JP2019173821A - Hydraulic control device for vehicle - Google Patents

Abstract

To provide a hydraulic control device for a vehicle capable of suppressing degradation of followability to a target gear ratio of a continuously variable transmission, and erroneous switching to a fail-safe state of a sequence valve.SOLUTION: In a sequence valve 88, energization force to have a normal state is applied from a spring 88sp, a modulator pressure PLPM and an output pressure Psl2 of a solenoid valve SL2 for C2 are supplied as thrust to have the normal state, and an output pressure Pslp of a solenoid valve SLP for primary is supplied as thrust to have a fail-safe state. When the second clutch C2 is after start of engagement of clutch-to-clutch replacement control in a step S82, an upper limit guard pressure Pslpg of an SLP output pressure Pslp is determined by adding an effective clutch pack end pressure Ppeeff of the second clutch C2 to the sum of a spring load pressure SPld by the spring 88sp and a hydraulic pressure calculated by a PLPM pressure map in a step 86.SELECTED DRAWING: Figure 11

Description

  The present invention relates to a hydraulic control device for a vehicle including a plurality of power transmission paths provided in parallel between an engine and drive wheels.

  A vehicle in which a first power transmission path and a second power transmission path are provided in parallel between an input rotating member to which engine output torque is transmitted and an output rotating member for transmitting the output torque to drive wheels. The first power transmission path is via a gear mechanism which is a stepped transmission formed by engagement of the first clutch and the meshing clutch, and the second power transmission path is the second power transmission path. 2. Description of the Related Art There is known a hydraulic control device for a vehicle through a continuously variable transmission in which a transmission belt is wound between a primary pulley and a secondary pulley formed by engagement of the clutch. For example, this is the vehicle hydraulic control apparatus described in Patent Document 1. In the vehicle hydraulic control device described in Patent Document 1, the first clutch, the second clutch, the meshing clutch, the first brake, the primary brake, using hydraulic oil supplied from an oil pump that is rotationally driven by an engine. A hydraulic control circuit including a plurality of solenoid valves that output hydraulic pressures for operating the pulleys, the secondary pulleys, and the lock-up clutch is provided. In addition, when an abnormality occurs in any of the above solenoid valves, the oil path is switched by switching the sequence valve provided in the hydraulic control circuit from the normal state to the fail safe state, and the fail safe function is exhibited. It has come to be.

JP 2017-48898 A

  In the vehicle hydraulic control apparatus, a spring biasing force is applied to bring the sequence valve into a normal state. Further, the modulator pressure of the hydraulic oil and the hydraulic pressure for controlling the operation of the second clutch are used to bring the sequence valve into a normal state, and the primary pulley for controlling the operation of the primary pulley to make the sequence valve into a fail-safe state. Hydraulic pressure is used. At this time, if the modulator pressure is reduced, there is a possibility that an erroneous switching occurs in which the sequence valve is unintentionally switched to the fail-safe state. When the sequence valve is erroneously switched, the first clutch, the second clutch, the primary pulley, or the lock-up clutch operates differently from the driver's intention, and drivability deteriorates. Here, if the modulator pressure is increased by engine idle-up control, erroneous switching is suppressed, but in this case, fuel consumption is deteriorated. In addition, erroneous switching is suppressed by increasing the hydraulic pressure for controlling the operation of the second clutch. In this case, the second clutch is hardly released, and the first power transmission path via the gear mechanism is used. Driving becomes difficult. In addition, the erroneous switching is suppressed by lowering the hydraulic pressure for primary pulley operation control. However, if it is simply lowered, the primary pulley operation control is affected and the continuously variable transmission shift control is adversely affected. End up. For this reason, an upper limit guard pressure is provided in the hydraulic pressure for controlling the operation of the primary pulley in accordance with the operating state of the second clutch (full release, engagement transition, release transition, and full engagement states) to suppress erroneous switching. It is possible to do. However, in the transitional transition state where the second clutch transitions from the disengaged state to the engaged state, the upper limit guard pressure is provided in the hydraulic pressure for controlling the operation of the primary pulley. The followability to the target gear ratio of the step transmission may be deteriorated.

  The present invention has been made against the background of the above circumstances, and the purpose of the present invention is to prevent a sequence valve from being intended while suppressing deterioration in follow-up performance to a target gear ratio of a continuously variable transmission. An object of the present invention is to provide a vehicle hydraulic control device that can prevent erroneous switching to a fail-safe state.

  The gist of the present invention is (a) an input rotating member coupled to an engine, an output rotating member coupled to a drive wheel, and the input rotating member and the output rotating member are stepped transmissions. A first power transmission path that transmits the output torque of the engine via a gear mechanism, and a first power transmission path that transmits the output torque of the engine via a continuously variable transmission through the input rotation member and the output rotation member. Two power transmission paths, a first clutch that is provided in the first power transmission path, and connects and disconnects the first power transmission path, and a second clutch that is provided in the second power transmission path. An oil pump that is rotationally driven by the engine and hydraulic oil is supplied from the oil pump with respect to a vehicle that includes a second clutch that connects and disconnects the power transmission path of the vehicle and a first brake that engages during reverse travel. T A modulator valve for generating a modulator pressure, a C2 solenoid valve capable of supplying a first engagement pressure for operating the second clutch based on the modulator pressure, and the continuously variable transmission based on the modulator pressure. When both the solenoid valve group including the primary solenoid valve capable of supplying the second engagement pressure capable of adjusting the primary control pressure for operating the primary pulley and the solenoid valve included in the solenoid valve group are normal The normal state in which the first oil passage is formed and the fail-safe state in which the second oil passage is formed can be switched when any of the solenoid valves included in the solenoid valve group is abnormal. , A biasing force of a spring is applied as the normal state thrust, and the modulator pressure and the front And a sequence valve to which the second engagement pressure is supplied as a thrust to be brought into the fail-safe state, and (b) the engine When the second clutch is after the engagement of the clutch-to-clutch switching control is started when the engine rotation speed, which is a rotation speed of The upper limit guard pressure of the second engagement pressure is determined in advance so that the sequence valve does not switch to the second state according to the engine speed and the oil temperature of the hydraulic oil. The effective clutch pack end pressure of the second clutch is added to the sum of the load pressure of the urging force and the modulator pressure.

