JP2010210076A - Control device for vehicular continuously variable transmission - Google Patents

Abstract

A control device for a continuously variable transmission for a vehicle capable of suitably preventing hunting that occurs during downshifting in a control device for a continuously variable transmission controlled by a solenoid valve for speed change.
When a continuously variable transmission is switched from upshift to downshift, a thrust ratio is configured such that a differential pressure calculation means 168 can communicate with a drive side hydraulic actuator 42c via a speed ratio control valve DN116. A differential pressure Pdiff between the thrust ratio control hydraulic pressure Pτ output from the output port 118t of the control valve 118 and the transmission pressure Pin of the drive side hydraulic actuator 42c is calculated, and the down switching control means 164 is based on the differential pressure Pdiff. Since the solenoid valve DS1 and the solenoid valve DS2 are controlled, it is possible to suitably prevent hunting that occurs during the speed change of the continuously variable transmission 18.
[Selection] Figure 7

Description

  The present invention relates to a control device for a continuously variable transmission for a vehicle, and more particularly to suppression of hunting that occurs during downshifting.

  A drive-side hydraulic actuator for changing the groove width of a drive-side pulley and its drive in a transmission control device for a continuously variable transmission for a vehicle having a drive-side pulley, a driven-side pulley, and a belt wound around both pulleys And a shift control valve that changes the groove width of the drive pulley by controlling the supply and discharge of hydraulic fluid to the side hydraulic actuator, and shifts based on the deviation between the target value and the actual value of the rotation speed of a predetermined rotating member. It is well known that gears of a continuously variable transmission are controlled by controlling the supply and discharge of hydraulic fluid to and from a drive side hydraulic actuator by a control valve. In a vehicle state where the rotational speed of the rotating member is less than a predetermined vehicle speed that is difficult to detect, the gear ratio of the continuously variable transmission is set to a predetermined gear ratio by confining the hydraulic oil in the drive side hydraulic actuator by the shift control valve. It is also well known.

  For example, when the vehicle travels above a predetermined vehicle speed, the drive oil is supplied to the drive-side hydraulic actuator by the shift control valve, thereby narrowing the groove width of the drive-side pulley and upshifting the continuously variable transmission, By discharging the hydraulic oil from the drive side hydraulic actuator, the groove width of the drive side pulley is widened and the continuously variable transmission is downshifted. Further, when the vehicle is stopped, so-called closing control is performed in which the hydraulic oil is confined in the drive side hydraulic actuator by the speed change control valve, so that the speed ratio of the continuously variable transmission becomes the lowest speed side speed ratio. It is said.

  By the way, for example, when the continuously variable transmission for a vehicle is upshifted, the actual rotational speed becomes the target rotational speed while the actual rotational speed (actual rotational speed) of the predetermined rotating member follows the target rotational speed. When the downshift is performed after reaching, there is a problem that the downshift is greatly performed and hunting that repeats the upshift again occurs. For example, for an upshift, the shift control valve is controlled via a command hydraulic pressure from the upshift solenoid valve to supply hydraulic oil to the drive side hydraulic actuator, and for a downshift, the command hydraulic pressure from the downshift solenoid valve is supplied. By controlling the shift control valve via the control valve, the hydraulic fluid of the drive side hydraulic actuator is discharged. At this time, when switching from upshift to downshift, the oil passage of the shift control valve is switched accordingly, but a predetermined response delay occurs with respect to the shift output by the solenoid valve. Due to this response delay, the hydraulic oil may be discharged more than necessary and the downshift may increase.

  FIG. 8 is a diagram showing hunting generated by switching from the above-described upshift to downshift. 8 shows the DUTY command values of the upshift solenoid valve and the downshift solenoid valve, the solid line shows the upshift DUTY output from the upshift solenoid valve, and the broken line shows the downshift solenoid. Each of the downshifts DUTY output from the valve is shown. Further, the lower side of FIG. 8 shows the pulley position of the driving pulley corresponding to the gear ratio, and when the pulley position moves upward in FIG. 8, the gear is shifted to the low speed gear ratio side (downshift side) and moved downward. Then, the gear is shifted to the high speed gear ratio side (upshift side). Here, the solid line represents the actual pulley position, and the broken line represents the target pulley position. As shown in FIG. 8, when the upshift DUTY is output, the actual pulley position is upshifted toward the target pulley position. When the actual pulley position passes the target pulley position, a downshift DUTY is output, and the pulley position moves to the downshift side. At this time, the upshift DUTY is output due to a response delay of the shift control valve. Nevertheless, it is downshifted greatly. Then, the upshift is again performed by the upshift DUTY. Similarly, since the downshift is greatly reduced when the downshift DUTY is output, hunting in which the upshift DUTY and the downshift DUTY are repeated occurs.

JP 2006-46517 A JP 2005-20769 A JP 2003-227564 A

  On the other hand, in the continuously variable transmission control device of Patent Document 1, when the number of huntings of the control signal in a predetermined period exceeds a predetermined number when hunting occurs, the responsiveness of the gear ratio control is suppressed. A technique for preventing hunting is disclosed. However, the control device for a continuously variable transmission disclosed in Patent Document 1 has a problem in that the speed change response also decreases as the speed ratio control response decreases.

  The present invention has been made against the background of the above circumstances, and its object is to suitably prevent hunting that occurs during downshift in a control device for a continuously variable transmission controlled by a solenoid valve for shifting. Another object of the present invention is to provide a control device for a continuously variable transmission for a vehicle.

  In order to achieve the above object, the gist of the invention according to claim 1 is that: (a) a continuously variable transmission for a vehicle having a driving pulley, a driven pulley, and a belt wound around both pulleys. A drive-side hydraulic actuator for changing the groove width of the drive-side pulley, and a gear change for changing the gear ratio of the continuously variable transmission for the vehicle by controlling supply / discharge of hydraulic oil to / from the drive-side hydraulic actuator A control valve, an upshift solenoid valve for controlling the shift control valve to the upshift side, a downshift solenoid valve for controlling the shift control valve to the downshift side, and the vehicle continuously variable transmission In a vehicular continuously variable transmission control device, comprising: a shift control means for controlling the gear ratio to be a preset target gear ratio; (b) the continuously variable transmission is upshifted; Between the predetermined hydraulic pressure output from the output port of the control valve configured to be able to communicate with the driving hydraulic actuator via the shift control valve and the hydraulic pressure of the driving hydraulic actuator. A differential pressure calculating means for calculating a differential pressure; and (c) a down-switching control means for controlling the upshift solenoid valve and the downshift solenoid valve based on the differential pressure.

  According to a second aspect of the present invention, in the control device for a continuously variable transmission for a vehicle according to the first aspect, a deviation offset of the gear ratio with respect to the target gear ratio is set based on the differential pressure. The down-switching control means changes a downshift output line serving as a boundary line from which a downshift command is output from the downshift solenoid valve according to the deviation offset.

  According to a third aspect of the present invention, in the control device for a continuously variable transmission for a vehicle according to the second aspect, the down-switching time control means calculates a deviation between the actual gear ratio and the target gear ratio. When the deviation is smaller than the deviation offset, a preset upshift command is output by the upshift solenoid valve, and when the deviation is larger than the deviation offset, the downshift is performed according to the deviation. A downshift command is output from the solenoid valve for operation.

  According to a fourth aspect of the present invention for achieving the above object, (a) a continuously variable transmission for a vehicle having a driving pulley and a driven pulley, and a belt wound around both pulleys. , A drive-side hydraulic actuator for changing the groove width of the drive-side pulley, and a shift control valve for changing the pulley position of the drive-side pulley by controlling the supply and discharge of hydraulic oil to and from the drive-side hydraulic actuator An upshift solenoid valve for controlling the shift control valve to the upshift side, a downshift solenoid valve for controlling the shift control valve to the downshift side, and an actual pulley position of the continuously variable transmission for the vehicle A control device for a continuously variable transmission for a vehicle, comprising: (b) the continuously variable transmission. When switching from an upshift to a downshift, a predetermined hydraulic pressure output from an output port of a control valve configured to be able to communicate with the driving hydraulic actuator via the shift control valve and a hydraulic pressure of the driving hydraulic actuator A differential pressure calculating means for calculating a differential pressure between the downshift control means and (c) a down-switching control means for controlling the upshift solenoid valve and the downshift solenoid valve based on the differential pressure. To do.

  According to a fifth aspect of the present invention, in the control device for a continuously variable transmission for a vehicle according to the fourth aspect, a deviation offset of the pulley position with respect to the target pulley position is set based on the differential pressure. The down-switching control means changes a downshift output line serving as a boundary line from which a downshift command is output from the downshift solenoid valve according to the deviation offset.

  According to a sixth aspect of the present invention, in the control device for a continuously variable transmission for a vehicle according to the fifth aspect, the down-switching time control means calculates a deviation between the actual pulley position and the target pulley position. When the deviation is smaller than the deviation offset, a preset upshift command is output by the upshift solenoid valve, and when the deviation is larger than the deviation offset, the downshift is performed according to the deviation. A downshift command is output from the solenoid valve for operation.

  The gist of the invention according to claim 7 is the control device for a continuously variable transmission for a vehicle according to claim 3 or 6, wherein the down-switching control means is configured such that the deviation is smaller than the deviation offset. The output of the upshift solenoid valve is lowered to allow communication between the drive hydraulic actuator and the output port of the control valve.

  According to the control device for a continuously variable transmission for a vehicle according to the first aspect of the present invention, when the continuously variable transmission is switched from upshift to downshift, the differential pressure calculating means transmits the shift to the drive hydraulic actuator. The pressure difference between the predetermined hydraulic pressure output from the output port of the control valve configured to be able to communicate via the control valve and the hydraulic pressure of the drive side hydraulic actuator is calculated, and the down-switching time control means Since the up-shift solenoid valve and the down-shift solenoid valve are controlled based on the above, hunting that occurs during gear shifting of the continuously variable transmission can be suitably prevented. For example, even if the gear ratio of the continuously variable transmission is changed to the gear ratio on the downshift side with respect to the target gear ratio during the upshift, the output of the upshift command is controlled without outputting the downshift command. By doing so, the drive side hydraulic actuator and the output port of the control valve are communicated. As a result, the hydraulic oil of the drive side hydraulic actuator is discharged to the control valve side, the groove width of the drive side pulley is expanded, and as a result, the downshift is performed. Therefore, even if a downshift command is not output, a downshift is possible within a certain range. As mentioned above, since the output of a downshift command can be suppressed, hunting can be suppressed.

