JP2009068523A - Oil pressure control apparatus - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To surely put a lockup clutch in a released state, when engaging a frictional engaging element for travel, with a simple structure, in an oil pressure control apparatus. <P>SOLUTION: This oil pressure control apparatus has a belt-driven continuously variable transmission 40, the lockup clutch 26, an advancing clutch C1 as the frictional engaging element for travel, a primary regulator valve 110 that regulates a line oil pressure PL that becomes a source pressure of oil pressure of various parts, and a lockup control valve 140 switched when controlling engagement/release of the lockup clutch 26. Control of the primary regulator valve 110 and control of the lockup control valve 140 are performed by a linear solenoid valve SLT. When engaging the advancing clutch C1, a control oil pressure PSLT of the linear solenoid valve SLT is supplied to the opposite side of the supply side when controlling the lockup control valve 140. <P>COPYRIGHT: (C)2009,JPO&INPIT

Description

  The present invention relates to a hydraulic control device for a vehicle power transmission device.

  As a power transmission device mounted on a vehicle, a belt-type continuously variable transmission that changes the gear ratio by changing the belt engagement diameter while transmitting the power by clamping the belt with hydraulic pressure, and a power transmission path when the vehicle travels A hydraulic traveling friction engagement element (for example, a forward clutch) that is engaged to establish a hydraulic power transmission and a fluid power transmission device provided in the power transmission path, and a belt type continuously variable One having a hydraulic lockup clutch that directly connects to the transmission side is known.

  Such a hydraulic control device for a vehicle power transmission device is provided with a number of various control valves and electromagnetic valves for controlling the control valves. For example, a line oil pressure control valve that adjusts the line oil pressure that is the source pressure of each part, and a line oil pressure that controls the gear ratio of the belt type continuously variable transmission by adjusting the line oil pressure that is the source pressure. A transmission hydraulic pressure control valve supplied to the drive side pulley (primary pulley) of the continuously variable transmission, a belt hydraulic pressure that controls the belt clamping pressure of the belt type continuously variable transmission by regulating the line hydraulic pressure that is also the original pressure. The clamping hydraulic pressure control valve to be supplied to the driven pulley (secondary pulley) of the variable continuously variable transmission, the engagement hydraulic pressure to be supplied when the traveling friction engagement element is engaged, There are provided a garage control valve that can be switched to, a lockup control valve that can be switched during engagement / release control of the lockup clutch, and the like. In addition, a linear solenoid valve, an ON-OFF solenoid valve, and the like for controlling each of these control valves are provided.

Patent Document 1 discloses a technique in which only a line hydraulic control valve is controlled by one linear electromagnetic valve. Conventionally, the control hydraulic pressure of one linear solenoid valve is supplied to the line hydraulic control valve and the lockup control valve, and the line hydraulic control valve and the lockup control valve are controlled by one linear solenoid valve. Technology has also been proposed.
JP 2000-130574 A

  By the way, in the hydraulic control apparatus as described above, in order to avoid the occurrence of engine stall, the lockup clutch is forcibly released when the forward clutch is engaged. For this reason, conventionally, when the forward clutch is engaged, the control hydraulic pressure of the ON-OFF electromagnetic valve that controls the garage control valve is supplied to the lockup control valve. In this case, the lockup control valve is supplied with the control oil pressure of its ON-OFF solenoid valve and the control oil pressure of the linear solenoid valve that controls the engagement / release of the lockup clutch to the opposite hydraulic chambers. The

  However, the control oil pressure of the ON-OFF solenoid valve is set to be relatively low. On the other hand, as described above, in the configuration in which the control of the line hydraulic control valve and the lockup control valve is controlled by one linear electromagnetic valve, when the line hydraulic pressure is controlled to be high, the control hydraulic pressure of the linear electromagnetic valve is increased. There is a need. In this case, the control hydraulic pressure of the linear solenoid valve supplied to the lockup control valve may also be increased. For this reason, in order to put the lock-up clutch into a released state, the area of action of the control oil pressure of the ON-OFF solenoid valve of the lock-up control valve must be larger than the area of action of the control oil pressure of the linear solenoid valve. Therefore, the lock-up control valve needs to be a three-stage valve, which complicates the structure of the valve body and increases the manufacturing cost.

  The present invention has been made in view of such a problem, and has a simple structure and reliably releases a lock-up clutch when engaging a frictional engagement element for traveling such as a forward clutch. An object of the present invention is to provide a hydraulic control device that can be in a state.

  In the present invention, means for solving the above-described problems are configured as follows. That is, the present invention is a hydraulic control device, which is a belt-type continuously variable transmission for transmitting power by clamping a belt by hydraulic pressure and changing a gear ratio by changing a belt engagement diameter, a power source, and the belt A hydraulic lock-up clutch that is directly connected between the power source side and the belt-type continuously variable transmission side, and is provided for driving the vehicle. Hydraulic travel friction engagement elements engaged to establish a transmission path, line oil pressure control valve for adjusting the line oil pressure that is the source pressure of each part, and engagement / release of the lockup clutch A lock-up control valve that is switched at the time of control, and a garage control valve that is capable of switching the engagement hydraulic pressure supplied at the time of engagement of the traveling friction engagement element between the engagement transient hydraulic pressure and the engagement holding hydraulic pressure. prepare for That. When the garage control valve is switched so as to supply the engagement transient oil pressure to the traveling friction engagement element, the control oil pressure of the electromagnetic valve is supplied when the lockup control valve is controlled. It is characterized by being supplied to the opposite side. As the electromagnetic valve, for example, a linear type electromagnetic valve, a duty type electromagnetic valve, or the like can be used.

  According to the above configuration, for example, when the garage control valve is switched to supply the engagement transient hydraulic pressure to the traveling friction engagement element when the vehicle starts, the control hydraulic pressure of the electromagnetic valve is locked up. It is supplied to the side opposite to the side supplied during control of the control valve. That is, the control hydraulic pressure of the electromagnetic valve is supplied to the opposite hydraulic chambers (the hydraulic chambers at both ends of the spool). As a result, the lockup clutch is forcibly released. Such a forced OFF of the lockup clutch can prevent the engine stall. In addition, since the hydraulic pressure supplied to the hydraulic chambers at both ends of the spool is the same, the lockup clutch can be reliably returned to the released state regardless of the magnitude of the hydraulic pressure.

  Further, since the hydraulic pressure supplied to the hydraulic chambers at both ends of the spool is the same, it is possible to make the operating area of each hydraulic pressure to the spool the same. Therefore, the lockup control valve does not need to be a three-stage valve, and the structure of the valve body is simplified. In addition, the degree of freedom of the valve diameter is increased. As a result, an increase in manufacturing cost can be avoided.

  Here, it is preferable that the line hydraulic control valve and the lockup control valve are controlled by one electromagnetic valve.

  Further, it is preferable that when the garage control valve is switched to supply the engagement transient hydraulic pressure to the traveling friction engagement element, the engagement transient hydraulic pressure is controlled by the electromagnetic valve.

  Further, when the lockup control valve is controlled by the electromagnetic valve, it is preferable that the line hydraulic control valve is controlled by an electromagnetic valve other than the electromagnetic valve. In this case, as the other solenoid valve, for example, a clamping hydraulic pressure control valve that supplies a clamping hydraulic pressure for controlling the belt clamping pressure of the belt type continuously variable transmission to the driven pulley of the belt type continuously variable transmission. An electromagnetic valve that controls By doing so, it is possible to avoid a situation in which the line oil pressure regulation control is not performed when the lockup control valve is controlled by the electromagnetic valve.

  When the engine restarts after an engine stall has occurred, the garage control valve is switched to supply an engagement transient hydraulic pressure to the traveling friction engagement element, and the control hydraulic pressure of the electromagnetic valve is It is preferable to continue the control such that the supply is performed on the side opposite to the supply side during the control of the lockup control valve. In this way, when the vehicle restarts after the engine stall, the forced OFF control for forcibly releasing the lockup clutch is continued after the vehicle restarts. Therefore, it is possible to avoid the engine stall from occurring again even if the solenoid valve that performs the control of the lockup control valve is turned on.

  According to the present invention, in the hydraulic control device, the lock-up clutch can be surely released when engaging the frictional engagement element for traveling such as the forward clutch with a simple structure, The engine stall can be prevented.

  The best mode for carrying out the present invention will be described with reference to the accompanying drawings.

  FIG. 1 is a diagram illustrating a schematic configuration of a vehicle drive device according to the embodiment.

  The vehicle drive device illustrated in FIG. 1 is suitably employed in an FF (front engine / front drive) type vehicle. The vehicle drive device includes an engine (internal combustion engine) 10 that is a driving power source, a torque converter 20, a forward / reverse switching device 30, a belt-type continuously variable transmission (CVT) 40, a reduction gear device 50, and a differential. A gear device 60 is provided. In this vehicle drive device, the output of the engine 10 is transmitted from the torque converter 20 to the differential gear device 60 via the forward / reverse switching device 30, the belt-type continuously variable transmission 40, and the reduction gear device 50. Are distributed to the drive wheels 70L and 70R. The torque converter 20, the forward / reverse switching device 30, the belt type continuously variable transmission 40, and the like constitute a power transmission mechanism.