  According to the vehicle hydraulic control apparatus of the present invention, when all of the solenoid valves included in the solenoid valve group are normal, the normal state in which the first oil passage is formed and the solenoid included in the solenoid valve group. If any of the valves is abnormal, the fail-safe state in which the second oil passage is formed can be switched, and a biasing force of a spring is applied as the normal state thrust, and the modulator pressure and the A sequence valve to which the first engagement pressure is supplied and the second engagement pressure is supplied as a thrust for the fail-safe state is provided, and the engine rotation speed that is the rotation speed of the engine is a predetermined rotation speed And the second clutch is engaged in clutch-to-clutch switching control when the second clutch is in an engagement transition state. When it is after the start, the upper limit guard pressure of the second engagement pressure is determined in advance so that the sequence valve does not switch to the second state according to the engine speed and the oil temperature of the hydraulic oil. Further, the effective clutch pack end pressure of the second clutch is added to the sum of the load pressure of the biasing force of the spring and the modulator pressure. Thus, since the effective clutch pack end pressure of the second clutch is applied to the upper limit guard pressure of the second engagement pressure, the upper limit guard pressure of the second engagement pressure is relieved and continuously variable. It is possible to achieve both improvement in followability to the target speed ratio of the transmission and suppression of occurrence of erroneous switching of the sequence valve.

BRIEF DESCRIPTION OF THE DRAWINGS FIG. 1 is a skeleton diagram of a vehicle equipped with a vehicle hydraulic control apparatus according to an embodiment of the present invention, and is a diagram for explaining a control function of an electronic control device and various parts of a control system for various controls in the vehicle. It is a figure explaining the structure of the hydraulic control circuit of FIG. It is a block diagram of the sequence valve of FIG. FIG. 3 is a diagram illustrating a sequence valve switching operation when a second clutch is in a disengaged state in the hydraulic control circuit of FIG. 2. FIG. 3 is a diagram illustrating a sequence valve switching operation when a second clutch is engaged in the hydraulic control circuit of FIG. 2. 3 is an example of a method for setting an upper limit guard pressure of an SLP output pressure in accordance with the operating state of a second clutch in the hydraulic control circuit of FIG. FIG. 3 is a time chart when an upper limit guard pressure of an SLP output pressure is set according to an operation state of a second clutch in the hydraulic control circuit of FIG. 2. FIG. 5 is a state transition diagram showing the relationship between the operation state of the second clutch represented by four states of complete disengagement, complete engagement, disengagement transition, and engagement transition, and transition conditions for determining switching of the operation state. is there. 3 is a flowchart for explaining a control operation for setting an upper limit guard pressure of an SLP output pressure in accordance with an operation state of a second clutch in the hydraulic control circuit of FIG. 2. FIG. 10 is an example of a partial flowchart illustrating a control operation for determining an operation state of a second clutch in step S50 of the flowchart of FIG. 9; FIG. 10 is an example of a partial flowchart illustrating a control operation related to setting of an upper limit guard pressure at the time of engagement transition of the second clutch in step S80 of the flowchart of FIG. 9;

  In one embodiment of the present invention, the input rotation member is connected to the engine via a torque converter, the vehicle includes a meshing clutch provided on the first power transmission path, and the solenoid valve group includes: A lock-up solenoid valve capable of supplying a third engagement pressure for controlling engagement and disengagement of the lock-up clutch provided in the torque converter; and a fourth engagement pressure for operating the meshing clutch during forward running. And a solenoid valve for D1 that can supply the fourth engagement pressure that operates the first brake during reverse travel, and the sequence valve includes the primary solenoid valve, the lock When both the up solenoid valve and the D1 solenoid valve are normal, the normal state It is the primary solenoid valve, the lock-up solenoid valve, and wherein when any one of the abnormality of D1 solenoid valve is with the fail-safe state. As described above, when an abnormality occurs in any of the solenoids included in the solenoid valve group, the sequence valve is set in a fail-safe state to exhibit a fail-safe function.

  Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings. FIG. 1 is a skeleton diagram of a vehicle 10 equipped with a vehicle hydraulic control device 150 according to an embodiment of the present invention, as well as the control functions of the electronic control device 100 and various control systems for various controls in the vehicle 10. It is a figure explaining a part. The vehicle 10 includes, for example, an engine 12 that is used as a driving force source for traveling, a vehicle power transmission device (hereinafter referred to as a power transmission device) 16 that transmits the power of the engine 12 to driving wheels 14, and a hydraulic control circuit 46. And a vehicle hydraulic control device 150 having the electronic control device 100.