  Further, the optimum downshift is performed by determining whether to continue the upshift command or to output the downshift command based on the differential pressure. For example, when the differential pressure is large, when the drive side hydraulic actuator and the control valve are communicated with each other, the flow rate of hydraulic oil discharged from the drive side hydraulic actuator to the control valve side increases. Thus, a large downshift can be obtained by appropriately controlling the upshift command and communicating the above without outputting the downshift command. Therefore, hunting is prevented as the downshift command is suppressed by performing a suitable downshift in consideration of the downshift amount based on the differential pressure and the current shift state.

  According to the control device for a continuously variable transmission for a vehicle of the invention according to claim 2, a deviation offset of the speed ratio with respect to the target speed ratio is set based on the differential pressure, and the down-switching control is performed. The means changes a downshift output line serving as a boundary line from which the downshift command is output from the downshift solenoid valve in accordance with the deviation offset, so that a suitable downshift is performed based on the changed downshift output line. To be implemented.

  Further, according to the control device for a continuously variable transmission for a vehicle of the invention according to claim 3, the down-switching time control means calculates a deviation between an actual gear ratio and a target gear ratio, and the deviation is the deviation. When the deviation is smaller than the offset, a preset upshift command by the upshift solenoid valve is output, and when the deviation is larger than the deviation offset, a downshift command is issued from the dow shift solenoid valve according to the deviation. For output, a suitable downshift is performed by comparing the deviation with the deviation offset.

  According to the control device for a continuously variable transmission for a vehicle according to a fourth aspect of the present invention, when the continuously variable transmission is switched from upshift to downshift, the differential pressure calculating means is connected to the drive side hydraulic actuator. A down-switching time control means calculates a differential pressure between a predetermined hydraulic pressure output from an output port of a control valve configured to communicate with the shift control valve and a hydraulic pressure of the drive side hydraulic actuator. Since the upshift solenoid valve and the downshift solenoid valve are controlled based on the differential pressure, it is possible to suitably prevent hunting that occurs during shifting of the continuously variable transmission. For example, even if the pulley position of the continuously variable transmission is changed to the pulley position on the downshift side with respect to the target pulley position during the upshift, the downshift command is not output and the output of the upshift command is preferable. By controlling this, the drive side hydraulic actuator and the output port of the control valve are communicated. As a result, the hydraulic oil of the drive side hydraulic actuator is discharged to the control valve side, the groove width of the drive side pulley is expanded, and as a result, the downshift is performed. Therefore, even if a downshift command is not output, a downshift is possible within a certain range. As mentioned above, since the output of a downshift command can be suppressed, hunting can be suppressed.

  Further, the optimum downshift is performed by determining whether to continue the upshift command or to output the downshift command based on the differential pressure. For example, when the differential pressure is large, when the drive side hydraulic actuator and the control valve are communicated with each other, the flow rate of hydraulic oil discharged from the drive side hydraulic actuator to the control valve side increases. Thus, a large downshift can be obtained by appropriately controlling the upshift command and communicating the above without outputting the downshift command. Therefore, hunting is prevented as the downshift command is suppressed by performing a suitable downshift in consideration of the downshift amount based on the differential pressure and the current shift state.

  According to the control device for a continuously variable transmission for a vehicle of the invention according to claim 5, a deviation offset of the pulley position with respect to the target pulley position is set based on the differential pressure, and the down-switching control is performed. The means changes a downshift output line serving as a boundary line from which the downshift command is output from the downshift solenoid valve in accordance with the deviation offset, so that a suitable downshift is performed based on the changed downshift output line. To be implemented.

  According to the control device for a continuously variable transmission for a vehicle according to claim 6, the down-switching control means calculates a deviation between the actual pulley position and the target pulley position, and the deviation is the deviation. When the deviation is smaller than the offset, a preset upshift command by the upshift solenoid valve is output, and when the deviation is larger than the deviation offset, a downshift command is issued from the dow shift solenoid valve according to the deviation. For output, a suitable downshift is performed by comparing the deviation with the deviation offset.

  According to the control device for a continuously variable transmission for a vehicle of the invention according to claim 7, the down-switching control means outputs the output of the upshift solenoid valve when the deviation is smaller than the deviation offset. In order to make the drive side hydraulic actuator and the output port of the control valve communicate with each other, the hydraulic oil of the drive side hydraulic actuator is discharged to the control valve side and the groove width of the drive side pulley is expanded. Therefore, by reducing the output of the upshift solenoid valve, it is possible to implement a downshift without outputting a downshift command.

1 is a skeleton diagram illustrating a vehicle drive device to which the present invention is applied. It is a block diagram explaining the principal part of the control system provided in the vehicle in order to control the vehicle drive device etc. of FIG. FIG. 3 is a hydraulic circuit diagram showing a main part relating to belt clamping pressure control of a continuously variable transmission, gear ratio control, and engagement hydraulic control of a forward clutch or reverse brake accompanying operation of a shift lever in the hydraulic control circuit. It is a figure which shows an example of the shift map used when calculating | requiring a target input rotational speed in the shift control of a continuously variable transmission. It is a figure which shows an example of the required hydraulic pressure map which calculates | requires required hydraulic pressure according to gear ratio etc. in the clamping pressure control of a continuously variable transmission. This is a relationship obtained and stored in advance between the gear ratio and the thrust ratio with the vehicle speed as a parameter. It is a functional block diagram explaining the principal part of the control function of the electronic control apparatus of FIG. It is a figure which shows the hunting which generate | occur | produces by switching from an upshift to a downshift. It is a figure which shows the relationship between differential pressure | voltage and deviation offset. 10 is a time chart showing a shift state when a deviation offset is set when the actual pulley position of the driving pulley is upshifted toward the shift target pulley position. FIG. 10 is another time chart showing a shift state when a deviation offset is set when the actual pulley position of the driving pulley is upshifted toward the shift target pulley position. It is a flowchart for demonstrating the control action which can prevent the main part of the control action of an electronic controller, ie, the hunting which generate | occur | produces at the time of switching from an upshift to a downshift.

  Here, preferably, the hydraulic pressure in the drive-side actuator is calculated based on the hydraulic pressure in the driven-side hydraulic actuator. According to this configuration, the drive-side hydraulic actuator is directly provided with the hydraulic sensor based on the preset relationship (relation map, relational expression) based on the hydraulic pressure in the driven-side hydraulic actuator, the gear ratio, the accelerator opening, and the like. The hydraulic pressure of the drive side hydraulic actuator can be calculated without any problem. Note that the hydraulic pressure in the driven hydraulic actuator is detected by a hydraulic pressure sensor.

  Preferably, the hydraulic pressure of the control valve is calculated by a relationship (relation map, relational expression) based on a preset hydraulic pressure of the driven hydraulic actuator. In this way, the hydraulic pressure of the control valve can be calculated by detecting the hydraulic pressure of the driven hydraulic actuator.

  Preferably, the differential pressure and the deviation offset are in a proportional relationship, and the deviation offset increases as the differential pressure increases. For example, when the differential pressure is large, the differential pressure between the hydraulic pressure of the drive side hydraulic actuator and the hydraulic pressure of the output port of the control valve is large, so when the drive side hydraulic actuator communicates with the control valve, it is discharged to the control valve side. Since the flow rate of hydraulic oil increases, the amount of downshift increases. In such a case, the deviation offset increases in proportion to the differential pressure, making it difficult for the downshift command to be output, but the downshift amount is large due to the reduced output of the upshift command, so the downshift command is output. Even if this is not done, a sufficient downshift is possible.

  On the other hand, the deviation offset decreases as the differential pressure decreases. When the differential pressure is small, the differential pressure between the hydraulic pressure of the drive side hydraulic actuator and the hydraulic pressure of the output port of the control valve is small, so when the drive side hydraulic actuator and the control valve communicate with each other, the operation is discharged to the control valve side. The oil flow rate decreases and the downshift amount decreases. In such a case, the deviation offset becomes smaller in proportion to the differential pressure, and the downshift command is likely to be output, but the downshift amount that is insufficient with the downshift by reducing the output of the upshift command is reduced. This can be achieved by outputting a shift command.

  Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings. In the following embodiments, the drawings are appropriately simplified or modified, and the dimensional ratios, shapes, and the like of the respective parts are not necessarily drawn accurately.

  FIG. 1 is a skeleton diagram illustrating the configuration of a vehicle drive device 10 to which the present invention is applied. This vehicle drive device 10 is a horizontal automatic transmission, which is suitably employed in an FF (front engine / front drive) type vehicle, and includes an engine 12 as a driving power source. The output of the engine 12 composed of an internal combustion engine is the crankshaft of the engine 12, the torque converter 14 as a fluid transmission device, the forward / reverse switching device 16, the belt type continuously variable transmission (CVT) 18, the reduction gear. It is transmitted to the differential gear device 22 via the device 20 and distributed to the left and right drive wheels 24L, 24R.

  The torque converter 14 includes a pump impeller 14p connected to the crankshaft of the engine 12 and a turbine impeller 14t connected to the forward / reverse switching device 16 via a turbine shaft 34 corresponding to an output side member of the torque converter 14. And power transmission is performed via a fluid. A lock-up clutch 26 is provided between the pump impeller 14p and the turbine impeller 14t, and a lock-up control valve (L not shown) in the hydraulic control circuit 100 (see FIGS. 2 and 3) is provided. The hydraulic pressure supply to the engagement side oil chamber and the release side oil chamber is switched by the (C / C control valve) or the like, thereby being engaged or released. The turbine impeller 14t is rotated integrally. The pump impeller 14p has a hydraulic pressure for controlling the transmission of the continuously variable transmission 18, generating a belt clamping pressure, controlling the engagement release of the lockup clutch 26, or supplying lubricating oil to each part. Is coupled to a mechanical oil pump 28 that is generated by being driven to rotate by the engine 12.

  The forward / reverse switching device 16 is composed mainly of a forward clutch C1, a reverse brake B1, and a double pinion type planetary gear device 16p, and the turbine shaft 34 of the torque converter 14 is integrally connected to the sun gear 16s. The input shaft 36 of the continuously variable transmission 18 is integrally connected to the carrier 16c, while the carrier 16c and the sun gear 16s are selectively connected via the forward clutch C1, and the ring gear 16r is connected to the reverse brake B1. And is fixed to the housing selectively. The forward clutch C1 and the reverse brake B1 correspond to an intermittent device, both of which are hydraulic friction engagement devices that are frictionally engaged by a hydraulic actuator.

  When the forward clutch C1 is engaged and the reverse brake B1 is released, the forward / reverse switching device 16 is brought into an integral rotation state, whereby the turbine shaft 34 is directly connected to the input shaft 36, and the forward power The transmission path is established (achieved), and the driving force in the forward direction is transmitted to the continuously variable transmission 18 side. When the reverse brake B1 is engaged and the forward clutch C1 is released, the forward / reverse switching device 16 establishes (achieves) the reverse power transmission path, and the input shaft 36 is connected to the turbine shaft 34. On the other hand, it is rotated in the opposite direction, and the driving force in the reverse direction is transmitted to the continuously variable transmission 18 side. When both the forward clutch C1 and the reverse brake B1 are released, the forward / reverse switching device 16 enters a neutral state (power transmission cut-off state) in which power transmission is cut off.