  The torque converter 20 is a fluid transmission device that transmits power via fluid (fluid), and includes a pump impeller 22 that is provided integrally with a front cover 21 to which the output shaft 11 of the engine 10 is coupled, and the pump impeller. And a turbine runner 23 provided adjacent to the inner surface of the front cover 21 and connected to the forward / reverse switching device 30 via the turbine shaft 28. Specifically, the pump impeller 22 and the turbine runner 23 are provided with a large number of blades (not shown), and the pump impeller 22 rotates to generate a fluid spiral flow. By sending it to the runner 23, the turbine runner 23 is rotated by applying torque.

  A stator 24 that changes the flow direction of the fluid sent out from the turbine runner 23 and flows into the pump impeller 22 is disposed in the inner peripheral portion of the pump impeller 22 and the turbine runner 23. The stator 24 is connected to a predetermined fixed portion via a one-way clutch 25. Further, the pump impeller 22 is a mechanical oil pump (hydraulic pressure generation) that is generated when the engine 10 is rotationally driven to supply hydraulic oil to each part of the hydraulic control circuit 100 (see FIG. 3). Source) 27 is provided.

  The torque converter 20 includes a lockup clutch 26. The lock-up clutch 26 is arranged in parallel to the substantial torque converter including the pump impeller 22, the turbine runner 23, and the stator 24, and faces the inner surface of the front cover 21 in the turbine runner 23. Is held in. The lock-up clutch 26 is pressed against the inner surface of the front cover 21 by hydraulic pressure, so that torque is directly transmitted from the front cover 21 as an input member to the turbine runner 23 as an output member. Here, the clutch capacity of the lockup clutch 26 can be controlled by controlling the hydraulic pressure. Specifically, the lockup clutch 26 includes a lockup engagement hydraulic pressure PON and a release hydraulic chamber 262 supplied to the engagement hydraulic chamber 261 by the lockup control valve 140 of the hydraulic control circuit 100 (see FIG. 3). By controlling the differential pressure (lock-up differential pressure) ΔP from the lock-up releasing hydraulic pressure POFF supplied to the full-engaged, half-engaged (engaged in the slip state) or released.

  By completely engaging the lockup clutch 26, the front cover 21 (pump impeller 22) and the turbine runner 23 rotate integrally. Further, by engaging the lockup clutch 26 in a predetermined slip state (half-engaged state), the turbine runner 23 rotates following the pump impeller 22 with a predetermined slip amount during driving. On the other hand, the lockup clutch 26 is released by setting the lockup differential pressure ΔP to be negative. Engagement / release of the lockup clutch 26 by the hydraulic control circuit 100 will be described later.

  The forward / reverse switching device 30 includes a double pinion type planetary gear mechanism 31, a forward clutch C1, and a reverse brake B1.

  The sun gear 32 of the planetary gear mechanism 31 is integrally connected to the turbine shaft 28 of the torque converter 20, and the carrier 36 is integrally connected to the input shaft 47 of the belt type continuously variable transmission 40. The carrier 36 and the sun gear 32 are selectively connected via the forward clutch C1. Further, the ring gear 33 is selectively fixed to the housing via the reverse brake B1.

  An inner pinion gear 34 that meshes with the sun gear 32 and an outer pinion gear 35 that meshes with the inner pinion gear 34 and the ring gear 33 are disposed between the sun gear 32 and the ring gear 33. These pinion gears 34 and 35 are held by a carrier 36 so as to rotate and revolve freely.

  Both the forward clutch C1 and the reverse brake B1 are hydraulic travel friction engagement elements that are engaged and released by a hydraulic actuator. When the forward clutch C1 is engaged and the reverse brake B1 is released, the forward / reverse switching device 30 enters an integral rotation state, and the forward / reverse switching device 30 establishes a forward power transmission path. In this state, the driving force in the forward direction is transmitted to the belt type continuously variable transmission 40 side. On the other hand, a reverse power transmission path is established in the forward / reverse switching device 30 by engaging the reverse brake B1 and releasing the forward clutch C1. In this state, the input shaft 47 rotates in the reverse direction with respect to the turbine shaft 28, and the driving force in the reverse direction is transmitted to the belt type continuously variable transmission 40 side. Further, when both the forward clutch C1 and the reverse brake B1 are released, the forward / reverse switching device 30 is set to a neutral (interrupted state) that interrupts power transmission between the engine 10 and the belt-type continuously variable transmission 40. Become.

  More specifically, the forward clutch C1 and the reverse brake B1 are mechanically switched by the manual valve 170 of the hydraulic control circuit 100 (see FIG. 3) according to the operation of the shift lever 87 (see FIG. 2). Engaged / released. The shift lever 87 is, for example, disposed beside the driver's seat and is switched by the driver. The shift lever 87 is a parking position “P” for parking, a reverse position “R” for reverse travel, and power transmission. Each shift position such as a neutral position “N” for blocking and a drive position “D” for forward traveling is selectively operated. In the parking position “P” and the neutral position “N”, both the forward clutch C1 and the reverse brake B1 are released. In the reverse position “R”, the reverse brake B1 is engaged, while the forward clutch C1 is released. In the drive position “D”, the forward clutch C1 is engaged, while the reverse brake B1 is released. Engagement / release of the travel friction engagement elements (forward clutch C1 and reverse brake B1) of the forward / reverse switching device 30 by the hydraulic control circuit 100 will be described later.

  The belt-type continuously variable transmission 40 transmits power by clamping the transmission belt 45 with hydraulic pressure, and changes the gear ratio of the transmission belt 45 by changing the engagement diameter of the transmission belt 45. The belt-type continuously variable transmission 40 is wound around a driving pulley (primary pulley) 41 provided on the input shaft 47, a driven pulley 42 provided on the output shaft 48, and both the pulleys 41 and 42. The metal transmission belt 45 is provided. The belt type continuously variable transmission 40 is configured such that power transmission is performed via a frictional force between the pulleys 41 and 42 and the transmission belt 45.

  The driving pulley 41 is a variable pulley having a variable effective diameter, and a fixed sheave 411 fixed to the input shaft 47 and a movable sheave arranged on the input shaft 47 in a state in which only the axial direction can slide. 412. Similarly, the driven pulley 42 is a variable pulley having a variable effective diameter, and is provided with a fixed sheave 421 fixed to the output shaft 48 and a state in which the output shaft 48 can slide only in the axial direction. The movable sheave 422 is configured. A hydraulic actuator 413 for changing the V groove width between the fixed sheave 411 and the movable sheave 412 is disposed on the drive sheave 41 side of the movable sheave 412. Similarly, a hydraulic actuator 423 for changing the V groove width between the fixed sheave 421 and the movable sheave 422 is also arranged on the movable sheave 422 side of the driven pulley 42.

  In the belt-type continuously variable transmission 40, by controlling the hydraulic pressure (shift hydraulic pressure) PIN of the hydraulic actuator 413 of the driving pulley 41, the V groove widths of the pulleys 41 and 42 change, and the transmission belt 45 is engaged. The diameter (effective diameter) is changed, and the gear ratio γ (= input shaft rotational speed NIN / output shaft rotational speed NOUT) continuously changes. Further, the hydraulic pressure (clamping hydraulic pressure) POUT of the hydraulic actuator 423 of the driven pulley 42 is controlled so as to generate a predetermined belt clamping pressure (friction force) that transmits the transmission torque within a range where the transmission belt 45 does not slip. Is done.

  The transmission hydraulic pressure PIN of the hydraulic actuator 413 of the driving pulley 41 and the clamping hydraulic pressure POUT of the hydraulic actuator 423 of the driven pulley 42 are adjusted according to a command from the electronic control unit 80 (see FIG. 2). Here, the transmission hydraulic pressure PIN is pressure-controlled by the transmission hydraulic control valve 120 of the hydraulic control circuit 100 (see FIG. 3). The clamping oil pressure POUT is regulated by the clamping oil pressure control valve 130 of the oil pressure control circuit 100. The adjustment of the transmission hydraulic pressure PIN and the pinching hydraulic pressure POUT of the belt-type continuously variable transmission 40 by the hydraulic control circuit 100 will be described later.

  FIG. 2 is a block diagram showing an example of a control system of the power transmission mechanism of the vehicle drive device described above.

  The electronic control device 80 illustrated in FIG. 2 includes a CPU 801, a ROM 802, a RAM 803, and a backup RAM 804. Then, the CPU 801 performs signal processing in accordance with a program stored in the ROM 802 in advance using the temporary storage function of the RAM 803, thereby adjusting the transmission hydraulic pressure PIN and the clamping hydraulic pressure POUT of the belt-type continuously variable transmission 40. Engagement / release control of the frictional engagement elements for travel (forward clutch C1, reverse brake B1) of the forward / reverse switching device 30, control of engagement / release of the lockup clutch 26 of the torque converter 20, hydraulic pressure of each part Various controls such as pressure regulation control of the line hydraulic pressure PL, which is the original pressure, are executed.

