching valve75via the oil passage94for control of the operations of the selector switching valve75. The lock-up solenoid71supplies a signal pressure to the selector switching valve75via the oil passage95. The signal pressure from the lock-up solenoid71is fed to the selector control valve80through the selector switching valve75when the selector switching solenoid70is ON, and then, fed to a lock-up control valve (not shown) through the selector switching valve75when the selector switching solenoid70is OFF.

The selector control valve80opens by moving the valve spool82against the tension of the return spring81upon receipt of the signal pressure from the lock-up solenoid71and closes by biasing the valve spool82(to the right side ofFIG. 2) under the tension of the return spring81without no signal pressure from the lock-up solenoid71.

The pressure modifier valve73regulates the line pressure into a signal pressure under the control of the line pressure solenoid72and supplies the signal pressure to the pressure regulator valve40and the clutch regulator valve45such that the pressure regulator valve40and the clutch regulator valve45operate upon receipt of the signal pressure from the pressure modifier valve73. In the first embodiment, the signal pressure of the pressure modifier valve73is set higher than those of the solenoids70,71and72so as to obtain an improvement in pressure regulation performance even in a high hydraulic pressure range.

When the forward clutch25shifts from a disengaged state into an engaged state, the selector switching solenoid70and the lock-up solenoid71are switched on to interrupt direct communication between the oil passages92and97. The clutch engagement pressure Pc from the clutch regulator valve45is fed to the selector control valve80via the oil passage96a. The signal pressure from the lock-up solenoid71is concurrently supplied to the selector control valve80. The selector control valve80adjusts the clutch engagement pressure Pc under the signal pressure from the lock-up solenoid71and outputs the adjusted clutch engagement pressure Pc to the selector switching valve75via the oil passage96b. The clutch engagement pressure Pc is then supplied to the forward clutch25through the manual valve60and the oil passages97and98. It should be noted that there is a limit on the maximum attainable level of the clutch engagement pressure Pc when the fine adjustment of the clutch engagement pressure Pc is made by the selector control valve80under the signal from the lock-up solenoid71.

On the other hand, the selector switching solenoid70and the lock-up solenoid71are switched off to allow direct communication between the oil passages42when the forward clutch25is engaged. The clutch engagement pressure Pc is directly passed from the oil passage92to the oil passage97through the selector switching valve75without being adjusted by the selector control valve80, and then, supplied to the forward clutch25through the manual valve60and the oil passages97and98.

The hydraulic pressure control line of the hydraulic control valve unit7is thus changed to adjust the clutch engagement pressure Pc by means of the clutch regulator valve45, the line pressure solenoid72and the pressure modifier valve73after the forward clutch25comes into full engagement and by means of not only the clutch regulator valve45, the line pressure solenoid72and the pressure modifier valve73but also the selector switching solenoid70, the lock-up solenoid71, the selector switching valve75and the selector control valve80before the forward clutch25comes into full engagement (e.g. during the engagement of the forward clutch25for vehicle starting) in the first embodiment. This makes it possible to adjust the clutch engagement pressure Pc more finely in a low hydraulic pressure range under the signal from the lock-up solenoid71during the engagement of the forward clutch25.

The idle stop control unit12performs idle stop control (automatic engine stop/restart control) to decide on the idle stop or restart of the engine11based on the sensor detection signals and output an engine idle stop or restart signal to the CVT control unit10and the engine control unit18.

The idle stop control of the idle stop control unit12is programmed in the following steps shown inFIG. 3in the first embodiment.

At step S101, the idle stop control unit12determines whether the vehicle is in a state of satisfying all of the following idle stop conditions: (a) the idle stop flag is ON; (b) the idle stop switch13is ON; (c) the vehicle speed is zero; (d) the brake switch15is ON; and (e) the steering angle is zero. The idle stop flag is set and cleared according to the logics of other control operations. In the first embodiment, the idle stop flag is set OFF in the case where a desirable engine restart operation cannot be made, more specifically, e.g. when the forward clutch25cannot be engaged under the after-mentioned clutch engagement control or when the starter motor19acannot be driven due to insufficient battery charge. The idle stop switch13is energized upon ignition switch actuation and operated by a vehicle driver to indicate a driver's intention to execute and cancel the idle stop control. In other words, the idle stop control unit12decides on the idle stop of the engine11when the vehicle is in a stop state in which the execution of the idle stop control is requested by the vehicle driver with the brake pedal being held down and the steering angle being set to zero. Further, the idle stop condition (e) is set up to inhibit the idle stop of the engine11during a temporary stop e.g. at a right turn of the vehicle in the first embodiment. If all of the idle stop conditions (a) to (e) are satisfied (Yes at step S101), the control proceeds to step S102. The control is overridden if one or more of the idle stop conditions (a) to (e) are not satisfied (No at step S101).