  The continuously variable transmission 18 is an input side member provided on the input shaft 36, a drive side pulley (primary pulley, primary sheave) 42 having a variable effective diameter, and an output side member provided on the output shaft 44. A driven pulley (secondary pulley, secondary sheave) 46 having a variable diameter and a transmission belt 48 (corresponding to the belt of the present invention) wound around these variable pulleys 42, 46 are provided. , 46 and the transmission belt 48, power is transmitted through a frictional force.

  The variable pulleys 42 and 46 are fixed rotation bodies 42 a and 46 a fixed to the input shaft 36 and the output shaft 44, respectively, and are not rotatable relative to the input shaft 36 and the output shaft 44 and are movable in the axial direction. Driven hydraulic actuator (primary pulley side hydraulic actuator) 42c and driven side hydraulic actuator (secondary pulley side) as hydraulic actuators that apply thrust to change the V-groove width between the movable rotors 42b and 46b provided Hydraulic actuator) 46c, and the hydraulic fluid supply / discharge flow rate to the drive side hydraulic actuator 42c is controlled by the hydraulic control circuit 100, so that the V-groove widths of both variable pulleys 42, 46 change. Thus, the engagement diameter (effective diameter) of the transmission belt 48 is changed, and the gear ratio γ ( The input shaft rotation speed Nin / output shaft rotation speed Nout) is continuously changed. Further, the secondary pressure (hereinafter referred to as belt clamping pressure) Pd, which is the hydraulic pressure of the driven hydraulic actuator 46c, is regulated by the hydraulic control circuit 100, so that the belt clamping pressure is reduced so that the transmission belt 48 does not slip. Be controlled. As a result of such control, a primary pressure (hereinafter referred to as shift pressure) Pin (corresponding to the hydraulic pressure of the drive side hydraulic actuator of the present invention), which is the hydraulic pressure of the drive side hydraulic actuator 42c, is generated.

  FIG. 2 is a block diagram for explaining a main part of a control system provided in the vehicle for controlling the vehicle drive device 10 of FIG. The electronic control unit 50 includes, for example, a so-called microcomputer having a CPU, a RAM, a ROM, an input / output interface, and the like. The CPU uses a temporary storage function of the RAM and follows a program stored in the ROM in advance. By performing signal processing, output control of the engine 12, shift control of the continuously variable transmission 18, belt clamping pressure control, torque capacity control of the lockup clutch 26, and the like are executed. This is divided into control and hydraulic control for the continuously variable transmission 18 and the lockup clutch 26.

The electronic control unit 50, representing the crankshaft rotation speed corresponding to the engine rotational speed crankshaft detected by the sensor 52 rotation angle (position) A _{CR (°)} and the rotational speed of the engine 12 (engine rotational speed) N _E Signal, a signal representing the rotational speed (turbine rotational speed) _NT of the turbine shaft 34 detected by the turbine rotational speed sensor 54, and an input that is the input rotational speed of the continuously variable transmission 18 detected by the input shaft rotational speed sensor 56. A signal representing the rotational speed of the shaft 36 (input shaft rotational speed) Nin, the rotational speed of the output shaft 44 (output shaft) which is the output rotational speed of the continuously variable transmission 18 detected by the vehicle speed sensor (output shaft rotational speed sensor) 58. Rotational speed) Nout, i.e., a vehicle speed signal representing the vehicle speed V corresponding to the output shaft rotational speed Nout, and the intake of the engine 12 detected by the throttle sensor 60 Pipe 32 throttle valve opening signal representing the throttle valve opening theta _TH of the electronic throttle valve 30 provided in (see FIG. 1), a signal representing the cooling water temperature T _W of the engine 12 detected by a coolant temperature sensor 62, CVT A signal that represents the oil temperature _TCVT of the hydraulic circuit such as the continuously variable transmission 18 detected by the oil temperature sensor 64, and an accelerator that represents the accelerator opening Acc that is the operation amount of the accelerator pedal 68 detected by the accelerator opening sensor 66. opening signal, a brake operation signal indicating whether B _ON operation of the foot brake is a service brake, which is detected by a foot brake switch 70, a lever position (operating position) of a shift lever 74 detected by a lever position sensor 72 P _SH An operation position signal representing the following, a driven hydraulic actuator 46c detected by the hydraulic sensor 75 A belt narrow pressure signal representing the belt narrow pressure Pd is supplied.

Further, the electronic control device 50 receives an engine output control command signal S _E for controlling the output of the engine 12, for example, a throttle signal for driving a throttle actuator 76 for controlling the opening / closing of the electronic throttle valve 30, and a fuel injection device 78. An injection signal for controlling the amount of fuel injected from the engine, an ignition timing signal for controlling the ignition timing of the engine 12 by the ignition device 80, and the like are output. Further, the shift control command signal S _T for example a solenoid valve DS1 (upshift solenoid valve of the present invention for controlling the flow rate of hydraulic oil to the drive side hydraulic actuator 42c for changing the speed ratio γ of the continuously variable transmission 18 equivalent) and a command signal for driving the corresponding) to the downshift solenoid valve of the solenoid valve DS2 (the present invention, the clamping pressure control command signal S _B for example, a belt clamping pressure Pd of the order to adjust the clamping pressure of the transmission belt 48 A command signal for driving the linear solenoid valve SLS to regulate pressure, a lockup control command signal for controlling the engagement, release, and slip amount of the lockup clutch 26, for example, the valve of the lockup control valve in the hydraulic control circuit 100 A command signal or lock-up for driving an ON / OFF solenoid valve DSU (not shown) that switches the position. A command signal for driving the solenoid valve DS2 for adjusting the torque capacity of the clutch 26, a linear solenoid valve SLT for controlling the line hydraulic pressure PL, a command signal for driving the linear solenoid valve SLS, and the like are output to the hydraulic control circuit 100. The

  The shift lever 74 is arranged, for example, in the vicinity of the driver's seat and is sequentially positioned in five lever positions “P”, “R”, “N”, “D”, and “L” (see FIG. 3). Any one of them is manually operated.

  The “P” position (range) releases the power transmission path of the vehicle drive device 10, that is, a neutral state (neutral state) where the power transmission of the vehicle drive device 10 is interrupted, and is mechanically output by the mechanical parking mechanism. The parking position (position) for preventing (locking) the rotation of 44, the “R” position is the reverse traveling position (position) for reversely rotating the output shaft 44, and the “N” position. Is a neutral position (position) for setting the neutral state in which the power transmission of the vehicle drive device 10 is interrupted, and the “D” position establishes an automatic transmission mode within a transmission range that allows the transmission of the continuously variable transmission 18. This is a forward travel position (position) that allows automatic shift control to be executed, and the “L” position is operated by a strong engine brake. An engine brake position (position). As described above, the “P” position and the “N” position are non-traveling positions that are selected when the power transmission path is in the neutral state and the vehicle is not traveling, and the “R” position, the “D” position, and the “L” position. Is a travel position that is selected when the vehicle travels with the power transmission path in a power transmission enabled state that enables power transmission through the power transmission path.

FIG. 3 shows the main part of the hydraulic control circuit 100 relating to the belt clamping pressure control, the transmission gear ratio control of the continuously variable transmission 18, and the engagement hydraulic pressure control of the forward clutch C1 or the reverse brake B1 accompanying the operation of the shift lever 74. FIG. In FIG. 3, a hydraulic pressure control circuit 100 adjusts a belt clamping pressure Pd that is the hydraulic pressure of the driven hydraulic actuator 46c so that the transmission belt 48 does not slip. Equivalent), a first position for outputting the control hydraulic pressure P _SLT as the first hydraulic pressure regulated by the linear solenoid valve SLT, and an output hydraulic pressure P _LM2 as the second hydraulic pressure from the line pressure modulator NO. A clutch apply control valve 112 that functions as a switching valve that is switched to the second position, and a speed change control valve that controls the flow rate of hydraulic oil to the drive side hydraulic actuator 42c so that the speed ratio γ is continuously changed. Gear ratio control valve UP114 (corresponding to the gearshift control valve of the present invention) and gear ratio control Lub DN 116 (corresponding to the transmission control valve of the present invention), thrust ratio control valve 118 (corresponding to the control valve of the present invention) having a predetermined ratio between the transmission pressure Pin and the belt clamping pressure Pd, the forward clutch A manual valve 120 or the like that mechanically switches the oil passage according to the operation of the shift lever 74 is provided so that C1 and the reverse brake B1 are engaged or released.

The line oil pressure PL is, for example, a relief-type primary regulator valve (pressure regulating valve) using, as a source pressure, the working oil pressure output (generated) from a mechanical oil pump 28 (see FIG. 1) that is rotationally driven by the engine 12. It adapted to be pressure adjusted to a value corresponding to the engine load and the like based on the signal pressure P _SLS from the signal pressure P _SLT or the linear solenoid valve SLS from the linear solenoid valve SLT by 124. The output hydraulic pressure _PLM2 is regulated based on the signal pressure P _SLT from the linear solenoid valve _SLT or the signal pressure P _SLS from the linear solenoid valve SLS by the line pressure modulator NO. It has come to be. The output oil pressure P _LM3 is a _source pressure of the control oil pressure (signal pressure) P _SLT and the signal pressure P _SLS , and is adjusted to a constant pressure by the line pressure modulator NO. It has come to be. Modulator pressure P _M is a used as the basic pressure of the electronic control unit 50 controls oil pressure P _DS2 is the output hydraulic pressure of the control oil pressure P _DS1 and the solenoid valve DS2 which is the output hydraulic pressure of the solenoid valve DS1 that is duty-controlled by, The output hydraulic pressure _PLM3 is adjusted to a constant pressure by the modulator valve 128 using the original pressure.

Wherein the manual valve 120, the engagement pressure _{P A} that is output from the clutch apply control valve 112 is supplied to the input port 120a. When the shift lever 74 is operated to the "D" position or "L" position, and for reverse engagement pressure P _A is supplied to the forward clutch C1 via a forward output port 120f as forward running output pressure The oil passage of the manual valve 120 is switched so that the hydraulic oil in the brake B1 is drained (discharged) to the atmospheric pressure, for example, from the reverse output port 120r through the discharge port EX, and the forward clutch C1 is engaged. The reverse brake B1 is released.