  More specifically, the ROM 802 stores various control programs, maps that are referred to when the various control programs are executed, and the like. The CPU 801 executes arithmetic processing based on various control programs and maps stored in the ROM 802. The RAM 803 is a memory that temporarily stores calculation results in the CPU 801, data input from each sensor, and the like. The backup RAM 804 is a nonvolatile memory that stores data to be saved when the engine 10 is stopped. It is memory. The CPU 801, ROM 802, RAM 803, and backup RAM 804 are connected to each other via a bidirectional bus 807, and are connected to an input interface 805 and an output interface 806.

  Various sensors are connected to the input interface 805 in order to detect the operation state (or running state) of the vehicle on which the vehicle drive device is mounted. Specifically, the input interface 805 includes a lever position sensor 81, an accelerator operation amount sensor 82, an engine rotation speed sensor 83, an output shaft rotation speed sensor 84 that also functions as a vehicle speed sensor, an input shaft rotation speed sensor 85, a turbine rotation speed. A sensor 86 or the like is connected. The lever position sensor 81 detects a plurality of ONs that detect that the shift lever 87 is operated at a shift position such as a parking position “P”, a reverse position “R”, a neutral position “N”, or a drive position “D”. -An OFF switch is provided.

  Then, the electronic control unit 80 receives from these various sensors the lever position (operation position) PSH of the shift lever 87, the operation amount (accelerator operation amount) θACC of an accelerator operation member such as an accelerator pedal, and the rotational speed ( Engine rotational speed NE, rotational speed of output shaft 48 of belt-type continuously variable transmission 40 (output shaft rotational speed) NOUT, rotational speed of input shaft 47 of belt-type continuously variable transmission 40 (input shaft rotational speed) NIN, A signal representing the rotational speed (turbine rotational speed) NT of the turbine shaft 28 of the torque converter 20 is supplied. The turbine rotational speed NT coincides with the input shaft rotational speed NIN during forward traveling when the forward clutch C1 of the forward / reverse switching device 30 is engaged. The output shaft rotational speed NOUT corresponds to the vehicle speed V. Further, the accelerator operation amount θACC represents the driver's requested output amount.

  The output interface 806 is connected to linear solenoid valves SLP, SLS, SLT, ON-OFF solenoid valve SL1 of the hydraulic control circuit 100, and the like. The electronic control unit 80 controls the excitation currents of the linear solenoid valves SLP, SLS, SLT of the hydraulic control circuit 100, and controls the control hydraulic pressures PSLP, PSLS, PSLT output from these linear solenoid valves SLP, SLS, SLT, respectively. While adjusting the pressure, the ON-OFF solenoid valve SL1 of the hydraulic control circuit 100 is switched between an ON state (excitation state) and an OFF state (non-excitation state). Thus, pressure regulation control of the transmission hydraulic pressure PIN and the clamping hydraulic pressure POUT of the belt-type continuously variable transmission 40, engagement / release control of the traveling friction engagement element of the forward / reverse switching device 30, and engagement of the lockup clutch 26 -Release control, pressure regulation control of line oil pressure PL, etc. are performed.

  FIG. 3 is a circuit diagram showing an example of a hydraulic control circuit for controlling the power transmission mechanism of the vehicle drive device described above.

  The hydraulic control circuit 100 illustrated in FIG. 3 includes the oil pump 27, the transmission hydraulic control valve 120, the clamping hydraulic control valve 130, the lockup control valve 140, and the manual valve 170 described above, and further includes a primary regulator valve 110, a garage. A shift valve 160 and a pressure reducing valve 180 are included. Further, the hydraulic control circuit 100 includes linear solenoid valves SLP, SLS, SLT, and ON-OFF solenoid valve SL1 connected to the electronic control device 80 described above. The hydraulic control circuit 100 shown in FIG. 3 schematically shows a part of the hydraulic control circuit of the power transmission mechanism of the vehicle drive device, and the actual hydraulic circuit is other than that shown in FIG. It includes valves and oil passages not shown.

  In the hydraulic control circuit 100, the hydraulic pressure generated by the oil pump 27 is adjusted by the primary regulator valve 110 to the line hydraulic pressure PL that becomes the source pressure of each part. The line oil pressure PL regulated by the primary regulator valve 110 is supplied to each part of the hydraulic control circuit 100 such as the transmission hydraulic pressure control valve 120 and the clamping hydraulic pressure control valve 130 via the oil passage 101.

  The primary regulator valve 110 includes a first spool 111a and a second spool 111b that are movable in the axial direction, and a spring 112 as a biasing means that biases the first spool 111a and the second spool 111b to one side. . In FIG. 3, the first spool 111a provided on the upper side and the second spool 111b provided on the lower side are both slidable up and down. The primary regulator valve 110 is provided with control ports 115a, 115b, and 115c, an input port 116, and an output port 117.

  The input port 116 and the output port 117 are communicated and blocked by the first spool 111a. The spring 112 is disposed in a compressed state in a control hydraulic chamber 113c provided on one end side (the lower end side in FIG. 3) of the second spool 111b. That is, the control hydraulic chamber 113c is a spring chamber in which the spring 112 is disposed. Due to the urging force of the spring 112, the second spool 111b and the first spool 111a are pressed in a direction to shut off the input port 116 and the output port 117 (upward in FIG. 3).

  The control port 115a is connected to a control hydraulic chamber 113a provided on the other end side (the upper end side in FIG. 3) of the first spool 111a. Further, the control port 115 a is connected to the oil passage 101. The line oil pressure PL is supplied to the control oil pressure chamber 113a via the control port 115a.

  The control port 115b is connected to a control hydraulic chamber 113b provided between one end side of the first spool 111a and the other end side of the second spool 111b. The control port 115b is connected to the output port SLSb of the linear solenoid valve SLS via the oil passage 102. The output hydraulic pressure (control hydraulic pressure) PSLS of the linear solenoid valve SLS is supplied to the control hydraulic pressure chamber 113b via the control port 115b.

  The control port 115c is connected to the control hydraulic chamber 113c. Further, the control port 115 c is connected to the output port 187 of the pressure reducing valve 180 through the oil passage 103. The output hydraulic pressure PCTL of the pressure reducing valve 180 is supplied to the control hydraulic pressure chamber 113c through the control port 115c.

  The input port 116 is connected to the oil passage 101. The line hydraulic pressure PL is input via the input port 116. The output port 117 is connected to a secondary regulator valve (not shown).

  The first spool 111a has a force generated by the line hydraulic pressure PL introduced into the control hydraulic chamber 113a, a force generated by the control hydraulic pressure PSLS introduced into the control hydraulic chamber 113b, or an output of the pressure reducing valve 180 introduced into the control hydraulic chamber 113c. It slides up and down due to the balance between the force of the hydraulic PCTL and the combined force of the urging force of the spring 112. The input port 116 and the output port 117 are cut off while the force due to the control oil pressure is greater than the force due to the line oil pressure PL. On the other hand, when the force due to the line oil pressure PL exceeds the force due to the control oil pressure or the like, the first spool 111a moves downward in FIG. 3, and the input port 116 and the output port 117 are brought into communication. As a result, the hydraulic pressure from the oil passage 101 is drained via the output port 117, so that the line hydraulic pressure PL is adjusted. Therefore, the control of the line oil pressure PL is controlled by controlling at least one of the control oil pressure PSLS of the linear solenoid valve SLS and the output oil pressure PCTL of the pressure reducing valve 180 (in other words, the control oil pressure PSLT of the linear solenoid valve SLT). It is possible to do.

  Here, the first spool 111a and the second spool 111b are formed to have the same diameter. Then, the operating area (pressure receiving area) of the control hydraulic pressure PSLS supplied to the first spool 111a via the control port 115b, the operating area (pressure receiving area) of the control hydraulic pressure PSLS to the second spool 111b, and control. The operating area (pressure receiving area) of the output hydraulic pressure PCTL supplied through the port 115c to the second spool 111b is the same.

  Then, the force (control oil pressure PSLS * (pressure receiving area)) introduced into the control oil pressure chamber 113b and the force (output oil pressure PCTL * (output oil pressure PCTL * (pressure receiving area)) and the force caused by the output oil pressure PCTL of the pressure reducing valve 180 introduced into the control oil pressure chamber 113c. The higher hydraulic pressure of the combined force of the pressure receiving area)) and the biasing force of the spring 112 is selected. Specifically, when the force by the control oil pressure PSLS is higher than the combined force of the force by the output oil pressure PCTL and the urging force of the spring 112, the primary regulator valve 110 has a higher capacity than the first spool 111a with respect to the second spool 111b. While moving up and down in a separated state, when the combined force of the force by the output oil pressure PCTL and the biasing force of the spring 112 is higher than the force by the control oil pressure PSLS, both the spools 111a and 111b come into contact with each other. It is configured to move up and down in a stable state. As described above, when controlling the line oil pressure PL, the oil pressure used for controlling the line oil pressure PL is automatically selected without calculating the two control oil pressures PSLS and PSLT. Control can be easily performed.