At step S102, the idle stop control unit12determines whether the transmission device is placed in D range. If the transmission device is in D range (Yes at step S102), the control proceeds to step S103. The control proceeds to step S104if the transmission device is in any other range (No at step S102).

At step S103, the idle stop control unit10determines whether the hydraulic oil temperature Toil is higher than a low temperature limit Tlow and lower than an upper temperature limit Thi, i.e., judges whether the transmission engagement element such as the start clutch25can be supplied with a sufficient engagement pressure. When the oil temperature Toil is not higher than the low temperature limit Tlow, there arises a possibility that the transmission engagement element cannot be supplied with a predetermined amount of hydraulic oil due to oil viscosity drag before the occurrence of complete combustion in the engine11. The volumetric efficiency of the oil pump8decreases due to low oil viscosity drag and the valve leakage increases in a high oil temperature range. There thus arises a possibility that the transmission engagement element cannot be supplied with a predetermined amount of hydraulic oil due to such pump performance deterioration and valve leakage problems before the occurrence of complete combustion in the engine11when the oil temperature Toil exceeds the upper temperature limit Thi. If Tlow<Toil<Thi (Yes at step S103), the control proceeds to step S104. The control goes back to step S102if the temperature condition of Tlow<Toil<Thi is not satisfied (No at step S103).

At step S104, the idle stop control unit12generates the idle stop signal and causes the engine control unit18to stop the engine11.

At step S105, the idle stop control unit12determines whether the brake switch15is OFF, i.e., judges whether there is the driver's intention to restart the engine1on release of the brake pedal. If the brake switch15is turned off (Yes at step S15), the control then proceeds to step S106. If the brake switch15is turned on (No at step S15), the control goes back to step S104.

The idle stop control unit12is also programmed to judge that there is driver's intention to restart the engine1on de-energization of the idle stop switch13even when the brake switch15is ON. This makes it possible to cancel the idle stop control at the request of the vehicle driver in order to e.g. avoid a load on vehicle battery during the idle stop of the engine11and prevent an air conditioning system of the vehicle from becoming disabled even though the vehicle interior is hot. The vehicle drive control can be thus made more responsive to the driver's intention.

At step S106, the idle stop control unit10generates the engine restart signal and causes the engine control unit18to actuate the starter motor19aand restart the engine11.

The idle stop signal from the idle stop control unit12may also be outputted to a brake hill hold control mechanism of the vehicle in order to prevent the vehicle from rolling back on e.g. a slop during the idle stop of the vehicle.

The engine control unit18stops the idling of the engine11upon receipt of the engine idle stop signal and then actuates the starter motor10aand thereby restarts the engine11upon receipt of the engine restart signal.

The starter motor drive control of the engine control unit18is programmed in the following steps shown inFIG. 10Bin the first embodiment.

At step S400, the engine control unit18judges whether complete combustion has occurred in the engine11. The method for judging the occurrence of complete combustion in the engine11is not particularly restricted. In the first embodiment, the judgment on the engine complete combustion is made by determining whether the engine speed Ne is higher than a given speed level. If the engine complete combustion has occurred (Yes at step S400), the control proceeds to step S420. If the engine complete combustion has not occurred (No at step S400), the control proceeds to step S410.

At step S410, the engine control unit18drives the starter motor19a. The control then goes back to step S400.

At step S420, the engine control unit18stops the starter motor19a. The control exits the program.

Herein, the oil pump8comes to a stop upon the idle stop of the engine11. When the oil pump8stops, the hydraulic oil drops out of the oil passages of the hydraulic control valve unit7to cause decreases in the clutch engagement pressure Pc and the pulley clamping pressures Ppri and Psec. The forward clutch25becomes disengaged due to such a decrease in the clutch engagement pressure Pc during the engine idle stop. It is accordingly preferable to supply a predetermined amount of hydraulic oil to the forward clutch25and control the clutch engagement pressure Pc to a desired level at the time of an engine restart after an idle stop. It should be noted that the hydraulic control valve unit7is designed to secure a certain hydraulic pressure in each pulley cylinder chamber33,37without significant oil drainage when the engine11stops for a short time under the idle stop control and to drain hydraulic oil gradually from the oil passages to the pulley cylinder chambers33,37when the engine11stops for a long time.