Also, operating the shift lever 74 when it is operated to the "R" position, the engagement pressure P _A is the reverse in the output port 120r are supplied to the reverse brake B1 via the and the forward clutch C1 as reverse running output pressure The oil passage of the manual valve 120 is switched so that the oil is drained (discharged) from the forward output port 120f through the discharge port EX to, for example, atmospheric pressure, the reverse brake B1 is engaged, and the forward clutch C1 is engaged. Be released.

  When the shift lever 74 is operated to the “P” position and the “N” position, the oil path from the input port 120a to the forward output port 120f and the oil path from the input port 120a to the reverse output port 120r are both And the oil passage of the manual valve 120 is switched so that the hydraulic oil in the forward clutch C1 and the reverse brake B1 is drained from the manual valve 120, and both the forward clutch C1 and the reverse brake B1 are connected. Be released.

The clutch apply control valve 112, the control hydraulic pressure _{P SLT} is supplied to the manual valve 120 as an engagement pressure _{P A} via the output port 112s from the input port 112i to and signal pressure _{P SLS} by being movable in the axial direction from the input port 112k first position constituting a first oil passage for supplying the input port 112j via an output port 112t to the line pressure modulator NO.2 valve 122 and the primary regulator valve 124 and (CONTROL position) the output hydraulic pressure _{P LM2} supplied through the output port 112t of the engagement pressure _{P a} is supplied to the manual valve 120 as a and the control pressure _{P SLT} through an output port 112s from the input port 112i to the line pressure modulator NO.2 valve 122 and the primary regulator valve 124 A spool valve element 112a positioned at a second position (NORMAL position) constituting the second oil passage, and a spring 112b as an urging means for urging the spool valve element 112a toward the second position side. And an oil chamber 112c that receives the control oil pressure _PDS1 to apply a thrust toward the first position to the spool valve element 112a, and a control oil pressure P to apply a thrust toward the first position to the spool valve element 112a. _{And a} diameter difference portion 112d for receiving _DS2 .

In the clutch apply control valve 112 configured as described above, for example, a garage shift in which the shift lever 74 is operated from the “N” position to the “D” position or the “R” position at a predetermined low vehicle speed or when the vehicle is stopped. N → D shift or N → R shift), the control oil pressure P _DS1 exceeding the predetermined pressure is supplied to the oil chamber 112c, and the control oil pressure P _DS2 exceeding the predetermined pressure is supplied to the diameter difference portion 112d. The control hydraulic pressure _PSLT is supplied to the forward clutch C1 or the reverse brake B1 via the manual valve 120 by switching to the first position shown in the right half of the line. Thereby, the engagement transient hydraulic pressure of the clutch C1 and the brake B1 at the time of the garage shift is regulated by the linear solenoid valve SLT as the first electromagnetic valve. For example, the control hydraulic pressure P _SLT is a hydraulic pressure for controlling the transient engagement state of the clutch C1 and the brake B1 in the N → D shift or the N → R shift, and the clutch C1 or the brake B1 is smoothly engaged. The pressure is regulated according to a predetermined rule so that a shock at the time of engagement is suppressed.

Further, for example, when the supply of at least one of the control hydraulic pressure _PDS1 and the control hydraulic pressure _PDS2 is stopped in the steady state in which the clutch C1 and the brake B1 are engaged after the garage shift, the left half of the center line second position side switched to output hydraulic pressure P _LM2 is supplied to the forward clutch C1 or the reverse brake B1 via the manual valve 120 shown. As a result, the engagement of the clutch C1 and the brake B1 after the garage shift is held by the output hydraulic pressure _PLM2 . For example, the output hydraulic pressure P _LM2 is a predetermined hydraulic pressure for _bringing the clutch C1 and the brake B1 into a fully engaged state, and is adjusted to at least a predetermined constant pressure and a hydraulic pressure corresponding to the signal pressure P _SLT. To adjust the pressure.

Thus, the clutch apply control valve 112 controls the hydraulic oil supply passage to the clutch C1 or the brake B1 in order to control the transient engagement state of the forward clutch C1 or the reverse brake B1 during the garage shift. The second solenoid valve is connected to either the first oil passage that supplies the hydraulic pressure P _SLT and the second oil passage that supplies the output hydraulic pressure P _{LM2 in} order to bring the clutch C1 or the brake B1 into a fully engaged state in a steady state. functions as a switching valve for switching on the basis of the control oil pressure _{P DS1} and the control pressure _{P DS2} from the solenoid valve DS1 and the solenoid valve DS2 as.

In the present embodiment, although the description of the two types in the output hydraulic pressure of the linear solenoid valve SLT _{and the} control pressure P _SLT and the signal pressure P _SLT, the engagement transition oil pressure at the time of garage shifting the control pressure P _SLT The pilot pressure for regulating the line oil pressure PL is clearly distinguished and used as the signal pressure _PSLT . That is, the linear solenoid valve SLT outputs the control hydraulic pressure P _SLT to control the transient engagement state of the forward clutch C1 or the reverse brake B1 when the clutch apply control valve 112 is switched to the first position. When the clutch apply control valve 112 is switched to the second position, the signal pressure _PSLT is output to regulate the line oil pressure PL. The signal pressure _PSLT is a pilot pressure for adjusting the line hydraulic pressure PL by the primary regulator valve 124, and is directly supplied to the hydraulic actuators of the engaging devices in order to engage the clutch C1 or the brake B1. Therefore, the hydraulic pressure is smaller than the output hydraulic pressure _PLM2 .

The transmission ratio control valve UP114 is provided so as to be movable in the axial direction, so that the line hydraulic pressure PL can be supplied from the input port 114i to the driving pulley 42 via the input / output port 114j, and the input / output port 114k is closed. The spool valve element 114a is positioned at the original position where the position and the driving pulley 42 are communicated with the input / output port 114k via the input / output port 114j, and the spool valve element 114a is biased toward the original position side. A spring 114b as an urging means, an oil chamber 114c that accommodates the spring 114b and receives the control hydraulic pressure _PDS2 to _apply a thrust toward the original position to the spool valve element 114a, and an upshift to the spool valve element 114a the control oil pressure P _DS1 to apply a thrust force toward the position side And an only put the oil chamber 114d.

Further, the transmission ratio control valve DN116 is provided so as to be movable in the axial direction, whereby a downshift position where the input / output port 116j communicates with the discharge port EX and an original position where the input / output port 116j communicates with the input / output port 116k. A spool valve element 116a positioned at the first position, a spring 116b as an urging means for urging the spool valve element 116a toward the original position, and a spring 116b that accommodates the spool valve element 116a in the original position side. An oil chamber 116c that receives the control hydraulic pressure _PDS1 to apply a thrust toward the engine, and an oil chamber 116d that receives the control hydraulic pressure _PDS2 to apply a thrust toward the downshift position to the spool valve element 116a. .

  In the transmission ratio control valve UP114 and the transmission ratio control valve DN116 thus configured, in the closed state in which the spool valve element 114a is held in the original position in accordance with the urging force of the spring 114b as shown in the left half of the center line, The input / output port 114j and the input / output port 114k are communicated with each other, and the hydraulic fluid of the drive pulley 42 (drive hydraulic actuator 42c) is allowed to flow to the input / output port 116j. In the closed state in which the spool valve element 116a is held in the original position according to the urging force of the spring 116b as shown in the right half of the center line, the input / output port 116j and the input / output port 116k are communicated with each other, and the thrust ratio The thrust ratio control hydraulic pressure Pτ (corresponding to a predetermined hydraulic pressure of the control valve of the present invention) from the control valve 118 is allowed to flow to the input / output port 114k.

Further, when the control oil pressure _PDS1 is supplied to the oil chamber 114d, the spool valve element 114a is increased against the urging force of the spring 114b by a thrust according to the control oil pressure _PDS1 as shown in the right half of the center line. is moved to the shift position side, the line pressure PL is supplied to the control oil pressure P _DS1 to the corresponding flow rate through the output port 114j from the input port 114i drive side hydraulic actuator 42c, input-output port 114k is cut off The flow of hydraulic oil to the gear ratio control valve DN116 side is prevented. As a result, the transmission pressure Pin is increased, the V groove width of the drive pulley 42 is reduced, and the transmission ratio γ is reduced, that is, the continuously variable transmission 18 is upshifted.

When the control oil pressure _PDS2 is supplied to the oil chamber 116d, the spool valve element 116a is lowered against the urging force of the spring 116b by the thrust according to the control oil pressure _PDS2 , as shown in the left half of the center line. The hydraulic fluid of the drive side hydraulic actuator 42c is discharged from the input / output port 114j through the input / output port 114k and further through the input / output port 116j from the discharge port EX at a flow rate corresponding to the control hydraulic pressure _PDS2 . As a result, the transmission pressure Pin is reduced, the V groove width of the drive pulley 42 is increased, and the transmission ratio γ is increased, that is, the continuously variable transmission 18 is downshifted.

As described above, when the control hydraulic pressure _PDS1 is output, the line hydraulic pressure PL input to the transmission ratio control valve UP114 is supplied to the drive side hydraulic actuator 42c, and the transmission pressure Pin is increased and continuously upshifted. When the hydraulic pressure _PDS2 is output, the hydraulic fluid of the drive side hydraulic actuator 42c is discharged from the discharge port EX, and the shift pressure Pin is lowered and continuously downshifted.

For example, as shown in FIG. 4, an actual value is obtained from a previously stored relationship (shift map) between the vehicle speed V and the target input shaft rotational speed Nin ^* , which is the target input rotational speed of the continuously variable transmission 18, using the accelerator opening Acc as a parameter. actual input shaft rotational speed and the target input shaft rotation speed Nin ^* of the input shaft 36 as a predetermined rotation member to be set on the basis of the vehicle condition represented by the vehicle speed V and the accelerator opening Acc (hereinafter, referred to as the actual input shaft rotational speed ) Shifting of the continuously variable transmission 18 is executed by feedback control in accordance with the rotational speed difference (deviation) ΔNin (= Nin ^* −Nin) so that Nin matches, that is, the drive side hydraulic actuator 42c. As the hydraulic oil is supplied and discharged, the V groove widths of both variable pulleys 42 and 46 are changed, and the speed ratio γ is continuously changed by feedback control.

The shift map in FIG. 4 corresponds to a shift condition, and a target input shaft rotational speed Nin ^* that sets a larger gear ratio γ as the vehicle speed V is smaller and the accelerator opening Acc is larger is set. Further, since the vehicle speed V corresponds to the output shaft rotational speed Nout, the target input shaft rotational speed Nin ^*, which is the target value of the input shaft rotational speed Nin, corresponds to the target gear ratio γ ^* (= Nin ^* / Nout). It is determined within the range of the minimum speed ratio γmin and the maximum speed ratio γmax of the stage transmission 18.