  The power transmission device 16 includes a torque converter 20 that is a fluid transmission device, an input shaft 22, a forward / reverse switching device 26, a continuously variable transmission 24, a gear mechanism 28, an output shaft 30, a counter shaft 32, an output shaft 30, and a counter shaft. 32 includes a reduction gear device 34 including a pair of gears that are provided so as not to rotate relative to each other and mesh with each other, a gear 36 provided so as not to rotate relative to the counter shaft 32, a differential gear 38 connected to the gear 36, and an axle 40. The power (torque) generated by the engine 12 is transmitted to the input shaft 22 via the torque converter 20. The power transmission device 16 includes a first power transmission path PT1 for transmitting power from the input shaft 22 to the output shaft 30 via the forward / reverse switching device 26 and the gear mechanism 28, and the continuously variable transmission 24 from the input shaft 22. The second power transmission path PT2 that transmits power to the output shaft 30 via the second power transmission path PT2 is configured in parallel so that it can be selectively established. The output shaft 30 transmits power to the drive wheels 14 via the reduction gear device 34, the counter shaft 32, the gear 36, the differential gear 38 and the axle 40. In order to switch between the first power transmission path PT1 and the second power transmission path PT2 in accordance with the traveling state of the vehicle 10, the power transmission device 16 includes a first clutch C1 as a forward clutch, which will be described later, A plurality of engagement devices including a first brake B1 as a brake, a second clutch C2 as a belt running mode clutch, and a meshing clutch D1 are provided. The input shaft 22 corresponds to the “input rotation member” in the present invention, and the output shaft 30 corresponds to the “output rotation member” in the present invention.

  The torque converter 20 includes a pump impeller 20p connected to a crankshaft of the engine 12 and a turbine impeller 20t connected to a forward / reverse switching device 26 via an input shaft 22 corresponding to an output side member of the torque converter 20. . A lockup clutch LU is provided between the pump impeller 20p and the turbine impeller 20t, and the pump impeller 20p and the turbine impeller 20t are integrally rotated by fully engaging the lockup clutch LU. It is done. The torque converter 20 includes an engagement side oil chamber 20on to which a hydraulic pressure for engaging the lockup clutch LU is supplied, and a release side oil chamber 20off to which a hydraulic pressure for releasing the lockup clutch LU is supplied.

  The oil pump 44 is a mechanical oil pump connected to the pump impeller 20p. The oil pump 44 is rotationally driven by the engine 12 to control the transmission of the continuously variable transmission 24, generate belt clamping pressure in the continuously variable transmission 24, and to engage each of the plurality of engagement devices. The hydraulic oil for switching the operation state such as engagement and release or switching the operation state of the lockup clutch LU is supplied to the hydraulic control circuit 46 provided in the vehicle 10.

  The forward / reverse switching device 26 is mainly composed of a first clutch C1, a first brake B1, and a double pinion type planetary gear device 26p. The planetary gear device 26p includes a sun gear 26s, a carrier 26c, and a ring gear 26r. The carrier 26c and the input shaft 22 are connected so as to be integrally rotated. The ring gear 26r is selectively coupled to the case 18 that is a non-rotating member via the first brake B1. The sun gear 26s and the carrier 26c are selectively coupled via the first clutch C1. The forward / reverse switching device 26 sets the first clutch C1 in the engaged state and releases the first brake B1 to move the vehicle 10 forward, and sets the first clutch C1 in the released state. It is possible to switch to the reverse mode in which the first brake B1 is engaged and the vehicle 10 travels backward. The first clutch C1 and the first brake B1, and the second clutch C2 and the meshing clutch D1 are all hydraulic engagement devices whose engagement state is controlled by a hydraulic actuator.

  The sun gear 26 s is connected to a small-diameter gear 48 that constitutes the gear mechanism 28. The gear mechanism 28 is provided with a small diameter gear 48, a gear mechanism counter shaft 50, and a large diameter which is provided around the gear mechanism counter shaft 50 coaxially with the gear mechanism counter shaft 50 so as not to rotate relative to the small diameter gear 48. And a gear 52. The large diameter gear 52 has a larger diameter than the small diameter gear 48. The gear mechanism 28 includes an idler gear 54 provided around the gear mechanism countershaft 50 so as to be rotatable relative to the gear mechanism countershaft 50 and coaxially with the output shaft 30 around the output shaft 30. And an output gear 42 which is provided so as not to be relatively rotatable and meshes with the idler gear 54. The output gear 42 has a larger diameter than the idler gear 54. Therefore, the gear mechanism 28 functions as a stepped transmission having one gear stage in the first power transmission path PT1 formed between the input shaft 22 and the output shaft 30.

  The gear mechanism 28 is provided between the large-diameter gear 52 and the idler gear 54 around the gear mechanism counter shaft 50, and is a meshing clutch D1 that selectively connects and disconnects the power transmission path between them. Is provided. The meshing clutch D1 is an engagement device that selectively connects or disconnects the first power transmission path PT1, and forms the first power transmission path PT1 when engaged. The meshing clutch D1 is engaged with the first clutch C1 or the first brake B1 to form a first power transmission path PT1. When both the first clutch C1 and the first brake B1 are released, or when the meshing clutch D1 is released, the first power transmission path PT1 is disconnected. The meshing clutch D1 includes a known synchromesh mechanism S1 as a synchronization mechanism that synchronizes rotation when engaged. The operating state of the meshing clutch D <b> 1 is switched by the operation of the hydraulic actuator 56 provided in the power transmission device 16. The meshing clutch D1 is provided with a spring 56a, and the meshing clutch D1 is applied with a biasing force in a direction to be released from the spring 56a.