The CVT control unit10thus enables the torque transmission controller10ato perform clutch engagement control to determine a target clutch engagement pressure Pc* based on the accelerator opening θ and control the hydraulic control valve unit7to adjust the forward clutch engagement pressure Pc to the target clutch engagement pressure Pc* for proper engagement of the forward clutch25at the restart of the engine11. The CVT control unit10also enables the torque transmission controller10ato perform so-called torque-down control, in tandem with the clutch engagement control, to reduce engine torque output Te according to the engagement state of the forward clutch25and prevent excessive torque input to the automatic transmission device.

It is herein assumed that the forward clutch25comes into engagement through the following phases: a precharge phase during which a clutch piston stroke is started to push clutch plates against the biasing force of a disc spring and eliminate play between the clutch plates; an engagement phase during which the clutch piston stroke is coming to an end so that the clutch plates are engaged to allow torque transmission with clutch disc slippage; an engagement termination phase during which the clutch plate slippage is not minimized or eliminated even after the expiration of a predetermined time period; and then, a full engagement phase during which the clutch discs are fully engaged to allow torque transmission with no clutch plate slippage.

In particular, the clutch engagement and torque-down controls of the torque transmission controller10aduring the forward clutch engagement phase is programmed in the following steps shown inFIG. 10Ain the first embodiment.

At step S100, the torque transmission controller10afirst determines a target clutch engagement time Δt* of the forward clutch25based on the accelerator opening θ.

At step S200, the torque transmission controller10anext determines a target clutch engagement torque Tc* of the forward clutch25, which corresponds to the target clutch engagement pressure Pc*, based on the accelerator opening θ and the target clutch engagement time Δt*.

At step S300, the torque transmission controller10athen determines a target engine torque Te* based on the target clutch engagement torque Tc*.

The torque transmission controller10athen generates control signals responsive to these control signals to control the operations of the hydraulic control valve unit7and the engine control unit18.

In this way, the torque transmission controller10ais configured to determine the target clutch engagement time Δt* and the target clutch engagement torque Tc* based on the accelerator opening θ and cause the hydraulic control valve unit7to control the engagement state of the forward clutch25according to the target clutch engagement time Δt* and the target clutch engagement torque Tc* in such a manner as to bring the forward clutch25into engagement during a lapse of the clutch engagement time Δt*. The target clutch engagement time Δt* and the target clutch engagement torque Tc* are determined based on the accelerator opening θ until the forward clutch25comes into full engagement. It is therefore possible according to the first embodiment to control a driving force of the vehicle, which generally depends on the torque capacity Tc of the forward clutch25, in response to changes in the accelerator opening θ and allow the vehicle to make a smooth start without causing driver discomfort.

The vehicle inertia becomes suddenly exerted upon full engagement of the forward clutch25. If the engine torque Te has not increased to compare favorably with the vehicle inertia at the time of full engagement of the forward clutch25, the vehicle cannot attain a sufficient driving force. The engine speed Ne becomes once decreased and then increased to a desired level under such an insufficient engine torque condition. This results in vehicle shaking. Further, the transient target clutch engagement torque Tc* is inevitably smaller than the final target clutch engagement torque Tcf* so that the target engine torque Te* is smaller during a transient period before the full engagement of the forward clutch25than after the full engagement of the forward clutch25. If the forward clutch25comes into full engagement during such a transient period, there arises vehicle shaking due to insufficient engine torque.

In the first embodiment, however, the forward clutch25can be prevented from coming into full engagement during a transient period before the expiration of the clutch engagement time Δt*. This makes it possible to avoid the occurrence of vehicle shaking due to insufficient engine torque and provide improved driving comfort.

The torque transmission controller10ais further configured to determine the target engine torque Te* based on the target clutch engagement torque Tc* such that the target engine torque Te* gradually increases with the target clutch engagement torque Tc*, and then, cause the engine control unit18to control the operations of the engine11according to the target engine torque Te*. This makes it possible to secure a sufficient engine torque Te for vehicle driving at the time of full engagement of the forward clutch25without lengthening the clutch engagement time.

More specifically, the torque transmission controller10aof the CVT control unit10is structured as follows.

As shown inFIG. 4, the torque transmission controller10aincludes a target clutch engagement time calculation block100, a target clutch torque calculation block200, a target engine torque calculation block300and a timer101.

The target clutch engagement time calculation block100calculates the target engagement time Δt* of the forward clutch25(i.e. the time from the initiation of the clutch precharge phase to the completion of the clutch engagement phase) based on the accelerator opening θ, and then, outputs the target clutch engagement time Δt* to the transient target clutch torque calculation circuit220. Herein, the target clutch engagement time calculation block100may also serve as the after-mentioned target clutch engagement time setting portion314aof the engine torque calculation block300.