  The transmission hydraulic pressure control valve 120 includes a spool 121 that is movable in the axial direction, and a spring 122 as a biasing means that biases the spool 121 in one direction. The transmission hydraulic pressure control valve 120 is configured to continuously control the line hydraulic pressure PL, which is the original pressure, using the output hydraulic pressure (control hydraulic pressure) PSLP of the linear solenoid valve SLP as a pilot pressure. The hydraulic pressure (shift hydraulic pressure PIN) adjusted by the shift hydraulic pressure control valve 120 is supplied to the hydraulic actuator 413 of the drive pulley 41 via the oil passage 109a.

  Therefore, the pressure adjustment control of the transmission hydraulic pressure PIN is performed by controlling the control hydraulic pressure PSLP of the linear solenoid valve SLP. Since the control hydraulic pressure PSLP changes linearly according to the excitation current, the gear ratio γ of the belt type continuously variable transmission 40 is continuously changed according to the control hydraulic pressure PSLP. In this case, for example, the target input shaft rotational speed set based on the vehicle state indicated by the actual vehicle speed V and the accelerator opening θACC from the shift map stored in advance in the ROM 802 and the actual input shaft rotational speed NIN are obtained. The gear ratio γ of the belt-type continuously variable transmission 40 is changed in accordance with the rotational speed difference (deviation). The shift map corresponds to a shift condition, and is, for example, the relationship between the vehicle speed V and the target input shaft rotation speed that is the target input rotation speed of the belt-type continuously variable transmission 40 using the accelerator opening θACC as a parameter.

  The clamping hydraulic pressure control valve 130 includes a spool 131 that is movable in the axial direction, and a spring 132 as a biasing means that biases the spool 131 in one direction. The clamping oil pressure control valve 130 is configured to continuously control the line oil pressure PL, which is the original pressure, using the control oil pressure PSLS of the linear solenoid valve SLS as a pilot pressure. The hydraulic pressure (clamping hydraulic pressure POUT) adjusted by the clamping hydraulic pressure control valve 130 is supplied to the hydraulic actuator 423 of the driven pulley 42 via the oil passage 109b.

  Therefore, the pressure regulation control of the clamping oil pressure POUT is performed by controlling the control oil pressure PSLS of the linear solenoid valve SLS. Since the control hydraulic pressure PSLS changes linearly according to the excitation current, the belt clamping pressure of the belt type continuously variable transmission 40 is continuously changed according to the control hydraulic pressure PSLS. In this case, for example, the driven pulley 42 is used to obtain a necessary target transmission hydraulic pressure that is set based on the vehicle state indicated by the actual transmission ratio γ and the accelerator opening θACC from the clamping pressure map stored in advance in the ROM 802. The clamping pressure POUT of the hydraulic actuator 423 is regulated, and the belt clamping pressure of the belt type continuously variable transmission 40 is changed according to the clamping pressure POUT. The sandwiching pressure map is a relationship between the gear ratio γ and the required target transmission hydraulic pressure with the accelerator opening θACC as a parameter, and is a relationship that is experimentally obtained in advance so that belt slip does not occur.

  Here, since the transmission hydraulic pressure PIN and the pinching hydraulic pressure POUT are obtained by adjusting the line hydraulic pressure PL as the original pressure, the line hydraulic pressure PL needs to be at least equal to or greater than the transmission hydraulic pressure PIN and the clamping hydraulic pressure POUT. For this reason, the oil pump 27 is set so that the target transmission hydraulic pressure and the line hydraulic pressure PL equal to or higher than the target clamping hydraulic pressure required to control the transmission gear ratio γ and the clamping pressure of the belt-type continuously variable transmission 40 are obtained. Need to drive. In this case, the required target transmission hydraulic pressure and target clamping hydraulic pressure are set as shown in FIG. 6, for example. FIG. 6 shows the set values of the target transmission hydraulic pressure and the target clamping hydraulic pressure that are required according to the transmission gear ratio γ of the belt-type continuously variable transmission 40 under the condition that the input shaft rotational speed NIN and the input torque are constant. An example of the change is shown. The broken line in the figure indicates the change in the target transmission hydraulic pressure, and the alternate long and short dash line indicates the change in the target clamping hydraulic pressure.

  On the speed increasing side (left side in the figure) where the speed ratio γ is lower than γ1, the target transmission hydraulic pressure is set higher than the target clamping hydraulic pressure, and the difference increases as the acceleration increases. On the other hand, on the deceleration side (right side in the figure) where the gear ratio γ is higher than γ1, the target clamping hydraulic pressure is set higher than the target transmission hydraulic pressure, and the difference increases as the deceleration increases. That is, the set values of the target transmission hydraulic pressure and the target clamping hydraulic pressure reverse with the change of the transmission gear ratio γ (using the transmission gear ratio γ1 as a switching point). In order to suppress the drive loss of the oil pump 27, when the speed ratio γ is higher than γ1, it is preferable to set the line oil pressure PL to be equal to or slightly higher than the target clamping oil pressure. When γ is lower than γ1, it is preferable to set the line oil pressure PL equal to or slightly higher than the target transmission oil pressure. A specific value of the gear ratio γ1 is a gear ratio “1”, but is not limited to this because it depends on the area of the hydraulic actuators 413 and 423.

  The pressure reducing valve 180 is a pressure regulating valve that adjusts the hydraulic pressure introduced into the control hydraulic chamber 113 c of the primary regulator valve 110. The pressure reducing valve 180 is provided between the primary regulator valve 110 and the linear solenoid valve SLT. Specifically, the pressure reducing valve 180 includes a spool 181 that is movable in the axial direction, and a spring 182 as a biasing means that biases the spool 181 in one direction. In FIG. 3, a spool 181 is provided so as to be slidable up and down. The pressure reducing valve 180 is provided with a control port 185, an input port 186, an output port 187, and a feedback port 188.

  The spool 181 communicates and blocks the input port 186 and the output port 187. The spring 182 is disposed in a compressed state in a spring chamber 184 provided on one end side (the lower end side in FIG. 3) of the spool 181. Due to the biasing force of the spring 182, the spool 181 is pressed in a direction (upward in FIG. 3) in which the input port 186 and the output port 187 are blocked.

  The control port 185 is connected to a control hydraulic chamber 183 provided on the other end side (the upper end side in FIG. 3) of the spool 181. The control port 185 is connected to the output port SLTb of the linear solenoid valve SLT via the oil passage 104. Via this control port 185, the output hydraulic pressure (control hydraulic pressure) PSLT of the linear solenoid valve SLT is supplied to the control hydraulic pressure chamber 183.

  The input port 186 is connected to a first modulator valve (not shown). Via this input port 186, the first modulator oil pressure PM1 regulated by the first modulator valve using the line oil pressure PL as the original pressure is inputted. The output port 187 is connected to the control port 115 c of the primary regulator valve 110 described above via the oil passage 103. The hydraulic pressure (output hydraulic pressure PCTL) reduced by the pressure reducing valve 180 is output from the output port 187.

  The feedback port 188 is connected to the spring chamber 184. The feedback port 188 is connected to the oil passage 103. A hydraulic pressure equal to the output hydraulic pressure PCTL is supplied to the spring chamber 184 via the feedback port 188.

  The spool 181 slides up and down by a balance between the control hydraulic pressure PSLT introduced into the control hydraulic chamber 183 and the combined force of the output hydraulic pressure PCTL introduced into the spring chamber 184 and the urging force of the spring 182. In this case, while the urging force of the spring 182 exceeds the force by the control hydraulic pressure PSLT, the spool 181 is fixed to the other end side, and the input port 186 and the output port 187 are blocked. Yes. In this state, the output hydraulic pressure PCTL is not introduced into the spring chamber 184 and the control hydraulic chamber 113c of the primary regulator valve 110.

  On the other hand, when the force by the control hydraulic pressure PSLT exceeds the urging force of the spring 182, movement of the spool 181 to one end side is allowed. Accordingly, the input port 186 and the output port 187 communicate with each other, so that the output hydraulic pressure PCTL is introduced into the spring chamber 184 and the control hydraulic chamber 113c of the primary regulator valve 110 via the oil passage 103. become.

  The output hydraulic pressure PCTL output from the output port 187 has an urging force (load) of the spring 182 as W1, and an action area (pressure receiving area) of the output hydraulic pressure PCTL supplied to the spool 181 via the feedback port 188 as S1. Then, [PSLT-W1 / S1] is obtained. Therefore, as shown in FIG. 4, by controlling the control oil pressure PSLT of the linear solenoid valve SLT, the pressure adjustment control of the output oil pressure PCTL of the pressure reducing valve 180 can be performed.

  Thus, the output hydraulic pressure PCTL is reduced by the pressure reducing valve 180 by the hydraulic pressure corresponding to the urging force of the spring 182 with respect to the control hydraulic pressure PSLT of the linear solenoid valve SLT. As a result, the linear solenoid valve SLT can be used not only for pressure regulation control of the line hydraulic pressure PL but also for engagement / release control of the lockup clutch 26, as will be described later.