  The continuously variable transmission 24 includes an input-side member provided on the input shaft 22 side that has a variable effective diameter primary pulley 58, an output-side member that has a variable effective diameter variable pulley 60, a primary pulley 58, and a secondary pulley. A transmission belt 62 wound around the pulley 60. The continuously variable transmission 24 is a belt type continuously variable transmission that transmits power via a frictional force between the primary pulley 58 and the secondary pulley 60 and the transmission belt 62. The frictional force is the same as the clamping pressure, and is also referred to as the belt clamping pressure. The primary pulley 58 is provided with a fixed sheave 58a as an input-side fixed rotating body coaxially attached to the input shaft 22, and is relatively non-rotatable around the axis and movable in the axial direction with respect to the fixed sheave 58a. A movable sheave 58b as an input side movable rotating body and a hydraulic actuator 58c that generates a thrust for moving the movable sheave 58b to change the V groove width therebetween are provided. The secondary pulley 60 includes a fixed sheave 60a serving as an output-side fixed rotating body, and a movable sheave 60b serving as an output-side movable rotating body provided so as not to rotate relative to the fixed sheave 60a around the axis and to be movable in the axial direction. And a hydraulic actuator 60c for generating a thrust force for moving the movable sheave 60b in order to change the V groove width therebetween. In the continuously variable transmission 24, the V groove width in the primary pulley 58 is changed to change the engagement diameter (effective diameter) of the transmission belt 62, whereby the transmission ratio γcvt (= input shaft rotational speed Nin / output shaft). The rotational speed Nout) is continuously changed. For example, when the V groove width of the primary pulley 58 is narrowed, the gear ratio γcvt is decreased. That is, the continuously variable transmission 24 is upshifted. When the V groove width of the primary pulley 58 is increased, the gear ratio γcvt increases, that is, the continuously variable transmission 24 is downshifted.

  The electronic control device 100 includes, for example, a so-called microcomputer having a CPU, a RAM, a ROM, an input / output interface, and the like, and the CPU uses a temporary storage function of the RAM according to a program stored in the ROM in advance. Various controls of the vehicle 10 are executed by performing signal processing. The electronic control unit 100 executes shift control of the continuously variable transmission 24, belt clamping pressure control, hydraulic control for switching the operation states of the plurality of engagement devices (C1, B1, C2, D1), and the like.

  The electronic control device 100 includes various sensors (for example, various rotational speed sensors 64, 66, 68, 70, an accelerator operation amount sensor 72, a throttle opening sensor 74, a shift position sensor 76, etc.) provided in the vehicle 10. Detection signals and the like (for example, the engine rotation speed Ne (rpm), the input shaft rotation speed Nin (rpm) and the primary rotation speed Npri (rpm), the secondary rotation speed Nsec (rpm), and the vehicle speed V (km / h) Corresponding output shaft rotation speed Nout (rpm), accelerator operation amount θacc (%) indicating the magnitude of acceleration operation by the driver, throttle opening degree tap (%), operation position POSsh of the shift lever 98 provided in the vehicle 10 Etc.) are input respectively. From the electronic control unit 100, various command signals (for example, a hydraulic control command signal for controlling a shift of the continuously variable transmission 24, a belt clamping pressure, etc.) to each device provided in the vehicle 10, such as a hydraulic control circuit 46, etc. Scvt, a hydraulic control command signal Scbd for controlling the operating state of each of the plurality of engagement devices, etc.) are respectively output. The input shaft rotational speed Nin is the same as the turbine rotational speed Nt (rpm) that is the rotational speed of the turbine impeller 20t and the primary rotational speed Npri that is the rotational speed of the primary pulley 58, and the secondary rotational speed Nsec is the secondary rotational speed Nsec. This is the rotational speed of the pulley 60. The electronic control unit 100 calculates a speed ratio γcvt (= Npri / Nsec) that is an actual speed ratio of the continuously variable transmission 24 based on the primary rotational speed Npri and the secondary rotational speed Nsec.

  The operation position POSsh of the shift lever 98 is, for example, a P, R, N, D operation position. The P operation position is a parking operation position for selecting the P position of the power transmission device 16 in which the power transmission device 16 is in a neutral state and the output shaft 30 is mechanically fixed so as not to rotate. The neutral state of the power transmission device 16 is realized, for example, by releasing all of the first clutch C1, the first brake B1, and the second clutch C2. That is, the neutral state of the power transmission device 16 is a state in which neither the first power transmission path PT1 nor the second power transmission path PT2 is formed. The R operation position is a reverse travel operation position for selecting the R position of the power transmission device 16 that enables reverse travel in the gear travel mode. The N operation position is a neutral operation position for selecting the N position of the power transmission device 16 that places the power transmission device 16 in the neutral state. The D operation position allows the forward travel in the gear travel mode or the automatic transmission control of the continuously variable transmission 24 in the belt travel mode to enable the forward travel. This is the forward travel operation position for selecting the position.

  FIG. 2 is a diagram illustrating the configuration of the hydraulic control circuit 46 of FIG. In this specification, “supplying hydraulic pressure (including engagement pressure, output pressure, and control pressure)” means “supplying hydraulic oil having such hydraulic pressure”.