The target clutch torque calculation block200has a final target clutch engagement torque calculation circuit210and a transient target clutch engagement torque calculation circuit220as shown inFIG. 4.

As shown inFIG. 5, the final target clutch engagement torque calculation circuit210has a target driving force setting section211and a final target clutch engagement torque calculation section212as shown inFIG. 5.

The target driving force setting section211stores therein a target driving force setting map to set a target driving force Fd(θ) based on the accelerator opening θ.

The final target clutch engagement torque calculation section212calculates the final target engagement torque Tcf* of the forward clutch25to be achieved at the time of completion of the clutch engagement phase by multiplying the target driving force Fd(θ) by a wheel radius Rtire and dividing the multiplication result by a primary gear ratio Ip and a final gear ratio If to, and then, outputs the final target clutch engagement torque Tcf* to the transient target clutch engagement torque calculation circuit220.

As shown inFIG. 6, the transient target clutch engagement torque calculation circuit220has a target precharge pressure setting section221, a pressure-to-torque conversion section222, a first adding section223, a multiplication section224and a second adding section225.

The target precharge pressure setting section221sets a target precharge pressure Pp* of the forward clutch25(i.e. a target value of the hydraulic pressure supplied to the forward clutch25during the clutch precharge phase to make a clutch piston stroke for clutch play elimination) based on the accelerator opening θ and outputs the target precharge pressure Pp* to the pressure-to-torque conversion section222.

The pressure-to-torque conversion section222converts the target clutch precharge pressure Pp* into a target clutch precharge torque Tcp* and outputs the target precharge torque Tcp* to the first and second adding sections223and225.

The first adding section223inverts the sign of the target precharge torque Tcp* and calculates a difference ΔT between the final target clutch torque Tcf* and the target precharge torque Tcp* (ΔT=Tcf*−Tcp*). Since the target clutch precharge torque Tcp* corresponds to a torque of the forward clutch25for clutch play elimination during the precharge phase, the torque difference ΔT corresponds to a substantial target engagement torque of the forward clutch25for torque transmission.

The multiplication section224determines a target clutch torque increment ΔTc* by multiplying the torque difference ΔT by the time Δt lapsed during the clutch engagement phase (measured by the timer101) and dividing the multiplication result by the target clutch engagement time Δt* and outputs the target incremental clutch torque ΔTc* to the second adding section225.

The second adding section225adds the target incremental clutch torque ΔTc* to the target clutch precharge torque Tcp* to determine the target engagement torque Tc* of the forward clutch25to be achieved during the transient period before the completion of the clutch engagement phase (i.e. during the clutch engagement and engagement termination phases where the clutch plates are gradually engaged), and then, outputs the transient target clutch engagement torque Tc* to the target engine torque calculation block300.

The calculation of the target clutch torque Tc* is done by the target clutch torque calculation block200through the following steps as shown inFIG. 11.

At step S211, the target driving force setting section211reads the target driving force Fd(θ) from the target driving force setting map with reference to the accelerator opening θ. The control proceeds to step S211.

At step S212, the final target clutch engagement torque calculation section212calculates the final target clutch engagement torque Tcf* based on the target driving force Fd(θ), the wheel radius Rtire, the primary gear ratio Ip and the final gear ratio If. The control proceeds to step S221.

At step S221, the target precharge pressure setting section221sets the target clutch precharge pressure Pp* based on the accelerator opening θ. The control proceeds to step S222.

At step S222, the pressure-to-torque conversion section222converts the target clutch precharge pressure Pp* into the target clutch precharge torque Tcp*. The control proceeds to step S223.

At step S223, the first adding section223calculates the torque difference ΔT between the final target clutch engagement torque Tcf* and the target clutch precharge torque Tcp* (ΔT=Tcf*−Tcp*).

At step S224, the multiplication section224calculates the target incremental clutch torque ΔTc* based on the torque difference ΔT, the lapsed time Δt and the target clutch engagement time Δt* (ΔTc*=ΔT×Δt2/Δt*).

At step S225, the second adding section225determines the transient target clutch engagement torque Tc* based on the target incremental clutch torque ΔTc* and the target clutch precharge torque Tcp* (Tc*=Tcp*+ΔTc*).

The target engine torque calculation block300determines the target engine torque Te* based on the target clutch torque Tc* in such a manner that the target engine torque Te* changes depending on the accelerator opening θ without exceeding the maximum transmissible torque of the forward clutch25, and has a target feedforward torque calculation circuit310, a target feedback torque calculation circuit320and an adding section330as shown inFIG. 7.