  The manual valve 170 is a switching valve that switches hydraulic pressure supply to the forward clutch C1 and the reverse brake B1 of the forward / reverse switching device 30 in accordance with the operation of the shift lever 87. The manual valve 170 is switched corresponding to each shift position such as the parking position “P”, the reverse position “R”, the neutral position “N”, the drive position “D”, etc. of the shift lever 87.

  When the manual valve 170 is switched corresponding to the parking position “P” and the neutral position “N” of the shift lever 87, the hydraulic pressure is supplied to the hydraulic servo of the forward clutch C1 and the hydraulic servo of the reverse brake B1. Not. The hydraulic fluid of the hydraulic servo of the forward clutch C1 and the hydraulic servo of the reverse brake B1 is drained through the manual valve 170. As a result, both the forward clutch C1 and the reverse brake B1 are released.

  When the manual valve 170 is switched corresponding to the reverse position “R” of the shift lever 87, the hydraulic pressure is supplied to the hydraulic servo of the reverse brake B1, while the hydraulic pressure of the forward clutch C1 is hydraulic. Is not supplied. The hydraulic fluid of the hydraulic servo of the forward clutch C1 is drained through the manual valve 170. As a result, the reverse brake B1 is engaged and the forward clutch C1 is released.

  When the manual valve 170 is switched corresponding to the drive position “D” of the shift lever 87, the input port 176 and the output port 177 are communicated to supply hydraulic pressure to the hydraulic servo of the forward clutch C1. On the other hand, no hydraulic pressure is supplied to the hydraulic servo of the reverse brake B1. The hydraulic fluid of the hydraulic servo of the reverse brake B1 is drained through the manual valve 170. As a result, the forward clutch C1 is engaged and the reverse brake B1 is released. The hydraulic pressure supply accompanying the engagement of the forward clutch C1 is performed through a garage shift valve 160 described below.

  When the garage shift is performed, the garage shift valve 160 is used when the travel friction engagement elements (forward clutch C1 and reverse brake B1) of the forward / reverse switching device 30 are engaged and engaged (when fully engaged). This is a switching valve that switches the oil passage in response to. By switching the garage shift valve 160, for example, the shift lever 87 is operated from the non-traveling position such as the parking position “P” or the neutral position “N” to the traveling position such as the drive position “D” when starting the vehicle. At this time, the hydraulic pressure supplied to the hydraulic servo of the forward clutch C1 is switched between the engagement transient hydraulic pressure corresponding to the engagement transition and the engagement holding hydraulic pressure corresponding to the complete engagement. Similarly, when the shift lever 87 is operated to the reverse position “R”, switching of the garage shift valve 160 causes the hydraulic pressure supplied to the hydraulic servo of the reverse brake B1 to correspond to the engagement transient. It is switched between the hydraulic pressure and the engagement holding hydraulic pressure corresponding to the complete engagement. Here, the case where the hydraulic pressure supplied to the hydraulic servo of the forward clutch C1 is switched by the garage shift valve 160 will be described as a representative.

  Specifically, the garage shift valve 160 is switched to the control position shown in the left half of FIG. 3 when the forward clutch C1 is engaged, and to the right half of FIG. 3 when the forward clutch C1 is fully engaged. It is configured to be switched to the normal position shown. Switching of the garage shift valve 160 is performed by controlling the output hydraulic pressure (control hydraulic pressure) of the ON-OFF solenoid valve SL1.

  The garage shift valve 160 includes a spool 161 that is movable in the axial direction, and a spring 162 that serves as a biasing means that biases the spool 161 in one direction. In FIG. 3, a spool 161 is slidable up and down. The spring 162 is disposed in a compressed state in a spring chamber 164 provided on one end side (the lower end side in FIG. 3) of the spool 161. Due to the urging force of the spring 162, the spool 161 is pressed in a direction (upward in FIG. 3) to hold the garage shift valve 160 in the normal position. The garage shift valve 160 is provided with a control port 165, input ports 166a, 166b, 166c, output ports 167a, 167b, and drain ports 169a, 169b.

  The control port 165 is connected to a control hydraulic chamber 163 provided on the other end side (the upper end side in FIG. 3) of the spool 161. The control port 165 is connected to the output port SL1b of the ON-OFF solenoid valve SL1. The control oil pressure of the ON-OFF solenoid valve SL1 is supplied to the control oil pressure chamber 163 via the control port 165.

  The input port 166a is connected to a second modulator valve (not shown). Via this input port 166a, the second modulator oil pressure PM2 regulated by the second modulator valve using the line oil pressure PL as a source pressure is inputted. The input ports 166b and 166c are connected to the output port SLTc of the linear solenoid valve SLT via the oil passage 105, respectively. The control hydraulic pressure PSLT of the linear solenoid valve SLT is input via the input ports 166b and 166c. The second modulator valve is provided on the downstream side of the primary regulator valve 110, and the first modulator valve is provided on the downstream side of the second modulator valve. For this reason, the second modulator oil pressure PM2 is set higher than the first modulator oil pressure PM1.

  The output port 167a is connected to the input port 176 of the manual valve 170 via the oil passage 107. The output port 167 b is connected to the backup port 145 b of the lockup control valve 140 through the oil passage 108. The drain port 169b is connected to the spring chamber 164.

  Subsequently, the switching operation of the garage shift valve 160 will be described.

  In this embodiment, an ON-OFF solenoid valve SL1 is provided as a control valve for switching the garage shift valve 160. The ON-OFF solenoid valve SL1 is configured to switch between an ON state and an OFF state in accordance with a command sent from the electronic control unit 80. As the ON-OFF solenoid valve SL1, a normally closed type solenoid valve as described below can be used, but a configuration using a normally open type solenoid valve may be used. As a control valve for switching the garage shift valve 160, a linear type solenoid valve, a duty type solenoid valve, a three-way valve type solenoid valve, or the like may be used instead of the ON-OFF solenoid valve SL1. Is possible.

  Specifically, when the ON-OFF solenoid valve SL1 is in an ON state when energized, a predetermined control oil pressure is output from the output port SL1b, and the control oil pressure is supplied to the garage shift valve 160. The spool 161 moves downward against the urging force of the spring 162 by the control oil pressure. As a result, the garage shift valve 160 is held in the control position. On the other hand, when the ON-OFF solenoid valve SL1 is in the OFF state, which is a non-energized state, the output of the control hydraulic pressure is stopped. The spool 161 is moved upward by the biasing force of the spring 162. As a result, the garage shift valve 160 is held in the normal position. The first modulator hydraulic pressure PM1 regulated by the first modulator valve is introduced into the ON-OFF solenoid valve SL1 via the input port SL1a.

  The ON-OFF solenoid valve SL1 is fully engaged when the forward clutch C1 of the forward / reverse switching device 30 is engaged, in other words, after the forward clutch C1 is engaged. Until reaching the state, it is controlled to be in the ON state. Accordingly, the control oil pressure of the ON-OFF solenoid valve SL1 is introduced into the control oil pressure chamber 163 via the control port 165, and the garage shift valve 160 is held at the control position. As a result, the input port 166b and the output port 167a communicate with the input port 166c and the output port 167b, respectively.

  In this case, since the input port 176 and the output port 177 of the manual valve 170 are connected, the control hydraulic pressure PSLT of the linear solenoid valve SLT is supplied to the hydraulic servo of the forward clutch C1 by the communication of the input port 166b and the output port 167a. Will come to be. Accordingly, the engagement hydraulic pressure supplied to the hydraulic servo at the time of engagement transition of the forward clutch C1 is the control hydraulic pressure PSLT. Thus, the engagement transient control of the forward clutch C1 is performed by the linear solenoid valve SLT. Here, since the control hydraulic pressure PSLT of the linear solenoid valve SLT as the engagement transient hydraulic pressure changes linearly according to the excitation current (see FIG. 5), the forward clutch C1 can be smoothly engaged during the garage shift. Thus, the shock accompanying the engagement of the forward clutch C1 can be suppressed. Further, the control hydraulic pressure PSLT is supplied to the spring chamber 144a of the lockup control valve 140 through the oil passage 108 by the communication between the input port 166c and the output port 167b.

  On the other hand, the ON-OFF solenoid valve SL1 is controlled to be in an OFF state when the forward clutch C1 is completely engaged (for example, during steady running). In this case, since the supply of the control hydraulic pressure of the ON-OFF solenoid valve SL1 to the control hydraulic chamber 163 is stopped, the garage shift valve 160 is held at the normal position. Thereby, the input port 166a and the output port 167a are communicated. In this case, since the input port 176 and the output port 177 of the manual valve 170 are in communication, the communication of the input port 166a and the output port 167a supplies the second modulator hydraulic pressure PM2 to the hydraulic servo of the forward clutch C1. It becomes like this. Therefore, the engagement holding hydraulic pressure supplied to the hydraulic servo when the forward clutch C1 is completely engaged is the second modulator hydraulic pressure PM2. Here, the second modulator oil pressure PM2 is set to a constant oil pressure (clutch pressure) equal to or higher than the control oil pressure PSLT, and the forward clutch C1 can be reliably held in the fully engaged state.