However, although the target input shaft rotational speed Nin ^* may be set as it is as the target value of the input shaft rotational speed Nin, it is preferable to travel during acceleration travel, deceleration travel (coast travel), or during shift transitions. A basic target input shaft rotational speed Ninc, which is a value obtained by processing the target input shaft rotational speed Nin ^* according to the state, is set, and the final target value of the input shaft rotational speed Nin is set based on the basic target input shaft rotational speed Ninc. The transient target input shaft rotational speed Nint is set. Accordingly, in this case, the transient target input shaft rotational speed Nint and the actual input shaft rotational speed Nin are continuously variable in accordance with their rotational speed difference (deviation) ΔNtnt (= Nint−Nin). The shift of the transmission 18 is executed by feedback control.

Further, the control hydraulic pressure _PDS1 is supplied to the oil chamber 116c of the transmission ratio control valve DN116, and regardless of the control hydraulic pressure _PDS2 , the transmission ratio control valve DN116 is closed to limit the downshift, while the control hydraulic pressure _{PDS2 changes} the speed. The oil ratio is supplied to the oil chamber 114c of the ratio control valve UP114, and regardless of the control oil pressure _PDS1 , the transmission ratio control valve UP114 is closed to prohibit the upshift. That is, the control when the hydraulic _{P DS1} and the control pressure _{P DS2} are not supplied together but of course, also, the speed change ratio control valve UP114 and speed ratio control valve when the control oil pressure _{P DS1} and the control pressure _{P DS2} is supplied together Each of the DNs 116 is in a closed state held in its original position. As a result, one of the solenoid valves DS1 and DS2 does not function due to a failure in the electrical system, and a sudden upshift occurs even when the control hydraulic pressure _PDS1 or the control hydraulic pressure _PDS2 continues to be output at the maximum pressure. It is possible to prevent a downshift or a belt slip due to the sudden shift.

The pinching pressure control valve 110 is provided so as to be movable in the axial direction, thereby opening and closing the input port 110i to transfer the line oil pressure PL from the input port 110i to the driven pulley 46 and the thrust ratio control valve 118 via the output port 110t. The spool valve element 110a that can supply the clamping pressure Pd, the spring 110b as an urging means that urges the spool valve element 110a in the valve opening direction, and the spring 110b is accommodated and opened in the spool valve element 110a. an oil chamber 110c that receives the control oil pressure P _SLS to apply thrust direction, feedback oil chamber for receiving the belt clamping pressure Pd output from the output port 110t to apply thrust in the valve closing direction to the spool valve element 110a 110d and the thrust in the valve closing direction is applied to the spool valve element 110a. And an oil chamber 110e that accepts modulator pressure _{P M} in order.

In the pinching pressure control valve 110 configured as described above, the line oil pressure PL is continuously regulated by using the control oil pressure _PLSS as a pilot pressure so that the transmission belt 48 does not slip. The belt clamping pressure Pd is output.

For example, as shown in FIG. 5, the accelerator opening degree Acc corresponding to the transmission torque is used as a parameter, and is experimentally obtained and stored in advance so as not to cause belt slip between the gear ratio γ and the required hydraulic pressure (belt clamping pressure) Pd ^*. From the relationship (belt clamping pressure map), the belt of the driven hydraulic actuator 46c is obtained so that the belt clamping pressure Pd ^* determined (calculated) based on the vehicle state indicated by the actual gear ratio γ and the accelerator opening Acc is obtained. The clamping pressure Pd is controlled, and the belt clamping pressure, that is, the frictional force between the variable pulleys 42 and 46 and the transmission belt 48 is increased or decreased according to the belt clamping pressure Pd.

  The thrust ratio control valve 118 is provided so as to be movable in the axial direction, thereby opening and closing the input port 118i and supplying the line hydraulic pressure PL from the input port 118i to the transmission ratio control valve DN116 via the output port 118t. A spool valve element 118a that enables supply, a spring 118b as an urging means that urges the spool valve element 118a in the valve opening direction, and a spring 118b that accommodates the spring 118b and applies a thrust in the valve opening direction to the spool valve element 118a. An oil chamber 118c that receives the belt clamping pressure Pd for application and a feedback oil chamber 118d that receives the thrust ratio control hydraulic pressure Pτ output from the output port 118t for applying thrust in the valve closing direction to the spool valve element 118a. I have.

In the thrust ratio control valve 118 configured in this manner, the pressure receiving area of the belt clamping pressure Pd in the oil chamber 118c is a, the pressure receiving area of the thrust ratio control hydraulic pressure Pτ in the feedback oil chamber 118d is b, and the biasing force of the spring 118b is FS. Then, it will be in an equilibrium state by following Formula (1).
Pτ × b = Pd × a + FS (1)
Therefore, the thrust ratio control hydraulic pressure Pτ is expressed by the following equation (2) and is proportional to the belt clamping pressure Pd.
Pτ = Pd × (a / b) + FS / b (2)

Then, the control oil pressure _P or _DS1 and the control pressure _{P DS2} is not supplied together, or the predetermined pressure or more control pressure _{P DS1} and the predetermined pressure or more control pressure _{P DS2} is both supplied, the speed ratio control valve UP114 and the speed ratio control When the valve DN116 is in the closed state in which both are held in their original positions, the thrust ratio control hydraulic pressure Pτ is supplied to the drive side hydraulic actuator 42c, so that the transmission pressure Pin is made to coincide with the thrust ratio control hydraulic pressure Pτ. . That is, the thrust ratio control valve 118 outputs the thrust ratio control hydraulic pressure Pτ, that is, the transmission pressure Pin, which maintains the ratio between the transmission pressure Pin and the belt clamping pressure Pd in a predetermined relationship.

For example, because of the accuracy of the input shaft rotational speed sensor 56 and the vehicle speed sensor 58, the detection accuracy of the input shaft rotational speed Nin and the vehicle speed V is inferior in a low vehicle speed state below a predetermined vehicle speed V ′. sometimes instead feedback control of the gear ratio γ for eliminating the rotation speed difference (deviation) ΔNin, for example the speed ratio control valve UP114 and speed ratio control valve DN116 without supplying both the control pressure P _DS1 and the control pressure P _DS2 In either case, so-called closing control is executed to make the closing state. As a result, the shift pressure Pin proportional to the belt clamping pressure Pd is supplied to the drive-side hydraulic actuator 42c so that the ratio between the transmission pressure Pin and the belt clamping pressure Pd becomes a predetermined relationship during low vehicle speed traveling or starting. Thus, the belt slip of the transmission belt 48 from the time when the vehicle is stopped to the extremely low vehicle speed is prevented, and at this time, for example, the thrust ratio τ corresponding to the maximum gear ratio γmax (= driven hydraulic actuator thrust Wout / drive hydraulic actuator thrust) Win; Wout is the belt clamping pressure Pd × the pressure receiving area of the driven hydraulic actuator 46c, and Win is the transmission pressure Pin × the pressure receiving area of the driving hydraulic actuator 42c). When the first term (a / b) or FS / b is set, a good start is performed at the maximum gear ratio γmax or a gear ratio γmax ′ in the vicinity thereof. The predetermined vehicle speed V ′ is a lower limit vehicle speed at which a predetermined feedback control can be executed as a vehicle speed V at which the rotational speed of the predetermined rotating member, for example, the input shaft rotational speed Nin cannot be detected. For example, it is set to about 2 km / h.

  FIG. 6 shows the relationship obtained and stored in advance between the speed ratio γ and the thrust ratio τ using the vehicle speed V as a parameter, and the first term (a It is a figure which shows an example when / b) is set. The parameter of the vehicle speed V shown by the one-dot chain line in FIG. 6 is a thrust ratio τ calculated in consideration of the centrifugal hydraulic pressure in the drive side hydraulic actuator 42c and the driven side hydraulic actuator 46c, and the intersections with the solid lines (V0, V20, V50). To obtain a transmission gear ratio γ as a predetermined transmission gear ratio that can be held during the closing control. For example, as shown in FIG. 6, in the continuously variable transmission 18 according to this embodiment, the maximum speed ratio γmax can be maintained as a predetermined speed ratio when the vehicle speed V is 0 km / h, that is, the closing control is performed while the vehicle is stopped. .

FIG. 7 is a functional block diagram for explaining the main part of the control function by the electronic control unit 50. In FIG. 7, the target input rotation setting means 150 is a target of the input shaft rotation speed Nin based on the vehicle state indicated by the actual vehicle speed V and the accelerator opening Acc from a shift map stored in advance as shown in FIG. Set the input shaft rotation speed Nin ^* sequentially.

The shift control means 152 is arranged so that the actual input shaft rotation speed Nin matches the target input shaft rotation speed Nin ^* set by the target input rotation setting means 150, that is, the rotation speed difference (deviation) ΔNin (= Nin ^* −). Nin), the shift of the continuously variable transmission 18 is executed by feedback control in accordance with the rotational speed difference ΔNin. That is, the speed change outputs a shift control command signal (oil pressure command) S _T for changing the V groove widths of both variable pulleys 42 and 46 by controlling the flow rate of the hydraulic oil to the drive side hydraulic actuator 42c to the hydraulic control circuit 100 The ratio γ is continuously changed.

The belt clamping pressure setting means 154, for example, from the belt clamping pressure map obtained and stored experimentally in advance as shown in FIG. 5, for example, the actual accelerator opening Acc and the actual input shaft rotational speed Nin by the electronic control unit 50. The belt clamping pressure Pd ^* is set based on the vehicle state indicated by the actual speed ratio γ (= Nin / Nout) calculated based on the output shaft rotational speed Nout. That is, the belt clamping pressure setting means 154 sets the belt clamping pressure Pd of the output side hydraulic actuator 46c for obtaining the belt clamping pressure Pd ^* .

Belt clamping pressure control means 156, the belt clamping pressure setting means 154 clamping pressure control command signal S _B for pressurizing regulates the belt clamping pressure Pd of the driven side hydraulic actuator 46c as set belt clamping pressure Pd ^* is obtained by and outputs to the hydraulic control circuit 100 increases or decreases the belt clamping pressure Pd ^*.

The hydraulic control circuit 100, the supply of hydraulic oil to the solenoid valve DS1 and operates the solenoid valve DS2 driven side hydraulic actuator 42c as shifting is executed in the continuously variable transmission 18 according to the shift control command signal S _T · to control the emissions, by operating the linear solenoid valve SLS so that the belt clamping pressure Pd ^* is increased or decreased pressure of the belt clamping pressure Pd adjusted in accordance with the above clamping force control command signal S _B.