The target feedforward torque calculation circuit310calculates a target feedforward engine torque Te* (FF) from the target clutch torque Tc*, the target clutch engagement time Δt*, the engine speed Ne and the primary pulley rotation speed Npri and outputs the target feedforward engine torque Te* (FF) to the adding section330.

The target feedback torque calculation circuit320calculates a target feedback engine torque Te* (FB) from the target clutch engagement time Δt*, the primary pulley rotation speed Npri, the turbine speed Ntb and the turbine speed drop flag Ftb and outputs the target feedback engine torque Te*(FB) to the adding section330.

The adding section330determines a target engine torque Te* by superimposing (adding) the target feedforward engine torque Te* (FF) on the target feedback engine torque Te* (FB) and outputs the target engine torque Te* to the engine control unit18, and then, outputs the target engine torque Te* to the engine control unit18.

The calculation of the target engine torque Te* is consequently done by the target engine torque calculation block300through the following steps as shown inFIG. 12.

At step S310, the target feedfoward torque calculation circuit310calculates the target feedforward engine torque Te* (FF) based on the target clutch torque Tc*, the target clutch engagement time Δt*, the engine speed Ne and the primary pulley rotation speed Npri.

At step S320, the target feedback engine torque calculation circuit320calculates the target feedback engine torque Te* (FB) based on the target clutch engagement time Δt*, the primary pulley rotation speed Npri, the turbine speed Ntb and the turbine speed drop flag Ftb.

At step S330, the adding section330calculates the target engine torque Te* by superimposing the target feedforward engine torque Te* (FF) on the target feedback engine torque Te* (FB) and outputs the target engine torque Te*.

In the first embodiment, the target feedforward torque calculation circuit320has a speed ratio setting section311, a torque ratio setting section312, a multiplication section313, a torque corrector section314and an adding section315as shown inFIG. 8A.

The speed ratio setting sections311sets a speed ratio e(t) (=Npri/Ne) by dividing the primary pulley rotation speed Npri by the engine speed Ne and outputs the speed ratio e(t) to the torque ratio setting section312.

The torque ratio setting section312stores therein a torque ratio setting map to set a torque ratio tr of the torque converter1based on the speed ratio e(t) and outputs the torque ratio tr to the multiplier313.

The multiplication section313determines a transient target engine torque Tet* by multiplying the target clutch torque Tc* by the torque ratio tr and outputs the transient target engine torque Tet* to the adding section315.

The torque corrector section314calculates a torque correction amount ΔTe based on the accelerator opening θ and outputs the torque correction amount ΔTe to the adding section315. The torque correction amount ΔTe is set to make a correction to the target engine torque in view of the clutch engagement speed and thereby avoid the occurrence of shaking or shock in the vehicle upon abrupt engagement of the forward clutch25. The torque corrector section314includes a target clutch engagement time setting portion314aand a torque correction amount setting portion314bas shown inFIG. 8B. The target clutch engagement time setting portion314astores therein a target clutch engagement setting map to set the target clutch engagement time Δt* based on the accelerator opening θ, whereas the torque correction amount setting portion314stores therein a torque correction amount setting map to set the torque correction amount ΔTe based on the target clutch engagement time Δt* in view of the clutch engagement speed.

The adding section315adds the torque correction amount ΔTe to the transient target engine torque Tet* to determine the target feedforward engine torque Te* (FF), and then, outputs the target feedforward torque Te* (FF) to the adding section330.

It is conceivable to use as the speed ratio e(t) a turbine/engine speed ratio of Ntb/Ne in place of the pulley/engine speed ratio Npri/Ne for calculation of the target feedforward torque Te* (FF). However, the torque capacity Tc of the forward clutch25is small soon after the restart of the engine11. The rotation speed Ntb of the turbine1a, which is located on the input side of the forward clutch25, follows the engine speed Ne so that the turbine/engine speed ratio Ntb/Ne becomes nearly 1 soon after the engine restart. Accordingly the torque ratio tr becomes so small that the target engine torque Te* takes a large value when the turbine/engine speed ratio Ntb/Ne is used as the speed ratio e(t). There is no choice but to raise the engine speed Ne and adjust the engine torque Te to such a large target value Te. This results in a large vehicle starting time lag.

In the first embodiment, however, the pulley/engine speed ratio Npri/Ne is used as the speed ratio e(t). The rotation speed Npri of the primary pulley30a, which is located on the output side of the forward clutch25, is low soon after the restart of the engine11so that the speed ratio e(t) takes a small value to secure a large torque ratio tr and a smaller target engine torque Te*. This makes it possible to avoid a vehicle starting time lag due to an engine speed response delay soon after the engine restart.

The calculation of the target feedforward engine torque Te* (FF) is done by the target feedforward torque calculation circuit310through the following steps as shown inFIG. 13.