  In other cases than the above (when the engagement is transitional or not completely engaged), the ON-OFF solenoid valve SL1 is controlled to be in the OFF state, and the garage shift valve 160 is held at the normal position. However, if the manual valve 170 is switched corresponding to a position other than the traveling position such as the drive position “D” of the shift lever 87, the input port 176 and the output port 177 of the manual valve 170 are blocked, and thus The second modulator hydraulic pressure PM2 is not supplied to the hydraulic servo of the forward clutch C1.

  The lock-up control valve 140 controls engagement / release of the lock-up clutch 26. Specifically, the lockup control valve 140 controls the engagement / release of the lockup clutch 26 by controlling the lockup differential pressure ΔP (ΔP = lockup engagement hydraulic pressure PON−lockup release hydraulic pressure POFF). Is configured to do. Control of the lockup differential pressure ΔP by the lockup control valve 140 is performed by controlling the control hydraulic pressure PSLT of the linear solenoid valve SLT.

  The lockup control valve 140 includes a spool 141 that is movable in the axial direction, and a spring 142 that serves as a biasing unit that biases the spool 141 in one direction. In FIG. 3, the spool 141 is slidable up and down. The spring 142 is disposed in a compressed state in a spring chamber 144a provided on one end side (the lower end side in FIG. 3) of the spool 141. Due to the biasing force of the spring 142, the spool 141 is pressed in a direction (upward in FIG. 3) to hold the lockup control valve 140 in the OFF position shown in the left half of FIG. The lockup control valve 140 includes a control port 145a, a backup port 145b, input ports 146a and 146b, a release side port 147a, an engagement side port 147b, a feedback port 148, and drain ports 149a and 149b. Is provided.

  The control port 145a is connected to a control hydraulic chamber 143 provided on the other end side (the upper end side in FIG. 3) of the spool 141. The control port 145a is connected to the output port SLTc of the linear solenoid valve SLT via the oil passage 105. The control hydraulic pressure PSLT of the linear solenoid valve SLT is supplied to the control hydraulic pressure chamber 143 via the control port 145a.

  The backup port 145b is connected to the spring chamber 144a. The backup port 145b is connected to the output port 167b of the garage shift valve 160 via the oil passage 108. When the garage shift valve 160 is held at the control position, the control hydraulic pressure PSLT is supplied to the spring chamber 144a via the backup port 145b.

  The input ports 146a and 146b are respectively connected to a secondary regulator valve (not shown) connected to the output port 117 of the primary regulator valve 110. The secondary hydraulic pressure PSEC regulated by the secondary regulator valve is input via the input ports 146a and 146b.

  The release side port 147a is connected to the release side hydraulic chamber 262 of the lockup clutch 26 through the oil passage 106a. The engagement side port 147b is connected to the engagement side hydraulic chamber 261 of the lockup clutch 26 via the oil passage 106b.

  The feedback port 148 is connected to a feedback hydraulic chamber 144 b provided on one end side of the spool 141. The feedback port 148 is connected to the oil passage 106b. Through this feedback port 148, a hydraulic pressure equal to the lockup engagement hydraulic pressure PON is supplied to the feedback hydraulic chamber 144b. The drain port 149b communicates with a lubrication circuit (not shown).

  Next, the operation of the lockup clutch 26 by the lockup control valve 140 will be described.

  When the control hydraulic pressure PSLT of the linear solenoid valve SLT is introduced into the control hydraulic pressure chamber 143 via the control port 145a, the lock-up control valve 140 causes the spool 141 to resist the biasing force of the spring 142 in accordance with the control hydraulic pressure PSLT. Thus, the state is moved downward (ON state). In this case, the spool 141 moves downward as the control hydraulic pressure PSLT is increased. The right half of FIG. 3 shows a state where the spool 141 has moved downward as much as possible. In the state shown in the right half of FIG. 3, the input port 146b, the engagement side port 147b, the release side port 147a, and the drain port 149a are communicated with each other. At this time, the lock-up clutch 26 is in a completely engaged state.

  When the lockup control valve 140 is in the ON state, the spool 141 is a combined force of the control hydraulic pressure PSLT introduced into the control hydraulic pressure chamber 143 and the lockup release hydraulic pressure POFF acting on the release side port 147a and the input port 146a. And the lockup engagement hydraulic pressure PON introduced into the feedback hydraulic chamber 144b and the combined force of the urging force of the spring 142 slide up and down. Here, the lockup clutch 26 is engaged according to the lockup differential pressure ΔP. The lockup differential pressure ΔP is controlled by controlling the control hydraulic pressure PSLT of the linear solenoid valve SLT. Since the control hydraulic pressure PSLT changes linearly according to the excitation current (see FIG. 5), the lockup differential pressure ΔP can be continuously adjusted. Accordingly, the degree of engagement (clutch capacity) of the lockup clutch 26 can be continuously changed according to the lockup differential pressure ΔP.

  More specifically, as the control hydraulic pressure PSLT is increased, the lockup differential pressure ΔP is increased and the degree of engagement of the lockup clutch 26 is increased. In this case, the hydraulic oil from the secondary regulator valve is supplied to the engagement side hydraulic chamber 261 of the lockup clutch 26 via the input port 146b, the engagement side port 147b, and the oil passage 106b. On the other hand, the hydraulic fluid in the release side hydraulic chamber 262 is discharged through the oil passage 106a, the release side port 147a, and the drain port 149a. When the lockup differential pressure ΔP becomes equal to or greater than a predetermined value, the lockup clutch 26 reaches full engagement.

  Conversely, the lower the control hydraulic pressure PSLT, the smaller the lockup differential pressure ΔP and the smaller the degree of engagement of the lockup clutch 26. In this case, hydraulic oil from the secondary regulator valve is supplied to the release-side hydraulic chamber 262 via the input port 146a, the release-side port 147a, and the oil passage 106a. On the other hand, the hydraulic fluid in the engagement side hydraulic chamber 261 is discharged through the oil passage 106b, the engagement side port 147b, and the drain port 149b. When the lockup differential pressure ΔP becomes a negative value, the lockup clutch 26 is released.

  On the other hand, when the supply of the control hydraulic pressure PSLT of the linear solenoid valve SLT to the control hydraulic chamber 143 is stopped, the lock-up control valve 140 causes the spool 141 to be moved by the biasing force of the spring 142 as shown in the left half of FIG. It moves upward and is held in its original position (OFF state). In this OFF state, the input port 146a, the release side port 147a, the engagement side port 147b, and the drain port 149b are communicated with each other. At this time, the lock-up clutch 26 is in a released state.

  Further, when the garage shift valve 160 is held at the control position, the engagement / release control of the lockup clutch 26 as described above is not performed, and control for forcibly releasing the lockup clutch 26 is performed. Done. This control will be described later.

  The linear solenoid valves SLT, SLP, SLS are, for example, normally open type solenoid valves. That is, at the time of de-energization, the input port and the output port are communicated, and the hydraulic pressure determined by the urging force of the spring provided in the solenoid valve is output from the output port as the control hydraulic pressure. On the other hand, when energized, the hydraulic pressure that is pressure-controlled according to the excitation current determined by the duty signal sent from the electronic control unit 80 is output from the output port as the control hydraulic pressure. In this case, the pressure adjustment control is performed so that the control hydraulic pressure decreases as the excitation current increases. When the exciting current becomes equal to or greater than a predetermined value, the control hydraulic pressure becomes “0” and the output of the control hydraulic pressure is stopped. For example, the control hydraulic pressure PSLT of the linear solenoid valve SLT changes linearly according to the excitation current as shown in FIG. Similarly, the control hydraulic pressures PSLP and PSLS of the linear solenoid valves SLP and SLS also change linearly according to the excitation current. In addition, it is good also as a structure which uses a normally closed type solenoid valve as linear solenoid valve SLT, SLP, SLS.

  The linear solenoid valve SLT performs pressure regulation control of the line hydraulic pressure PL, engagement / release control of the lockup clutch 26, and engagement transient control (control of engagement transient hydraulic pressure) of the forward clutch C1 and the reverse brake B1. Provided to do. In addition, as a control valve for performing these controls, it is good also as a structure which uses a duty type solenoid valve instead of the linear solenoid valve SLT.

  Specifically, the second modulator hydraulic pressure PM2 regulated by the second modulator valve is introduced into the linear solenoid valve SLT via the input port SLTa. When not energized, the hydraulic pressure PSLTmax determined by the urging force of the spring provided in the linear solenoid valve SLT is output as the control hydraulic pressure PSLT. When energized, the second modulator hydraulic pressure PM2 is linearly adjusted according to the excitation current. The pressure-controlled hydraulic pressure is output as the control hydraulic pressure PSLT.

  The control hydraulic pressure PSLT output from the output port SLTb is supplied to the pressure reducing valve 180 via the oil passage 104. The pressure regulation control of the line oil pressure PL is not performed directly based on the control oil pressure PSLT, but based on the output oil pressure PCTL reduced by the pressure reducing valve 180 according to the control oil pressure PSLT. Further, the control hydraulic pressure PSLT output from the output port SLTc is supplied to the lockup control valve 140 and the garage shift valve 160 via the oil passage 105, respectively. Engagement / release control of the lockup clutch 26 and engagement transition control of the forward clutch C1 and the reverse brake B1 are performed based on the control hydraulic pressure PSLT.