  The hydraulic control circuit 46 adjusts the hydraulic pressure generated by the oil pump 44 by a primary regulator valve and a secondary regulator valve (not shown) according to the load of the engine 12 represented by, for example, the throttle opening degree tap. A first line pressure PL1 (MPa) and a second line pressure PL2 (MPa) are generated.

  The hydraulic control circuit 46 includes a C1 solenoid valve SL1, a C2 solenoid valve SL2, a D1 solenoid valve SLG, a primary solenoid valve SLP, a secondary solenoid valve SLS, and a lockup solenoid valve SLU. Are included in the solenoid valve group SLgr. The hydraulic control circuit 46 includes a modulator valve 80, a manual valve 82, a primary pressure control valve 84, a secondary pressure control valve 86, a sequence valve 88, a C1 control valve 90, an S1B1 control valve 92, an LU pressure control valve 94, and an accumulator. 96. The lockup solenoid valve SLU, the C1 solenoid valve SL1, the C2 solenoid valve SL2, and the D1 solenoid valve SLG shut off the input port and the output port when not energized (hereinafter also referred to as OFF) This is a so-called normally closed (N / C) type linear solenoid valve that communicates when energized (hereinafter also referred to as ON). The primary solenoid valve SLP and the secondary solenoid valve SLS are so-called normally open (N / O) type linear solenoid valves that shut off the input port and the output port when energized and communicate with each other when de-energized.

  The hydraulic control circuit 46 is operated by hydraulic pressure, and is operated by hydraulic pressure, a hydraulic actuator C1a capable of connecting / disconnecting the first clutch C1, a hydraulic actuator C2a operated by hydraulic pressure and capable of connecting / disconnecting the second clutch C2, and hydraulic pressure. The hydraulic actuator 56 capable of connecting / disconnecting the engagement clutch D1 and the hydraulic actuator B1a operated by hydraulic pressure and capable of connecting / disconnecting the first brake B1 are connected.

  The modulator valve 80 regulates the first line pressure PL1 to be lower than the first line pressure PL1 in a pressure regulation region in which the first line pressure PL1, which is the original pressure, is higher than the saturation modulator pressure PLPMmax (MPa), which is a constant pressure. A constant modulator pressure PLPM (MPa) is generated. However, in the non-regulation region where the first line pressure PL1, which is the original pressure, is lower than the saturation modulator pressure PLPMmax, the first line pressure PL1, which is the original pressure, is output as it is as the modulator pressure PLPM. Therefore, the modulator pressure PLPM increases with an increase in the engine rotational speed Ne of the engine 12, and is saturated and saturated when the engine rotational speed Ne becomes equal to or higher than a predetermined saturation rotational speed Neh (rpm). It has come to be. The modulator pressure PLPM is affected by the flow rate of hydraulic oil due to the operation of each actuator described later, and decreases or rises due to the flow rate balance due to the supply of hydraulic oil from the oil pump 44 and the consumption of hydraulic fluid by the actuator and the like. Or

  In the manual valve 82, the oil passage is mechanically switched in conjunction with the switching operation of the shift lever 98 by the driver. When the shift lever 98 is in the D operation position, the manual valve 82 outputs the input modulator pressure PLPM as the forward hydraulic pressure PD, and when the shift lever 98 is in the R operation position, the manual valve 82 outputs the input modulator pressure PLPM as the reverse hydraulic pressure. Output as PR. When the shift lever 98 is in the N operation position or the P operation position, the manual valve 82 cuts off the output of the hydraulic pressure, and the forward hydraulic pressure PD and the reverse hydraulic pressure PR are set to the drain pressure (the hydraulic pressure of the discharge oil passage EX).

  The C1 solenoid valve SL1 outputs an SL1 output pressure Psl1 that can be a C1 control pressure Pc1, which is a hydraulic pressure supplied to the hydraulic actuator C1a of the first clutch C1, using the forward hydraulic pressure PD as a base pressure. The C2 solenoid valve SL2 outputs an SL2 output pressure Psl2 that can be a C2 control pressure Pc2, which is a hydraulic pressure supplied to the hydraulic actuator C2a of the second clutch C2, using the forward hydraulic pressure PD as a base pressure. The SL2 output pressure Psl2 corresponds to the “first engagement pressure” in the present invention.

  The D1 solenoid valve SLG becomes a D1 control pressure Pd1, which is a hydraulic pressure supplied to the hydraulic actuator 56 in order to switch the operating state of the meshing clutch D1, using an output pressure Psbv output from an S1B1 control valve 92 described later as a source pressure. The obtained SLG output pressure Pslg is output. Note that the SLG output pressure Pslg is the hydraulic pressure B1 that is supplied to the hydraulic actuator B1a of the first brake B1 during reverse travel in which the shift lever 98 is set to the R operation position and the reverse hydraulic pressure PR is output from the manual valve 82. It can be the control pressure Pb1.

  The primary solenoid valve SLP outputs the SLP output pressure Pslp for controlling the primary control pressure Ppri, which is the hydraulic pressure supplied to the hydraulic actuator 58c of the primary pulley 58, using the modulator pressure PLPM as a source pressure. Secondary solenoid valve SLS outputs SLS output pressure Psls for controlling secondary control pressure Psec, which is the hydraulic pressure supplied to hydraulic actuator 60c of secondary pulley 60, using modulator pressure PLPM as a source pressure. The SLP output pressure Pslp corresponds to the “second engagement pressure” in the present invention.