In addition to the above-described function, the shift control means 152 provides feedback of the transmission ratio γ for eliminating the rotational speed difference ΔNin as the normal shift control on condition that the vehicle speed V is equal to or less than the predetermined vehicle speed V ′. Without performing the control, the thrust ratio control valve 118 executes the closing control for maintaining the ratio between the transmission pressure Pin and the belt clamping pressure Pd in a predetermined relationship. That is, by closing the transmission ratio control valve UP114 and the transmission ratio control valve DN116, the transmission ratio γ of the continuously variable transmission 18 is set to a predetermined transmission ratio with the hydraulic oil confined in the drive side hydraulic actuator 42c. A shift command (closed control command) signal S _T ′ for shifting control for low vehicle speed is output to the hydraulic control circuit 100 to establish a predetermined gear ratio.

The hydraulic control circuit 100 does not operate the solenoid valve DS1 and the solenoid valve DS2 so as to close the transmission ratio control valve UP114 and the transmission ratio control valve DN116 in accordance with the closing control command signal S _T ′, and controls the thrust ratio control. The valve 118 outputs a thrust ratio control hydraulic pressure Pτ that maintains the ratio between the transmission pressure Pin and the belt clamping pressure Pd in a predetermined relationship.

The engine output control means 158 outputs an engine output control command signal S _E , for example, a throttle signal, an injection signal, an ignition timing signal, etc., to the throttle actuator 76, the fuel injection device 78, and the ignition device 80, respectively, for output control of the engine 12. To do. For example, the engine output control means 158 controls the engine torque T _E and outputs a throttle signal for opening and closing the electronic throttle valve 30 such that the throttle opening theta _TH corresponding to the accelerator opening Acc to the throttle actuator 76.

Shift position determining means 160, or determines the current lever position P _SH based on based on the lever position P _SH i.e. ON signal of the lever position P _SH, it determines an operation history of the shift lever 74. For example, the shift position determination unit 160 performs N → D shift determination, N → R shift determination, “D” position determination, “N” position determination, “R” position determination, and the like based on the lever position _PSH .

The engagement control means 162 switches the clutch apply control valve 112 to the first position side and moves forward when the garage shift is determined that the N → D shift or the N → R shift has been performed by the shift position determination means 160. In order to control the transitional engagement state of the clutch C1 or the reverse brake B1, the control oil pressure P _SLT for gradually increasing the engagement oil pressure is output so that the engagement shock is suppressed, and the line oil pressure PL is set. _A garage shift command signal _SA that outputs the signal pressure P _SLS to adjust the pressure is output to the hydraulic control circuit 100. For example, the engagement control unit 162 outputs the engagement transient hydraulic command pressure pcltexc to the hydraulic control circuit 100 as a command signal SLTDUTY for duty-controlling the linear solenoid valve SLT.

Further, the engagement control means 162 is configured to move the forward clutch C1 or the reverse brake in a steady state such as after a garage shift, for example, after a lapse of a predetermined time from the start of the garage shift, or after the control hydraulic pressure _PSLT becomes a predetermined engagement hydraulic pressure or higher. A steady control command for switching the clutch apply control valve 112 to the second position side in order to supply the output hydraulic pressure _PLM2 to B1 and bring it into a fully engaged state, and to output the signal pressure P _SLT to regulate the line hydraulic pressure PL. and it outputs a signal _{S a} 'to the hydraulic control circuit 100. For example, the engagement control unit 162 outputs the line pressure command pressure plctgt to the hydraulic control circuit 100 as a command signal SLTDUTY for duty-controlling the linear solenoid valve SLT.

The shift control means 152 executes a shift of the continuously variable transmission 18 by feedback control so that the actual input shaft rotation speed Nin matches the target input shaft rotation speed Nin ^* set by the target input rotation setting means 150. In other words, in other words, the shift of the continuously variable transmission 18 is executed by feedback control so that the transmission gear ratio γ of the continuously variable transmission 18 matches the preset target transmission gear ratio γ ^*. . Further, since the gear ratio γ is uniquely determined according to the pulley position χ of the driving pulley 42, the stepless step is performed so that the pulley position χ of the driving pulley 42 matches the preset target pulley position χ ^*. It is also agreed that the shift of the transmission 18 is executed by feedback control. In the following description of the control, the object to be fed back in the shifting of the continuously variable transmission 18 will be described with a focus on the pulley position χ. No change.

FIG. 8 shows the upshift of the continuously variable transmission 18 with the pulley position χ of the drive sheave 42 as a reference. Since the pulley position χ and the gear ratio γ are uniquely determined as described above, the pulley position χ may be replaced with the gear ratio γ. The upper side of FIG. 8 shows the shift output (shift command) of the solenoid valve DS1 and the solenoid valve DS2. The solid line indicates the upshift DUTY output from the solenoid valve DS1, and the broken line indicates the downshift output from the solenoid valve DS2. DUTY is shown respectively. Further, the lower side of FIG. 8 shows the positional relationship of the actual pulley position χ indicated by the solid line with respect to the speed change target pulley position χ ^{* as the} speed change target indicated by the broken line.

In FIG. 8, when the upshift DUTY is output, the actual pulley position χ is upshifted toward the target pulley position χ ^* . At this time, hydraulic oil is supplied to the drive side hydraulic actuator 42c, and the movable rotating body 42b is moved in parallel with the axis in the direction in which the V groove width of the drive side sheave 42 is reduced. For example, as shown in FIG. 8, when the actual pulley position χ reaches the shift target pulley position χ ^* and is underlapped, the downshift DUTY is output from the solenoid valve DS2. At this time, the spool valve element 114a of the transmission ratio control valve UP114 is switched to the original position, the spool valve element 116a of the transmission ratio control valve DN116 is switched to the downshift position side, and the hydraulic oil of the drive side hydraulic actuator 42c is shifted. It is discharged from the discharge port EX of the ratio control valve DN116. As a result, the continuously variable transmission 18 is downshifted, and the actual pulley position χ is again moved to the downshift side with respect to the shift target pulley position χ ^* . At this time, as shown in FIG. 8, in the hydraulic control circuit 100 of the present embodiment, the amount of downshift when the downshift command is output tends to increase, and the actual pulley position χ is the shift target pulley position χ. ^* Moved greatly to the downshift side. On the other hand, since the actual pulley position χ is moved to the upshift side, the upshift DUTY is output again. As a result, the output of the upshift DUTY and the downshift DUTY is repeated and hunting occurs. there's a possibility that.

  The reason why the downshift amount increases as described above will be described. When the continuously variable transmission 18 is switched from upshift to downshift, the spool valve element 114a of the transmission ratio control valve UP114 is moved from the upshift position toward the original position as the output of the upshift DUTY is stopped. It is done. In the transition period in which the spool valve element 114a is moved, the input / output port 114j and the input / output port 114k communicate with each other. On the other hand, when the gear ratio control valve DN116 is switched to downshift, a downshift DUTY is output and the spool valve element 116a is moved from the original position to the downshift position side, but the spool valve element 116a is moved to the downshift position. In the transition period of switching, the input / output 116j and the input / output port 116k are in communication with each other. Therefore, the drive side hydraulic actuator 42c and the output port 118t of the thrust ratio control valve 118 are communicated with each other in the transition period to the downshift. Here, when the transmission pressure Pin of the drive side hydraulic actuator 42c is higher than the thrust ratio control oil pressure Pτ of the thrust ratio control valve 118, the hydraulic oil of the drive side hydraulic actuator 42c is discharged to the thrust ratio control valve 118 side. The continuously variable transmission 18 is downshifted. That is, the hydraulic control circuit 100 is configured such that the hydraulic oil is discharged from the drive-side hydraulic actuator 42c even in the transition period to the downshift.

When the shift to the downshift further proceeds, the input / output port 114j of the transmission ratio control valve UP114 and the input / output port 114k are communicated, and the input / output port 116j of the transmission ratio control valve DN116 and the discharge port EX are communicated. As a result, the hydraulic fluid of the drive side hydraulic actuator 42c is discharged from the discharge port EX. When the actual pulley position χ passes through the speed change target pulley position χ ^* , the output of the downshift DUTY is stopped and the upshift DUTY is output, but the responsiveness of the speed ratio control valve UP114 and the speed ratio control valve DN116 is increased. The hydraulic oil discharged from the drive side hydraulic actuator 42c increases with a delay. Further, even during the transition period when the shift from downshift to upshift is performed, the drive side hydraulic actuator 42c and the thrust ratio control valve 118 communicate with each other as described above, so that the hydraulic oil of the drive side hydraulic actuator 42c is controlled by the thrust ratio control. It is discharged to the valve 118 side. As described above, in the hydraulic control circuit 100, when the upshift DUTY is stopped at the time of switching to the downshift, the amount of hydraulic oil discharged from the drive side hydraulic actuator 42c increases. Further, the amount of downshift increases due to leakage of hydraulic oil from the drive side hydraulic actuator 42c. Therefore, in the hydraulic control circuit 100, hunting may occur as the downshift amount increases. In the hydraulic control circuit 100, the input / output port 114j and the input / output port 114k of the transmission ratio control valve UP114 communicate with each other even in a state where the output of the upshift DUTY is lowered. When the transmission pressure Pin of 42c is higher than the thrust ratio control hydraulic pressure Pτ of the thrust ratio control valve 118, the hydraulic fluid of the drive side hydraulic actuator 42c is discharged to the thrust ratio control valve 118 side.

On the other hand, when the actual pulley position χ of the continuously variable transmission 18 during the upshift is moved to the pulley position where the downshift is required with respect to the shift target pulley position χ ^* , in other words, during the upshift. When the gear ratio γ of the continuously variable transmission 18 is moved to the pulley position where a downshift is required with respect to the target gear ratio γ ^* , the gear ratio Pin of the drive side hydraulic actuator 42c and the thrust ratio of the thrust ratio control valve 118 Hunting is prevented by calculating a differential pressure (difference) with the control hydraulic pressure Pτ and optimally controlling the upshift DUTY of the solenoid valve DS1 and the downshift DUTY of the solenoid valve DS2 based on the differential pressure. Hereinafter, the control for preventing the hunting will be described.