  The linear solenoid valve SLP is provided to perform pressure regulation control of the transmission hydraulic pressure PIN of the belt type continuously variable transmission 40.

  Specifically, the second modulator hydraulic pressure PM2 regulated by the second modulator valve is introduced into the linear solenoid valve SLP via the input port SLPa. When not energized, the hydraulic pressure PSLPmax determined by the urging force of the spring provided in the linear solenoid valve SLP is output as the control hydraulic pressure PSLP. When energized, the second modulator hydraulic pressure PM2 is linearly adjusted according to the exciting current. The pressure-controlled hydraulic pressure is output as the control hydraulic pressure PSLP. The control hydraulic pressure PSLP output from the output port SLPb is supplied to the transmission hydraulic pressure control valve 120. The pressure regulation control of the transmission hydraulic pressure PIN of the belt type continuously variable transmission 40 is performed based on the control hydraulic pressure PSLP.

  The linear solenoid valve SLS is provided to perform pressure regulation control of the line hydraulic pressure PL and pressure regulation control of the pinching hydraulic pressure POUT of the belt type continuously variable transmission 40.

  Specifically, the second modulator hydraulic pressure PM2 regulated by the second modulator valve is introduced into the linear solenoid valve SLS via the input port SLSa. When not energized, the hydraulic pressure PSLSmax determined by the urging force of the spring provided in the linear solenoid valve SLS is output as the control hydraulic pressure PSLS. When energized, the second modulator hydraulic pressure PM2 is linearly adjusted according to the excitation current. The pressure-controlled hydraulic pressure is output as the control hydraulic pressure PSLS. The control hydraulic pressure PSLS output from the output port SLSb is supplied to the primary regulator valve 110 and the clamping hydraulic control valve 130 via the oil passage 102, respectively. The pressure regulation control of the line hydraulic pressure PL and the pressure regulation control of the clamping hydraulic pressure POUT of the belt type continuously variable transmission 40 are performed based on the control hydraulic pressure PSLS.

  In the hydraulic control circuit 100 configured as described above, the pressure regulation control of the line hydraulic pressure PL, the engagement / release control of the lock-up clutch 26, and the forward clutch C1 and the reverse drive are performed by a single solenoid valve (linear solenoid valve SLT). The engagement transient control of the brake B1 is performed. In this embodiment, the control range (line hydraulic control range) of the pressure regulation control of the line hydraulic pressure PL by the linear solenoid valve SLT, the control range (lock up control range) of the engagement / release control of the lockup clutch 26, Are set respectively. That is, the pressure regulation control of the line hydraulic pressure PL and the engagement / release control of the lockup clutch 26 are not performed simultaneously. It should be noted that the engagement transition control of the forward clutch C1 and the reverse brake B1 by the linear solenoid valve SLT is performed when the garage shift valve 160 is held at the control position, and when it is held at the normal position. It is not done.

  The line hydraulic pressure control range and the lockup control range will be described with reference to FIGS. Here, it is assumed that the garage shift valve 160 is held in the normal position, and the engagement transition control of the forward clutch C1 and the reverse brake B1 is not performed.

  As shown in FIGS. 4 and 5, the line hydraulic pressure control range X1 and the lockup control range X2 are set for the exciting current energized to the linear solenoid valve SLT, and the control hydraulic pressure PSLT of the linear solenoid valve SLT is set. On the other hand, a line hydraulic pressure control range Y1 and a lockup control range Y2 are set. The line hydraulic pressure control range X1 (Y1) and the lockup control range X2 (Y2) are set to different ranges that do not overlap each other. Here, the line oil pressure control range X1 (Y1) and the lockup control range X2 (Y2) are divided by the switching point X3 (Y3) as a boundary.

  Specifically, the line hydraulic pressure control range X1 is set in a range where the exciting current to the linear solenoid valve SLT is smaller than the switching point X3, and the lockup control range X2 is set in a large range. Further, the line hydraulic pressure control range Y1 is set in a range where the control hydraulic pressure PSLT of the linear solenoid valve SLT is larger than the switching point Y3, and the lockup control range Y2 is set in a small range. The minimum value of the control hydraulic pressure PSLT is “0”, and the maximum value is PSLTmax.

  The setting of the line hydraulic pressure control range X1 (Y1) and the lockup control range X2 (Y2) is performed by the pressure reducing valve 180 provided between the linear solenoid valve SLT and the primary regulator valve 110.

  As described above, the output hydraulic pressure PCTL of the pressure reducing valve 180 is set to “0” when the control hydraulic pressure PSLT is equal to or lower than the pressure (W1 / S1) by the urging force of the spring 182 of the pressure reducing valve 180. PCTL is not introduced into the control hydraulic chamber 113c of the primary regulator valve 110. In this case, pressure regulation control of the line hydraulic pressure PL is not performed based on the control hydraulic pressure PSLT, and only engagement / release control of the lockup clutch 26 is performed based on the control hydraulic pressure PSLT. In this case, the pressure regulation control of the line hydraulic pressure PL is performed based on the control hydraulic pressure PSLS of the linear solenoid valve SLS, and therefore the pressure regulation control of the line hydraulic pressure PL is not performed in the lockup control range X2 (Y2). The situation has been avoided.

  Therefore, a range in which the control hydraulic pressure PSLT is equal to or lower than the pressure (W1 / S1) by the urging force of the spring 182 is set as the lockup control range X2 (Y2). Here, the switching point X3 (Y3) corresponds to the case where the control hydraulic pressure PSLT is equal to the pressure (W1 / S1) by the urging force of the spring 182. In this case, the output hydraulic pressure PCTL of the pressure reducing valve 180 is “ 0 ". The size (width) of the lockup control range X2 (Y2) is set according to the pressure (W1 / S1) by the urging force of the spring 182. In this case, as the pressure (W1 / S1) increases, the switching point X3 (Y3) is set to a smaller excitation current side (a side where the control hydraulic pressure PSLT is larger), and as a result, the lockup control range X2 (Y2). Is set larger.

  On the other hand, the output hydraulic pressure PCTL of the pressure reducing valve 180 is introduced into the control hydraulic pressure chamber 113c of the primary regulator valve 110 when the control hydraulic pressure PSLT exceeds the pressure (W1 / S1) by the urging force of the spring 182. In this case, as described above, the pressure regulation control of the line hydraulic pressure PL includes the combined force of the force by the output hydraulic pressure PCTL and the biasing force of the spring 112 of the primary regulator valve 110 and the force by the control hydraulic pressure PSLS of the linear solenoid valve SLS. Based on the higher one.

  Therefore, a range in which the control hydraulic pressure PSLT is equal to or higher than the pressure (W1 / S1) by the urging force of the spring 182 is set as the line hydraulic pressure control range X1 (Y1). Here, the size (width) of the line hydraulic pressure control range X1 (Y1) is set according to the pressure (W1 / S1) by the urging force of the spring 182. In this case, as the pressure (W1 / S1) is smaller, the switching point X3 (Y3) is set to a larger exciting current side (a smaller side of the control hydraulic pressure PSLT). As a result, the line hydraulic pressure control range X1 (Y1) Is set larger.

  In this line oil pressure control range X1 (Y1), the lockup clutch 26 is not engaged / released, and the lockup clutch 26 is held in a fully engaged state. For this reason, in the line oil pressure control range X1 (Y1), even when the control oil pressure PSLT changes, the spool 141 of the lockup control valve 140 is moved downward as much as possible (the state shown in the right half of FIG. 3). Retained. It should be noted that a state in which the spool 141 is moved downward as much as possible at the switching point X3 (Y3) may be obtained, or the spool 141 has already been moved downward as much as possible in the lockup control range X2 (Y2). You may be made to obtain the state which carried out.

  Here, as described above, when the gear ratio γ is higher than γ1 (see FIG. 6), it is preferable to set the line oil pressure PL to be equal to or slightly higher than the target clamping oil pressure. The pressure loss control of the line oil pressure PL is performed based on the control oil pressure PSLS of the linear solenoid valve SLS, whereby the drive loss of the oil pump 27 can be suppressed. On the other hand, when the transmission gear ratio γ is lower than γ1, it is preferable to set the line oil pressure PL to be equal to or slightly higher than the target transmission oil pressure. By performing based on the output hydraulic pressure PCTL of 180 (the control hydraulic pressure PSLT of the linear solenoid valve SLT), the drive loss of the oil pump 27 can be suppressed.

  On the other hand, when the garage shift valve 160 is held at the control position and the engagement transition control of the forward clutch C1 and the reverse brake B1 is performed, the engagement / release control of the lockup clutch 26 as described above is performed. The lockup clutch 26 is forcibly released and is controlled. Hereinafter, this control will be described.