At step S311, the speed ratio setting sections311calculates the speed ratio e(t) from the engine speed Ne and the primary pulley rotation speed Npri. The control proceeds to step S312.

At step S312, the torque ratio setting section312reads the torque ratio tr from the torque ratio setting map with reference to the speed ratio e(t). The control proceeds to step S313.

At step S313, the multiplication section313calculates the transient target engine torque Tet* from the target clutch torque Tc* and the torque ratio tr. The control proceeds to step S314.

At step S314, the torque corrector section314determines the torque correction amount ΔTe based on the accelerator opening θ in view of the torque engagement speed. The control proceeds to step S315.

At step S315, the adding section315calculates the target feedforward engine torque Te* (FF) from the torque correction amount ΔTe and the transient target engine torque Tet*. The control proceeds to step S320.

As shown inFIG. 9, by contrast, the feedback torque calculation circuit320has an actual speed difference calculating section321, a target differential speed change rate calculating section322, a time differentiator section323, a subtractor section324, a feedback controller section325and a selector section326.

The speed difference calculating section321calculates an actual difference between the turbine speed Ntb and the primary pulley rotation speed Npri and outputs the speed difference (Ntb−Npri) to the target differential speed change rate calculating section322and the time differentiator section323.

The target differential speed change rate calculating section322stores therein a target speed change rate setting map to set a target speed change rate ΔN* based on the speed difference (Ntb−Npri) and outputs the target speed change rate ΔN* to the subtractor section324. In the target differential speed change rate setting map, the target speed change rate ΔN* is defined with respect to the actual speed difference (Ntb−Npri) such that the target speed change rate ΔN* increases with decrease in the actual speed difference (Ntb−Npri) and decreases with increase in the actual speed difference (Ntb−Npri). When the actual speed difference (Ntb−Npri) is so large as to result in high clutch heat generation, the target speed change rate ΔN* is set to a low valve in order to allow quick engagement of the forward clutch25and thereby limit total heat generation in the forward clutch25for prevention of clutch durability deterioration. On the other hand, the target speed change rate ΔN* is set to a high valve in order to allow gradual engagement of the forward clutch25and thereby avoid clutch engagement shock when the actual speed difference (Ntb−Npri) is small.

The time differentiator section323calculates from the actual speed difference (Ntb−Npri) an actual speed change rate ΔN corresponding to the target clutch engagement time Δt* and outputs the actual speed change rate ΔN to the subtractor section324.

The subtractor section324calculates a difference err(t) between the actual speed change rate ΔN and the target speed change rate ΔN* and outputs the speed change rate difference err(t) to the feedback controller section325.

The feedback controller section325calculates a target feedback engine torque Te* (FB) based on the accelerator opening θ in such a manner that the engine torque Te follows the accelerator opening θ and outputs the target feedback engine torque Te* (FB) to the selector section326.

The selector section326selects either the target feedback engine torque Te* (FB) or zero depending on the turbine speed drop flag Ftb and outputs the selected torque value of Te* (FB) or 0 to the adding section330.

Herein, the turbine speed drop flag Ftb is set to 1 when the turbine speed Ntb is higher than or equal to a given threshold value Nα and the turbine speed differentiation result ΔNtb is negative and set to 0 so as to allow zero output at other times. When the turbine speed Ntb is higher than or equal to the threshold value Nα and the turbine speed differentiation result ΔNtb is negative, the turbine speed Ntb decreases with increase in the engine speed Ne. It means that clutch plate slippage occurs in the forward clutch25. If it takes a longer time to bring the forward clutch25into full engagement in such a case, the amount of heat generated in the forward clutch25due to clutch plate slippage becomes increased to cause a deterioration of the durability of the forward clutch25. If the forward clutch25is engaged abruptly, by contrast, there arises vehicle shaking to cause driver's discomfort due to clutch engagement shock. In the first embodiment, the target feedback engine torque Te* (FB) is outputted so as to perform feedback control on the engine torque Te according to the accelerator opening θ when the turbine speed drop flag Ftb is set to 1. The turbine speed Ntb can be thus decreased promptly by determining the target engine torque Te* and the target clutch torque Tc* based on the same control parameter Δt* in the case where there is a possibility that it takes a long time to engage the forward clutch25. This makes it possible to prevent the durability of the forward clutch25from becoming deteriorated due to slow engagement of the forward clutch25and to avoid the occurrence of clutch engagement shock due to abrupt engagement of the forward clutch25. The zero output is produced when the turbine speed drop flag Ftb is set to 0.

The calculation of the target feedback engine torque Te* (FB) is consequently done by the target feedback torque calculation circuit320through the following steps as shown inFIG. 14.