  As described above, when the output of the upshift DUTY is decreased in the hydraulic control circuit 100, the spool valve element 114a of the transmission ratio control valve UP114 is moved to the original position side, and the input / output port 114j and the input / output port 114k are moved. Communicated. Therefore, the drive side hydraulic actuator 42c is communicated with the output port 118t of the thrust ratio control valve 118 via the transmission ratio control valve DN116. Here, when the transmission pressure Pin of the drive side hydraulic actuator 42c is higher than the thrust ratio control oil pressure Pτ, the hydraulic oil of the drive side hydraulic actuator 42c is discharged to the thrust ratio control valve 118 side as described above. Therefore, in the hydraulic control circuit 100, even when the downshift DUTY is not output, the hydraulic oil of the drive side hydraulic actuator 42c can be discharged and downshifted by reducing the upshift DUTY. Further, the amount of downshift increases the amount of hydraulic oil discharged from the drive side hydraulic actuator 42c in accordance with the differential pressure Pdiff between the transmission pressure Pin of the drive side hydraulic actuator 42c and the thrust ratio control oil pressure Pτ. Is proportional to

  Therefore, the down-switching control means 164 prevents hunting by optimally controlling the solenoid valve DS1 and the solenoid valve DS2 based on the configuration (characteristics) of the hydraulic control circuit 100. Specifically, the down-switching control means 164 calculates a differential pressure Pdiff between the transmission pressure Pin of the driven hydraulic actuator 42c and the thrust ratio control hydraulic pressure Pτ of the thrust ratio control valve 118, and the calculated differential pressure Pdiff is calculated. Based on this, the solenoid valve DS1 and the solenoid valve DS2 are controlled.

  The hydraulic pressure calculation means 166 outputs from the output port 118 of the thrust ratio control valve 118 configured to be able to communicate with the transmission pressure Pin of the driving hydraulic actuator 42c and the driving hydraulic actuator 42c via the transmission ratio control valve DN116. The thrust ratio control hydraulic pressure Pτ to be calculated is calculated. The shift pressure Pin of the drive side hydraulic actuator 42c is calculated based on the belt narrow pressure Pd detected by the hydraulic sensor 75 of the driven side hydraulic actuator 46c. The transmission pressure Pin of the drive side hydraulic actuator 42c is calculated based on a preset relationship (relation map, relational expression) including, for example, the detected belt narrow pressure Pd, the transmission gear ratio γ, the accelerator opening degree Acc, and the like. Further, the thrust ratio control hydraulic pressure Pτ of the thrust ratio control valve 118 is calculated based on the relationship of the above-described equation (1) and the actual belt narrow pressure Pd.

  The differential pressure calculation means 168 calculates a differential pressure (difference) Pdiff between the transmission pressure Pin of the drive side hydraulic actuator 42c calculated by the hydraulic pressure calculation means 166 and the thrust ratio control hydraulic pressure Pτ of the thrust ratio control valve 118.

  The offset setting means 170 sets a deviation offset α for the pulley position χ or a deviation offset α for the speed ratio γ based on the differential pressure Pdiff calculated by the differential pressure calculation means 168. The deviation offset α set when the down-switching control means 164 is implemented is substantially the same regardless of whether the pulley position χ or the gear ratio γ is used as a reference. Is a common α. However, specific numerical values are appropriately changed according to the selected variables (pulley position χ, gear ratio γ).

  FIG. 9 is a diagram showing the relationship between the differential pressure Pdiff and the deviation offset α. The deviation offset α is set based on either the gear ratio γ or the pulley position χ. However, since the gear ratio γ and the pulley position χ have a uniquely corresponding relationship, any deviation offset α may be used. As shown in FIG. 9, the differential pressure Pdiff and the deviation offset α are in a proportional relationship. When the differential pressure Pdiff increases, when the drive side hydraulic actuator 42c and the thrust ratio control valve 118 are communicated with each other, more hydraulic oil is discharged from the drive side hydraulic actuator 42c to the thrust ratio control valve 118, resulting in a downshift. The amount increases. Therefore, the deviation offset α is proportional to the amount of downshift.

When switching from upshift to downshift, the down-switching time control means 164 provides a downshift DUTY output line (one-dot chain line in FIG. 10) serving as a boundary line for outputting the downshift DUTY based on the deviation offset α. change. Specifically, when the deviation offset α is set, as shown in FIG. 10, the down-switching-time control means 164 causes the downshift DUTY output line indicated by the conventional broken line (which coincides with the shift target pulley position χ ^* ). On the other hand, the downshift DUTY output line is changed to the downshift DUTY output area side (downward in FIG. 10) by the deviation offset α as indicated by a one-dot chain line. Therefore, the upshift DUTY output area is enlarged according to the deviation offset α. Further, when the speed ratio γ is used as a reference, the down-switching control means 164 sets the downshift DUTY output line at one point with respect to the downshift DUTY output line indicated by the conventional broken line (which coincides with the target speed ratio γ ^* ). As indicated by the chain line, the shift is changed to the downshift DUTY output region side (downward in FIG. 10) by the deviation offset α.

Referring to FIG. 10, when the actual pulley position χ (speed ratio γ) of the driving pulley 42 is upshifted toward the speed change target pulley position χ ^* (target speed ratio γ ^* ), the deviation offset α is set. 6 is a time chart showing a shift state in a case where The following description is based on the pulley position χ. However, when the transmission gear ratio γ is used as a reference, the pulley position χ is read as the transmission gear ratio γ. As described above, the downshift DUTY output line is moved to the downshift DUTY output area side based on the deviation offset α. When the deviation offset α is set, the down-switching control means 164 calculates a deviation χdiff (= χ ^* −χ) between the current pulley position χ and the speed change target pulley position χ ^* . When the deviation χdiff is a negative region (in FIG. 10, above the shift target pulley position χ ^* indicated by the broken line), normal upshift transmission control according to the deviation χdiff is performed.

Further, when it is determined that the deviation χdiff has a positive value, the down-switching time control means 164 determines whether or not the deviation χdiff is smaller than the deviation offset α. When the above is affirmed, it is determined in FIG. 10 that the actual pulley position χ is in the upshift output region between the speed change target pulley position χ ^* and the downshift DUTY output line. At this time, the down-switching time control means 164 executes a downshift by reducing the upshift DUTY to a predetermined output and discharging the hydraulic oil of the drive side hydraulic actuator 42c to the thrust ratio control valve 118 side. From the above, the predetermined output of the upshift DUTY is communicated between the input / output port 114j and the input / output port 114k in the transmission ratio control valve UP114, and the drive side hydraulic actuator 42c is connected via the transmission ratio control valve DN116. It is set to such an extent that it communicates with the thrust ratio control valve 118.
On the basis of the gear ratio γ, the down-switching control means 164 calculates a deviation γdiff between the actual gear ratio γ and the target gear ratio γ ^*, and when the deviation γdiff is smaller than the deviation offset α, the upshift DUTY Is equivalent to performing a downshift by reducing the hydraulic pressure of the drive side hydraulic actuator 42c to the thrust ratio control valve 118 side.

Here, FIG. 10 shows a case where the actual pulley position χ does not exceed the downshift DUTY output line indicated by the alternate long and short dash line when the continuously variable transmission 18 is upshifted. In FIG. 10, even if the actual pulley position χ exceeds the speed change target pulley position χ ^* , the actual pulley position χ is downshifted by reducing the output of the upshift DUTY. That is, if the actual pulley position χ passes through the speed change target pulley position χ ^* and is in the conventional range where the downshift DUTY is output, but is within the range not exceeding the deviation offset α, the downshift DUTY is output. Not.

  The control of the down switching control means 164 at this time will be described in detail. When the actual pulley position χ enters the region in which the deviation offset α is set, the down switching control unit 164 outputs a command for reducing the output of the upshift DUTY to the shift control unit 152. Accordingly, in the hydraulic control circuit 100 of FIG. 3, the spool valve element 114a of the transmission ratio control valve UP114 is moved to the original position side, whereby the input / output port 114j and the input / output port 114k are communicated to drive. The side hydraulic actuator 42c communicates with the output port 118t of the thrust ratio control valve 118 via the transmission ratio control valve DN116. At this time, the hydraulic oil of the drive side hydraulic actuator 42c is discharged to the thrust ratio control valve 118 side, so that the continuously variable transmission 18 is downshifted.

  In the above control, the downshift amount is smaller than that when the downshift DUTY is output. Further, when an upshift is requested again, the upshift DUTY is output, so that a quick upshift is possible. Therefore, the downshift DUTY from the solenoid valve DS2 is suppressed, the fluctuation of the actual pulley position χ is reduced, and hunting is prevented.

On the other hand, when the deviation χdiff calculated by the down-switching control means 164 is larger than the deviation offset α, a downshift DUTY corresponding to the deviation χdiff is output. As shown in FIG. 11, when the shift target pulley position χ ^* is passed at time t1 and further exceeds the downshift DUTY output line moved to the upshift side by the cutoff value α, the deviation χdiff is larger than the deviation offset α. It becomes big. In such a state, it is determined that the required downshift amount cannot be obtained by the downshift due to the output decrease of the upshift DUTY. At this time, the down-switching time control means 164 outputs a command to output a downshift DUTY corresponding to the deviation χdiff to the shift control means 152 from time t1 to time t2. Accordingly, the actual pulley position χ is greatly moved toward the downshift side to follow the speed change target pulley position χ ^* . If the gear ratio γ is used as a reference, the down-switching control means 164 determines that the downshift DUTY is made according to the deviation γdiff when the deviation γdiff between the actual gear ratio γ and the target gear ratio γ ^* is larger than the deviation offset α. A command to be output is output to the shift control means 152.

  As described above, the down-switching control means 164 changes the downshift DUTY output line for outputting the downshift DUTY based on the deviation offset α set in accordance with the differential pressure Pdiff, and the deviation χdiff becomes the deviation offset α. The downshift DUTY is output only when the actual pulley position χ exceeds the changed downshift DUTY output line.

  Here, the cut-off value α is set based on the above-described experimentally set relationship of FIG. 9, but FIG. 9 shows that the downshift amount increases as the differential pressure Pdiff increases. This means that the cut-off value α increases as the amount increases. When the amount of downshift when the output of the upshift DUTY is reduced is large, a sufficient downshift can be achieved by the output reduction of the upshift DUTY without outputting the downshift DUTY. In such a case, it is possible to widen the upshift DUTY output region, and the cutoff value α is set large. Specifically, in the region up to the downshift DUTY output line that is changed according to the cut-off value α, it is set experimentally or analytically so that it can be implemented by downshift due to the output decrease of the upshift DUTY. The At this time, the deviation cutoff value α may be set in consideration of leakage of hydraulic oil from the drive side hydraulic actuator 42c.

  FIG. 12 is a flowchart for explaining a main part of the control operation of the electronic control unit 50, that is, a control operation that can prevent hunting that occurs at the time of switching from upshift to downshift. It is repeatedly executed with an extremely short cycle time of about msec.