  As described above, when the garage shift valve 160 is held at the control position, the control hydraulic pressure PSLT of the linear solenoid valve SLT is introduced into the spring chamber 144a of the lockup control valve 140. At this time, the control oil pressure PSLT is also introduced into the control oil pressure chamber 143 of the lockup control valve 140. In other words, the control hydraulic pressure PSLT is supplied to the lockup control valve 140 to the hydraulic chambers 143 and 144a on opposite sides (the hydraulic chambers at both ends of the spool 141).

  Here, if the lock-up control valve 140 is a two-stage valve in which the pressure receiving area on the spring hydraulic chamber 144a side of the operating area (pressure receiving area) of each control hydraulic pressure PSLT with respect to the spool 141 is increased, the control hydraulic chamber 143 is controlled. The combined force of the force by the control hydraulic pressure PSLT supplied to the spring hydraulic chamber 144a and the biasing force of the spring 142 exceeds the force by the control hydraulic pressure PSLT supplied to. Accordingly, the spool 141 is pressed upward. Accordingly, the lockup control valve 140 is held in the OFF state shown in the left half of FIG. 3 regardless of whether or not the control hydraulic pressure PSLT of the linear solenoid valve SLT is supplied to the control hydraulic pressure chamber 143. Along with this, the lockup clutch 26 is forcibly released.

  By such a forced OFF of the lock-up clutch 26, for example, even when a linear solenoid valve SLT ON failure occurs during a garage shift such as when the vehicle starts, the lock-up clutch 26 is reliably returned to the released state. It is possible to prevent engine stall. Moreover, since the hydraulic pressure (the control hydraulic pressure PSLT) supplied to the hydraulic chambers 143 and 144a at both ends of the spool 141 is the same, the lockup clutch 26 can be reliably returned to the released state regardless of the magnitude of the hydraulic pressure.

  By the way, in order to forcibly turn off the lockup clutch 26, it is conceivable to introduce the control hydraulic pressure of the ON-OFF solenoid valve SL1 for switching the garage shift valve 160 to the control position into the spring chamber 144a. However, in this case, since the hydraulic pressure supplied to the hydraulic chambers 143 and 144a at both ends of the spool 141 is different, it is difficult to make the lockup control valve 140 a two-stage valve as described above. Specifically, when the control hydraulic pressure of the ON-OFF solenoid valve SL1 is introduced into the spring chamber 144a and the lockup clutch 26 is forcibly turned OFF, the lockup control valve 140 remains a two-stage valve, and the control hydraulic pressure PSLT There is a possibility that the gain becomes small, and there is a possibility that the lockup capacity is insufficient in the lockup control range described above. On the other hand, in this embodiment, since the hydraulic pressure (the control hydraulic pressure PSLT) supplied to the hydraulic chambers 143 and 144a at both ends of the spool 141 is the same, the above-described two-stage valve can be achieved. Therefore, the lockup control valve 140 need not be a three-stage valve, and the structure of the valve body is simplified. In addition, the degree of freedom of the valve diameter is increased. As a result, an increase in manufacturing cost can be avoided.

  Here, it is preferable to perform the following control as the control at the time of restart after the engine stall occurs.

  First, one cause of the occurrence of engine stall is, for example, the ON failure of the linear solenoid valve SLT, specifically, the case where the linear solenoid valve SLT fails in a state where the control hydraulic pressure PSLT is at or near the maximum. When such an ON failure occurs, a situation occurs in which the control hydraulic pressure PSLT in a maximum state or a state close thereto is continuously input to the control hydraulic chamber 143 of the lockup control valve 140. Along with this, a situation occurs in which the lockup clutch 26 is held in the fully engaged state, and as a result, engine stall is likely to occur.

  In this embodiment, as described above, the lock-up clutch 26 is controlled by controlling the ON-OFF solenoid valve SL1 to the ON state and holding the garage shift valve 160 in the control position in order to prevent the engine stall. Is forcibly turned off to force the release state. This forced OFF control is performed while the garage shift valve 160 is held at the control position when the vehicle starts. Thereafter, the ON-OFF solenoid valve SL1 is controlled to be in the OFF state, and the garage shift is performed. When the valve 160 is held in the normal position, it is terminated. That is, when the garage shift valve 160 is held at the normal position, the forced OFF control of the lockup clutch 26 is released. However, when the ON failure of the linear solenoid valve SLT as described above has occurred, when the forced OFF control of the lockup clutch 26 is released, the lockup clutch 26 is again held in the fully engaged state, and the engine A situation where a stall is likely to occur is caused.

  Therefore, in this embodiment, when the engine restarts after the engine stall occurs (after the engine stall is detected), the forced OFF control of the lockup clutch 26 is continuously performed after the restart. ing. Specifically, at the time of restart after the engine stall, the ON-OFF solenoid valve SL1 is controlled to be in an ON state and is controlled to be kept in the ON state. As a result, even after restarting, the garage shift valve 160 remains held at the control position, and the control hydraulic pressure PSLT of the linear solenoid valve SLT continues to be supplied to the spring chamber 144a. As a result, the forced OFF control of the lockup clutch 26 is continuously performed. Therefore, even if the linear solenoid valve SLT is turned ON, it is possible to avoid the engine stall again. Then, the vehicle continues to travel with the garage shift valve 160 held at the control position, and the vehicle can be moved to a repair shop, for example.

  Note that the garage shift valve 160 may be switched to the normal position if the vehicle speed becomes high so that engine stall does not occur even when the lock-up clutch 26 is held in the fully engaged state. When the low vehicle speed that may occur is reached, the garage shift valve 160 is switched to the control position again.

1 is a diagram illustrating a schematic configuration of a vehicle drive device according to an embodiment. It is a block diagram which shows an example of the control system of the power transmission mechanism of the vehicle drive device of FIG. It is a circuit diagram which shows an example of the hydraulic control circuit for controlling the power transmission mechanism of the vehicle drive device of FIG. It is a figure which shows the characteristic of the output hydraulic pressure of the pressure-reduction valve with respect to the control hydraulic pressure of a linear solenoid valve. It is a figure which shows the characteristic of the control hydraulic pressure with respect to the exciting current of a linear solenoid valve. It is a figure which shows the change of the setting value of the target transmission hydraulic pressure and target clamping pressure oil pressure according to the gear ratio of a belt-type continuously variable transmission.

Explanation of symbols

DESCRIPTION OF SYMBOLS 20 Torque converter 26 Lockup clutch 30 Forward / reverse switching device C1 Forward clutch 40 Belt type continuously variable transmission 80 Electronic control device 100 Hydraulic control circuit 110 Primary regulator valve (line hydraulic control valve)
140 Lock-up control valve (lock-up control valve)
160 Garage shift valve (garage control valve)
180 Pressure reducing valve SLS, SLT Linear solenoid valve SL1 ON-OFF solenoid valve

Claims (6)

A belt-type continuously variable transmission that changes the gear ratio by changing the belt engagement diameter while transmitting the power by clamping the belt with hydraulic pressure;
A hydraulic lockup clutch provided in a fluid power transmission device provided between a power source and the belt type continuously variable transmission, and directly connecting the power source side and the belt type continuously variable transmission side;
A hydraulic travel friction engagement element that is engaged to establish a power transmission path during travel of the vehicle;
A line hydraulic control valve that regulates the line hydraulic pressure, which is the source pressure of each part,
A lockup control valve that is switched during engagement / release control of the lockup clutch;
In a hydraulic control device comprising a garage control valve capable of switching an engagement hydraulic pressure to be supplied when the travel friction engagement element is engaged to an engagement transient hydraulic pressure and an engagement holding hydraulic pressure;
When the garage control valve is switched so as to supply the engagement transient hydraulic pressure to the traveling friction engagement element, the control hydraulic pressure of the electromagnetic valve is the side supplied when the lockup control valve is controlled. A hydraulic control device that is supplied to the opposite side. The hydraulic control device according to claim 1,
The hydraulic control apparatus according to claim 1, wherein the control of the line hydraulic control valve and the control of the lockup control valve are performed by one electromagnetic valve. In the hydraulic control device according to claim 1 or 2,
The hydraulic control device according to claim 1, wherein when the garage control valve is switched so as to supply an engagement transient oil pressure to the frictional engagement element for traveling, the engagement transient oil pressure is controlled by the electromagnetic valve. In the hydraulic control device according to any one of claims 1 to 3,
A hydraulic control device configured to control the line hydraulic control valve by another electromagnetic valve other than the electromagnetic valve when the lockup control valve is controlled by the electromagnetic valve. The hydraulic control device according to claim 4,
The other solenoid valve is an electromagnetic valve that controls a clamping hydraulic control valve that supplies a clamping hydraulic pressure that controls the belt clamping pressure of the belt type continuously variable transmission to the driven pulley of the belt type continuously variable transmission. There is a hydraulic control device. In the hydraulic control device according to any one of claims 1 to 5,
When the engine restarts after an engine stall occurs, the garage control valve is switched to supply an engagement transient hydraulic pressure to the traveling friction engagement element, and the control hydraulic pressure of the electromagnetic valve is locked up. A hydraulic control device that continuously performs control to supply to a side opposite to a side to be supplied when controlling a control valve.

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