At step S321, the speed difference calculating section321calculates the difference between the turbine speed Ntb and the primary pulley rotation speed Npri. The control proceeds to step S322.

At step S322, the target differential speed change rate calculating section322calculates the target speed change rate ΔN* based on the speed difference (Ntb−Npri) with reference to the target speed change rate setting map. The control proceeds to step S323.

At step S323, the time differentiator section323calculates the actual speed change rate ΔN from the speed difference (Ntb−Npri). The control proceeds to step S324.

At step S324, the subtractor section324calculates the difference err(t) between the actual speed change rate ΔN and the target speed change rate ΔN*. The control proceeds to step S325.

At step S325, the feedback controller section325calculates the target feedback engine torque Te* (FB) based on the accelerator opening θ. The control proceeds to step S326.

At step S326, the selector section326determines whether the turbine speed drop flag Ftb is set to 1. If Ftb=1 (Yes at step S326), the control proceeds to step S327. If Ftb=0 (No at step S326), the control proceeds to step S328.

At step S327, the selector section326outputs the target feedback engine torque Te* (FB). The control proceeds to step S330.

At step S327, the selector section326produces zero output. The control proceeds to step S330.

The torque transmission controller10ais configured to set the target engine torque Tep* to a level corresponding to the target clutch precharge torque Tcp* so as to limit the engine output Te to the target clutch precharge torque level during the forward clutch precharge phase, and then, to keep the target engine torque Te* constant so as to cancel the clutch plate slip gradually and promptly during the forward clutch engagement termination phase. The engine torque Te may alternatively be decreased during the forward clutch engagement termination phase. The torque transmission controller10ais further configured to, at the time of controlling the hydraulic control valve unit7to switch its hydraulic pressure control line upon full engagement of the forward clutch25, decrease the amount of torque-down control on the engine torque Te gradually until the engine torque Te reaches a given degree in order to avoid the occurrence of belt slippage in the CVT3due to abrupt torque input.

The engine torque control of the first embodiment is effected as follows as shown inFIG. 15.

At time t1, the supply of hydraulic oil to the forward clutch25is allowed in preparation for engine restart upon a driver's brake pedal releasing operation. The forward clutch25shifts into the precharge phase at this time.

Between time t1and time t2, the secondary pulley clamping pressure Psec starts increasing by accelerator pedal actuation. The forward clutch25remains in the precharge phase during the time period between t1and t2. The actual hydraulic pressure Pc of the forward clutch25is controlled to the target start clutch pressure Pp* (i.e. the torque Tc of the forward clutch25is controlled to the target precharge torque Tp*) according to the accelerator opening θ. Further, the forward clutch25is hardly capable of torque transmission in the precharge phase. The engine torque Te is thus limited to the target forward clutch precharge torque level.

At time t2, the actual hydraulic pressure Pc of the forward clutch25starts increasing toward the target clutch pressure Pc*.

Between time2and time3, the actual hydraulic pressure Pc of the forward clutch25still keeps on increasing toward the target clutch pressure Pc*. The engine torque Te does not increase during this time period between t2and t3. The secondary pulley clamping pressure Psec reaches its target level and is held constant at the target level after that time.

At time t3, the forward clutch25completes its piston stroke for clutch play elimination and shifts into the engagement phase. The torque transmission controller10astarts its engine torque controls to calculate the target clutch engagement time Δt* based on the accelerator opening θ, calculate the target clutch torque Tc* based on the target clutch engagement time Δt* and then calculate the target engine torque Te* based on the target clutch engagement time Δt* and the target clutch torque Tc*.

Between time t3and time t4, the actual hydraulic pressure Pc of the forward torque25gradually increases so as to control the engagement torque Tc of the forward clutch25to the target clutch torque Tc*. Further, the clutch plate slippage condition of Ntb≧Nα and ΔNtb<0 is not yet satisfied so that the turbine speed drop flag Ftb is set to 0 during this time period between t3and t4. Thus, the target engine torque Te* becomes equal to the target feedforward engine torque Te* (FF) since the output of the target feedback engine torque Te* (FB) is not allowed. The engine torque Te starts increasing to the target engine torque Te* so as not to exceed the maximum transmissible torque of the forward clutch25.

At time t4, the actual hydraulic pressure Pc of the forward torque25keeps on increasing so as to control the engagement torque Tc of the forward clutch25to the target clutch torque Tc*. The turbine speed drop flag Ftb is set to 1 to allow the output of the target feedback engine torque Te* (FB) upon satisfaction of the clutch plate slippage condition of Ntb≧Nα and ΔNtb<0 at this time. The target engine torque Te* becomes thus equal to the sum of the target feedforward engine torque Te* (FF) and the target feedback engine torque Te* (FB). The engine torque Te keeps on increasing to the target engine torque Te*.