First, in step SA1 (hereinafter, step is omitted) corresponding to the differential pressure calculation means 168, a differential pressure Pdiff between the transmission pressure Pin of the drive side hydraulic actuator 42c and the thrust ratio control hydraulic pressure Pτ of the thrust ratio control valve 118 is calculated. . Next, in SA2 corresponding to the offset setting means 170, the deviation offset α with respect to the pulley position χ of the driving pulley 42 or the gear ratio γ of the continuously variable transmission 18 is set based on the differential pressure Pdiff and the relationship map shown in FIG. Is done. In SA3 corresponding to the down-switching control means 164, a deviation χdiff between the target pulley position χ ^* and the actual pulley position χ is calculated, and it is determined whether or not the deviation χdiff is a negative value. When the gear ratio γ is used as a reference, a deviation γdiff between the target gear ratio γ ^* and the actual gear ratio γ is calculated, and it is determined whether or not the deviation γdiff is a negative value. When SA3 is positive, it is determined that the actual pulley position χ (or gear ratio γ) is in the upshift DUTY output region. Based on the upshift DUTY is output.

  On the other hand, if SA3 is negative, it is determined in SA4 corresponding to the down-switching control means 164 whether or not the deviation χdiff (deviation γdiff) calculated in SA3 is smaller than the deviation offset α calculated in SA2. The When SA2 is negative, the deviation χdiff (deviation γdiff) exceeds the deviation offset α, so that the current pulley position χ (gear ratio γ) passes through the downshift DUTY output line and is in the downshift DUTY output region. And a downshift DUTY based on the deviation χdiff (deviation γdiff) is output. On the other hand, when SA4 is affirmed, the deviation χdiff (deviation γdiff) is within the deviation offset α, so it is determined that it does not exceed the downshift DUTY output line. In SA5 corresponding to the down-switching control means 164, A preset low output upshift DUTY is output. Accordingly, the hydraulic oil of the drive side hydraulic actuator 42c is discharged to the thrust ratio control valve 118 side, whereby the continuously variable transmission 18 is downshifted.

As described above, according to this embodiment, when the continuously variable transmission 18 is switched from upshift to downshift, the differential pressure calculation means 168 communicates with the drive side hydraulic actuator 42c via the speed ratio control valve DN116. The down-switching control means 164 calculates a differential pressure Pdiff between the thrust ratio control hydraulic pressure Pτ output from the output port 118t of the thrust ratio control valve 118 configured to be possible and the shift pressure Pin of the drive side hydraulic actuator 42c. Since the solenoid valve DS1 and the solenoid valve DS2 are controlled based on the differential pressure Pdiff, hunting that occurs during the speed change of the continuously variable transmission 18 can be suitably prevented. For example, during the upshift, the pulley position χ (speed ratio γ) of the continuously variable transmission 18 reaches the pulley position χ (speed ratio γ) on the downshift side with respect to the speed change target pulley position χ ^* (target speed ratio γ ^* ). Even if it is changed, the downshift DUTY is not output, and the output of the upshift DUTY is controlled so that the drive side hydraulic actuator 42c and the output port 118t of the thrust ratio control valve 118 are communicated. As a result, the hydraulic oil of the drive side hydraulic actuator 42c is discharged to the thrust ratio control valve 118 side, the V groove width of the drive side pulley 42 is expanded, and as a result, it is downshifted. Therefore, even if the downshift DUTY is not output, the downshift can be performed within a certain range. From the above, since the output of the downshift DUTY can be suppressed, hunting can be suppressed.

  Further, the optimum downshift is performed by determining whether to continue the upshift DUTY or to output the downshift DUTY based on the differential pressure Pdiff. For example, when the differential pressure Pdiff is large, when the drive side hydraulic actuator 42c and the thrust ratio control valve 118 communicate with each other, the flow rate of the hydraulic oil discharged from the drive side hydraulic actuator 42c to the thrust ratio control valve 118 side increases. Thus, even if the downshift DUTY is not output, a large downshift can be obtained by appropriately controlling the upshift DUTY and communicating the above. Therefore, hunting is prevented as the downshift DUTY is suppressed by implementing a suitable downshift in consideration of the downshift amount based on the differential pressure Pdiff and the current shift state.

Further, according to the present embodiment, the deviation offset α of the actual pulley position χ (speed ratio γ) with respect to the speed change target pulley position χ ^* (target speed ratio γ ^* ) is set based on the differential pressure Pdiff, and the down-switching is performed. The time control means 164 changes the downshift DUTY output line that becomes the boundary line from which the downshift DUTY is output from the solenoid valve DS2 in accordance with the deviation offset α, so that the suitable downshift is based on the changed downshift output line. A shift is performed.

Further, according to the present embodiment, the down-switching control means 164 determines the deviation χdiff (deviation γdiff) between the actual pulley position χ (actual transmission ratio γ) and the transmission target pulley position χ ^* (target transmission ratio γ ^* ). If the deviation (χdiff, γdiff) is smaller than the deviation offset α, a preset upshift DUTY by the solenoid valve DS1 is output, and if the deviation (χdiff, γdiff) is larger than the deviation offset α, the deviation Since the downshift DUTY is output from the solenoid valve DS2 in accordance with (χdiff, γdiff), a suitable downshift is performed by comparing the deviation (χdiff, γdiff) with the deviation offset α.

  Further, according to the present embodiment, when the deviation (χdiff, γdiff) is smaller than the deviation offset α, the down-switching time control means 164 reduces the output of the solenoid valve DS1 and the thrust ratio with the drive side hydraulic actuator 42c. In order to communicate with the output port 118t of the control valve 118, the hydraulic oil of the drive side hydraulic actuator 42c is discharged to the thrust ratio control valve 118 side, and the V groove width of the drive side pulley 42 is expanded. Therefore, by reducing the output of the solenoid valve DS1, the downshift can be performed without outputting the downshift DUTY.

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

  For example, in the above-described embodiment, the hydraulic pressure of the driving hydraulic actuator 42c and the thrust ratio control hydraulic pressure Pτ of the thrust ratio control valve 118 are calculated based on the belt narrow pressure Pd of the driven hydraulic actuator 42c. It may be detected by a sensor.

  In the above-described embodiment, the hydraulic control circuit 100 is configured such that the drive side hydraulic actuator 42c and the thrust ratio control valve 118 communicate with each other when the upshift DUTY is lowered. The ratio control valve 118 is not limited. Specifically, when the upshift DUTY is lowered, the present invention can be applied as long as it is configured to communicate with a predetermined control valve or the like whose hydraulic pressure is lower than the transmission pressure Pin of the drive side hydraulic actuator 42. .

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

18: continuously variable transmission for vehicle 42: driving pulley 42c: driving hydraulic actuator 46: driven pulley 48: transmission belt (belt)
114: Transmission ratio control valve UP (transmission control valve)
116: Transmission ratio control valve DN (transmission control valve)
118: Thrust ratio control valve (control valve)
118t: Output port 152: Shift control means 164: Down switching control means 168: Differential pressure calculation means DS1: Solenoid valve (upshift solenoid valve)
DS2: Solenoid valve (solenoid valve for downshift)
Pτ: Thrust ratio control hydraulic pressure (predetermined hydraulic pressure of control valve)
Pin: Speed change pressure (hydraulic pressure of drive side hydraulic actuator)
Pdiff: Differential pressure α: Deviation offset γ: Gear ratio γ ^* : Target gear ratio γdiff: Deviation χ: Pulley position χ ^* : Shift target pulley position χdiff: Deviation

Claims (7)

A continuously variable transmission for a vehicle having a driving pulley and a driven pulley and a belt wound around both pulleys, a driving hydraulic actuator for changing a groove width of the driving pulley, and a driving hydraulic pressure thereof A shift control valve that changes a gear ratio of the continuously variable transmission for the vehicle by controlling supply and discharge of hydraulic fluid to and from the actuator; an upshift solenoid valve that controls the shift control valve to the upshift side; A downshift solenoid valve for controlling the control valve to the downshift side, and a shift control means for controlling the actual gear ratio of the continuously variable transmission for the vehicle to a preset target gear ratio. A control device for a continuously variable transmission for a vehicle,
When the continuously variable transmission is switched from upshift to downshift, a predetermined hydraulic pressure output from an output port of a control valve configured to be able to communicate with the drive-side hydraulic actuator via the shift control valve; Differential pressure calculating means for calculating the differential pressure with the hydraulic pressure of the drive side hydraulic actuator;
And a down-switching control means for controlling the upshift solenoid valve and the downshift solenoid valve based on the differential pressure. A deviation offset of the gear ratio with respect to the target gear ratio is set based on the differential pressure,
The vehicle control unit according to claim 1, wherein the down-switching control means changes a downshift output line serving as a boundary line from which a downshift command is output from a downshift solenoid valve according to the deviation offset. Control device for step transmission. The down-switching control means calculates a deviation between an actual gear ratio and a target gear ratio, and outputs a preset upshift command by the upshift solenoid valve when the deviation is smaller than the deviation offset. And
3. The control device for a continuously variable transmission for a vehicle according to claim 2, wherein when the deviation is larger than the deviation offset, a downshift command is output from the dow shift solenoid valve according to the deviation. A continuously variable transmission for a vehicle having a driving pulley and a driven pulley and a belt wound around both pulleys, a driving hydraulic actuator for changing a groove width of the driving pulley, and a driving hydraulic pressure thereof A shift control valve that changes the pulley position of the drive pulley by controlling the supply and discharge of hydraulic oil to and from the actuator, an upshift solenoid valve that controls the shift control valve to the upshift side, and the shift control valve A vehicle shifter comprising: a solenoid valve for downshift that is controlled to the downshift side; and a shift control unit that controls the actual pulley position of the continuously variable transmission for the vehicle to be a preset target pulley position. A control device for a step transmission,
When the continuously variable transmission is switched from upshift to downshift, a predetermined hydraulic pressure output from an output port of a control valve configured to be able to communicate with the drive-side hydraulic actuator via the shift control valve; Differential pressure calculating means for calculating the differential pressure with the hydraulic pressure of the drive side hydraulic actuator;
And a down-switching control means for controlling the upshift solenoid valve and the downshift solenoid valve based on the differential pressure. A deviation offset of the pulley position with respect to the target pulley position is set based on the differential pressure,
The vehicle control unit according to claim 4, wherein the down-switching control means changes a downshift output line serving as a boundary line from which a downshift command is output from a downshift solenoid valve according to the deviation offset. Control device for step transmission. The down-switching control means calculates a deviation between an actual pulley position and a target pulley position, and outputs a preset upshift command by the upshift solenoid valve when the deviation is smaller than the deviation offset. And
6. The control device for a continuously variable transmission for a vehicle according to claim 5, wherein when the deviation is larger than the deviation offset, a downshift command is output from the dow shift solenoid valve in accordance with the deviation.   When the deviation is smaller than the deviation offset, the down-switching control means reduces the output of the upshift solenoid valve and causes the drive side hydraulic actuator and the output port of the control valve to communicate with each other. The control device for a continuously variable transmission for a vehicle according to claim 3 or 6.

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