  The lockup solenoid valve SLU is a hydraulic pressure supplied to the engagement side oil chamber 20on and the release side oil chamber 20off in the torque converter 20 functioning as a hydraulic actuator of the lockup clutch LU using the modulator pressure PLPM as a source pressure. An SLU output pressure Pslu for controlling the lockup control pressure Plu is output. The lockup control pressure Plu is a differential pressure between an on pressure Pon supplied to the engagement side oil chamber 20on and an off pressure Poff supplied to the release side oil chamber 20off.

  The primary pressure control valve 84 regulates the primary control pressure Ppri using the first line pressure PL1 as a source pressure, and outputs the primary control pressure Ppri to the hydraulic actuator 58c of the primary pulley 58. The primary pressure control valve 84 includes a spool valve element whose position can be freely switched between a fully open state and a fully closed state, and a spring that urges the spool valve element in the fully open state position direction. The primary pressure control valve 84 receives the SLP output pressure Pslp as a thrust toward the fully open position of the spool valve element, and the primary control pressure Ppri and a sequence valve described later as a thrust toward the fully closed position of the spool valve element. The output pressure Pseq1 output from 88 is input. Primary pressure control valve 84 regulates primary control pressure Ppri based on SLP output pressure Pslp, primary control pressure Ppri, and output pressure Pseq1. In the primary pressure control valve 84, the primary control pressure Ppri is increased when the SLP output pressure Pslp is increased.

  The secondary pressure control valve 86 regulates the secondary control pressure Psec using the first line pressure PL <b> 1 as a source pressure, and outputs the secondary control pressure Psec to the hydraulic actuator 60 c of the secondary pulley 60. The secondary pressure control valve 86 includes a spool valve element whose position can be freely switched between a fully open state and a fully closed state, and a spring that urges the spool valve element in the fully open state position direction. The secondary pressure control valve 86 receives the SLS output pressure Psls as the thrust toward the fully open position of the spool valve element, and receives the secondary control pressure Psec as the thrust toward the fully closed position of the spool valve element. The secondary pressure control valve 86 regulates the secondary control pressure Psec based on the SLS output pressure Psls and the secondary control pressure Psec. In the secondary pressure control valve 86, the secondary control pressure Psec is increased as the SLS output pressure Psls increases.

  The LU pressure control valve 94 receives the second line pressure PL2. The LU pressure control valve 94 outputs the second line pressure PL2 to the engagement side oil chamber 20on as the on pressure Pon, and the on position where the release side oil chamber 20off communicates with the unillustrated discharge oil passage EX, and the off pressure. A spool valve (not shown) that is operated between an off position that outputs the second line pressure PL2 to the release side oil chamber 20off as Poff and communicates the engagement side oil chamber 20on with a discharge oil passage EX (not shown). A child 94sv and a spring 94sp (not shown) that urges the spool valve element 94sv in the off position direction are provided. The LU pressure control valve 94 receives the off pressure Poff and the SLU output pressure Pslu as thrust in the on-position direction of the spool valve element 94sv, and the on-pressure Pon and the sequence valve 88 as thrust in the off-position direction of the spool valve element 94sv. Output pressure Pseq1. The LU pressure control valve 94 controls a lockup control pressure Pl, which is a differential pressure between the on-pressure Pon in the engagement-side oil chamber 20on and the off-pressure Poff in the release-side oil chamber 20off, according to the SLU output pressure Pslu. Thus, the connection / disconnection state of the lockup clutch LU is controlled. When the SLU output pressure Pslu is input to the LU pressure control valve 94, the spool valve element 94sv is switched to the on position side. When an SLU output pressure Pslu that is equal to or higher than a predetermined pressure is input, the lockup clutch LU is switched from the released state to the locked up state that is in the engaged state.

  The accumulator 96 is connected to an oil passage through which the forward hydraulic pressure PD flows. The accumulator 96 is a known accumulator that includes a seal member that suppresses leakage of a spring and hydraulic oil, and is capable of accumulating hydraulic pressure and supplying the accumulated hydraulic pressure. When the hydraulic pressure in the oil passage through which the forward hydraulic pressure PD circulates is higher than the hydraulic pressure in the accumulator 96, the hydraulic pressure is supplied from the oil passage to the accumulator 96, and the oil passage through which the forward hydraulic pressure PD circulates rather than the hydraulic pressure in the accumulator 96. When the hydraulic pressure is low, the hydraulic pressure is supplied from the accumulator 96 to the oil passage.

  The C1 control valve 90 includes a spool valve element 90sv (not shown) whose position is alternatively switched between a normal position and a tie-up prevention position, and a spring 90sp that biases the spool valve element 90sv in the normal position direction. When the spool valve element 90sv is in the normal position, the C1 control valve 90 is in the normal state, and when the spool valve element 90sv is in the tie-up prevention position, the C1 control valve 90 is in the tie-up prevention state. To do. The C1 control valve 90 receives the C2 control pressure Pc2, the SL1 output pressure Psl1, and the reverse hydraulic pressure PR as thrust in the tie-up prevention position direction of the spool valve element 90sv, and as the thrust in the normal position direction of the spool valve element 90sv. The output pressure Pseq2 of the sequence valve 88 is input. The C1 control valve 90 receives the SL1 output pressure Psl1, the SLG output pressure Pslg, and the reverse hydraulic pressure PR as input pressures for the switching oil passage. In the normal state, as shown by the solid line in FIG. 2, the SL1 output pressure Psl1 is output as the C1 control pressure Pc1, the SLG output pressure Pslg is output as the output pressure Pc1v1, and the input of the reverse hydraulic pressure PR is cut off. Not output. In the tie-up prevention state, as shown by the broken line in FIG. 2, the C1 control pressure Pc1 is the drain pressure, the reverse hydraulic pressure PR is output as the output pressure Pc1v1 and the output pressure Pc1v2, and the SL1 output pressure Psl1 and the SLG output pressure Pslg The input is blocked and not output to either.