At time t5, the forward clutch25shifts into the engagement termination phase after a lapse of the target clutch engagement time Δt* from the initiation time t3of the clutch engagement phase. The clutch plate slippage has not yet been eliminated at this time.

Between time t5and time t6, the engine torque Te is kept constant so as to cancel the clutch plate slip gradually and promptly. The engine torque Te may alternatively be decreased during the period of the clutch engagement termination phase between time t5and time t6.

At time t6, the forward clutch25comes into full engagement upon elimination of the clutch plate slippage.

At time t7, the hydraulic pressure control line of the hydraulic control valve unit7is changed to directly supply the pressure regulated by the clutch regulator valve45to the forward clutch25. At this time, the amount of torque-down control on the engine torque Te starts gradually decreasing so that the engine torque Te increases. When the engine torque Te reaches a given degree, the engine torque control is cancelled. Then, the engine torque Te reaches its maximum degree.

Further, the engine speed Ne changes over time as follows as shown inFIG. 16under the engine torque control of the first embodiment.

At time t11, the clutch engagement pressure Pc starts increasing toward the target precharge pressure Pp* upon a driver's accelerator pedal operation.

At time t12, the engine speed Ne starts increasing in response to changes in the accelerator opening θ although complete combustion has not yet occurred in the engine11.

At time t13, the turbine speed Ntb starts increasing.

At time t14, complete combustion occurs in the engine11so that the engine speed Ne becomes sharply increased. The turbine speed Ntb follows the engine speed Ne to become also increased sharply.

Between time t14and time t15, the turbine speed Ntb reaches and exceeds the given threshold value Nα.

At time t15, the forward clutch25shifts into the engagement phase upon completion of the piston stroke for clutch play elimination. Both of the engine speed Ne and the turbine speed Ntb keep on increasing.

At time t16, the time differential value ΔNe of the engine speed Ne starts decreasing so that the time differential value ΔNtb of the turbine speed Ntb becomes negative. Namely, the turbine speed drop flag Ftb is set to 1 upon satisfaction of the clutch plate slippage condition of Ntb≧Nα and ΔNtb<0. The feedback control of the engine torque Te is performed based on the target feedback engine torque Te* (FB).

At time t18, the turbine speed Ntb starts increasing once again so that the time differential value ΔNtb of the turbine speed Ntb becomes positive. Further, the primary pulley rotation speed Npri gets agreed with the turbine speed Ntb and then increases with the turbine speed Ntb. The forward clutch25comes into full engagement.

As described above, it is possible according to the first embodiment to control the output of the engine11and the engagement of the forward clutch25in such a manner as to allow the vehicle to make a smooth start upon the restart of the engine11after the idle stop.

A drive control system for an automotive vehicle according to the second embodiment of the present invention will be next explained below. The vehicle drive control system of the second embodiment is structurally similar to that of the first embodiment, except that the target feedback torque calculation circuit320is configured to make, based on the speed difference (Ntb−Npri), a correction to the target speed change rate ΔN*.

The target feedback torque calculation circuit320of the second embodiment has a correction factor setting section322a, a division section322band a multiplication section322cas shown inFIG. 17in place of the target differential speed change rate calculating section322of the first embodiment.

The correction factor setting section322astores therein a correction factor setting map to set a correction factor K based on the speed difference (Ntb−Npri) and outputs the correction factor to the multiplication section322c. In the correction factor setting map, the correction factor K is defined in such a manner that it decreases with the speed difference (Ntb−Npri).

The division section322bdivides the speed difference (Ntb−Npri) by the target clutch engagement time Δt* to determine the target speed change rate ΔN* and outputs the multiplication section322c.

The multiplication section322multiplies the target speed change rate ΔN* by the correction factor K and outputs the corrected target speed change rate ·KΔN* to the subtractor section324.

By such correction, the target speed change rate ΔN* can be corrected to the more accurate value K·ΔN* so that the speed change rate difference err(t) becomes smaller before the forward clutch25comes into full engagement. This makes it possible to achieve smooth engagement of the forward clutch25.

The entire contents of Japanese Patent Application No. 2005-036410 (filed on Feb. 14, 2005) are herein incorporated by reference.

Although the present invention has been described with reference to the specific embodiments of the invention, the invention is not limited to the above-described embodiments. Various modification and variation of the embodiments described above will occur to those skilled in the art in light of the above teaching. The scope of the invention is defined with reference to the following claims.