  In the C1 control valve 90, the pressure receiving area of the C2 control pressure Pc2 acting as a thrust on the spool valve element 90sv and the pressure receiving area of the output pressure Pseq2 acting as a thrust on the spool valve element 90sv are set to be the same. In the present specification, the pressure receiving area means an area where the spool valve element that contributes to the thrust toward the normal position or the fail-safe position receives the hydraulic pressure, and the normal position and the fail-safe position in the same hydraulic oil chamber. In the case where a thrust force directed in both directions acts, the area of the difference between the one contributing to the thrust force directed to the normal position and the one directed toward the fail-safe position in the area where the spool valve element receives the hydraulic pressure. The biasing force of the spring 90sp is set to be smaller than the thrust in the tie-up prevention position direction with respect to the spool valve element 90sv by the SL1 output pressure Psl1. Thus, when the C1 solenoid valve SL1 and the C2 solenoid valve SL2 are simultaneously operated and the SL1 output pressure Psl1 and the SL2 output pressure Psl2 are output simultaneously, respectively, the thrust acting on the spool valve element 90sv is the C2 control pressure Pc2. (Ie, the SL2 output pressure Psl2) and the output pressure Pseq2 (ie, the modulator pressure PLPM) cancel each other, and the SL1 output pressure Psl1 resists the biasing force of the spring 90sp, so that the spool valve element 90sv is in the tie-up prevention position. Can be switched.

  The biasing force of the spring 90sp is set to be smaller than the thrust in the tie-up prevention position direction with respect to the spool valve element 90sv by the reverse hydraulic pressure PR. Thus, when the C2 solenoid valve SL2 is operated at the reverse travel operation position and the reverse hydraulic pressure PR and the SL2 output pressure Psl2 are simultaneously output, the thrust acting on the spool valve element 90sv is the C2 control pressure Pc2 (ie, SL2 The output pressure Psl2) and the output pressure Pseq2 (that is, the modulator pressure PLPM) cancel each other, and the reverse hydraulic pressure PR is switched to the tie-up prevention position against the urging force of the spring 90sp.

  Therefore, when the C1 solenoid valve SL1 is operated and the C2 solenoid valve SL2 is not operated, the C1 control valve 90 remains in the normal state, and the SL1 output pressure Psl1 is supplied to the hydraulic actuator C1a. When the C2 solenoid valve SL2 is activated and the SL2 output pressure Psl2 is output from the sequence valve 88 as the C2 control pressure Pc2, and the C1 solenoid valve SL1 is not activated, the C1 control valve 90 remains in the normal state. , SL2 output pressure Psl2 is supplied to the hydraulic actuator C2a. When the C2 solenoid valve SL2 is activated and the SL2 output pressure Psl2 is output from the sequence valve 88 as the C2 control pressure Pc2, and the C1 solenoid valve SL1 is also activated, the C1 control valve 90 is switched to the tie-up prevention state. The hydraulic actuator C1a is set to the drain pressure, and the SL2 output pressure Psl2 is supplied to the hydraulic actuator C2a. As described above, since the engagement pressure is prevented from being simultaneously supplied to the hydraulic actuator C1a and the hydraulic actuator C2a, the occurrence of the tie-up state due to the simultaneous engagement of the first clutch C1 and the second clutch C2 is prevented. can do. The C1 control valve 90 cuts off the supply of the SL1 output pressure Psl1 as the C1 control pressure Pc1 to the hydraulic actuator C1a, so that the tie-up state by the simultaneous engagement of the first clutch C1 and the second clutch C2 is achieved. Functions as a fail-safe valve to prevent.

  The S1B1 control valve 92 includes a spool valve element 92sv (not shown) whose position is alternatively switched between a non-reverse position and a reverse position, and a spring 92sp that urges the spool valve element 92sv in the non-reverse position direction. The S1B1 control valve 92 is said to be in the non-reverse position when the spool valve element 92sv is in the non-reverse position, and the S1B1 control valve 92 is in the reverse condition when the spool valve element 92sv is in the reverse position. The S1B1 control valve 92 receives the reverse hydraulic pressure PR as a thrust toward the reverse position of the spool valve element 92sv, and receives the SLS output pressure Psls as the thrust toward the non-reverse position of the spool valve element 92sv. In the S1B1 control valve 92, the output pressures Pc1v1, Pc1v2, the modulator pressure PLPM, and the reverse hydraulic pressure PR of the C1 control valve 90 are input as the input pressure of the switching oil passage. In the non-reverse position, the output pressure Pc1v2 is output as the B1 control pressure Pb1, the output pressure Pc1v1 is output as the D1 control pressure Pd1, and the modulator pressure PLPM is output as the output pressure Psbv as shown by the solid line in FIG. The input of the hydraulic pressure PR is cut off and is not output to either. In the reverse position, as indicated by the broken line in FIG. 2, the output pressure Pc1v1 is output as the B1 control pressure Pb1, the modulator pressure PLPM is output as the D1 control pressure Pd1, the reverse hydraulic pressure PR is output as the output pressure Psbv, and the output pressure The input of Pc1v2 is blocked and is not output to any.

  In the S1B1 control valve 92, the pressure receiving area of the reverse hydraulic pressure PR acting as a thrust on the spool valve element 92sv and the pressure receiving area of the SLS output pressure Psls acting as a thrust on the spool valve element 92sv are set to be the same. The urging force of the spring 92sp is set to be smaller than the thrust in the reverse drive position direction with respect to the spool valve element 92sv by the reverse hydraulic pressure PR. Thus, when the reverse hydraulic pressure PR is input and the SLS output pressure Psls is not input, the spool valve element 92sv is switched to the reverse position, and when the reverse hydraulic pressure PR and the SLS output pressure Psls are input simultaneously, the S1B1 control is performed. The valve 92 is switched to the non-reverse state. Therefore, the SLS output pressure Psls is input even if the reverse hydraulic pressure PR is not output from the manual valve 82 when the shift lever 98 is other than the reverse travel operation position, or the reverse hydraulic pressure PR is output when the shift lever 98 is the reverse travel operation position. In this case, the S1B1 control valve 92 is in a non-reverse state.

  The sequence valve 88 includes a spool valve 88sv (not shown) that can be switched between a normal position and a fail-safe position, and a spring 88sp that urges the spool valve 88sv in the normal position direction. The sequence valve 88 is in a normal state when the spool valve element 88sv is in a normal position, and the sequence valve 88 is in a fail-safe state when the spool valve element 88sv is in a fail-safe position. The sequence valve 88 receives the SLP output pressure Pslp as the thrust toward the fail-safe position of the spool valve 88sv, and receives the modulator pressure PLPM and the SL2 output pressure Psl2 as the thrust toward the normal position of the spool valve 88sv. . In the sequence valve 88, the forward hydraulic pressure PD, the SL2 output pressure Psl2, the SLG output pressure Pslg, and the modulator pressure PLPM are input as the input pressure of the switching oil passage. When the spool valve 88sv is in the normal position, the sequence valve 88 forms a first oil passage, and the SL2 output pressure Psl2 is output as the C2 control pressure Pc2 and the modulator pressure PLPM is output as shown by the solid line in FIG. It is output as Pseq2, the output pressure Pseq1 is set as the drain pressure, and the inputs of the forward hydraulic pressure PD and the SLG output pressure Pslg are blocked and are not output to any of them. When the spool valve 88sv is in the fail-safe position, the sequence valve 88 forms a second oil passage, and the forward hydraulic pressure PD is output as the C2 control pressure Pc2 and the SLG output pressure Pslg is output as shown by the broken line in FIG. The pressure Pseq1 is output, the output pressure Pseq2 is the drain pressure, and the SL2 output pressure Psl2 and the modulator pressure PLPM are blocked and are not output to any of them.

  FIG. 3 is a block diagram of the sequence valve 88 of FIG. The sequence valve 88 includes a spool valve 88sv that can be switched between a normal position (right half position) and a fail-safe position (left half position), and a spring that includes a compression coil spring that biases the spool valve 88sv toward the normal position. 88sp. The sequence valve 88 includes a first hydraulic oil chamber 88c1 to which a modulator pressure PLPM that is a thrust toward the normal position of the spool valve 88sv is input, and an SL2 output pressure that is a thrust toward the normal position of the spool valve 88sv. A second hydraulic oil chamber 88c2 to which Psl2 is input; and a third hydraulic oil chamber 88c3 to which an SLP output pressure Pslp that is a thrust toward the fail-safe position of the spool valve element 88sv is input. The sequence valve 88 includes a first input port 88i1 to which the SL2 output pressure Psl2 is input, a second input port 88i2 to which the forward hydraulic pressure PD is input, and a third input port 88i3 to which the SLG output pressure Pslg is input. And a fourth input port 88i4 to which the modulator pressure PLPM is input. The sequence valve 88 outputs a first output port 88o1 that outputs the C2 control pressure Pc2 to the hydraulic actuator C2a of the second clutch C2, and a second output that outputs the output pressure Pseq1 to the LU pressure control valve 94 and the primary pressure control valve 84. Output port 88o2, a third output port 88o3 for outputting the output pressure Pseq2 to the C1 control valve 90, and a drain port 88ex.

  Next, the operation of the hydraulic control circuit 46 will be described in detail.

  After the engine 12 is driven, when the oil pump 44 is rotationally driven to generate the first line pressure PL1, the modulator valve PL generates the modulator pressure PLPM. The sequence valve 88 is switched to a normal state by the modulator pressure PLPM. The modulator pressure PLPM is input to the C1 control valve 90 as the output pressure Pseq2 of the sequence valve 88, and the C1 control valve 90 is switched to the normal state. As a result, during normal travel where no failure occurs, the sequence valve 88 is in a normal state and the C1 control valve 90 is in a normal state.

  When the vehicle 10 travels at a low speed, the power transmission device 16 enters the forward gear travel mode. At this time, control is performed so that only the first clutch C1 and the meshing clutch D1 are engaged. Specifically, the forward hydraulic pressure PD is output from the manual valve 82, and the SL1 output pressure Psl1 is output from the C1 solenoid valve SL1 using the forward hydraulic pressure PD as a source pressure. The SL1 output pressure Psl1 is supplied to the hydraulic actuator C1a via the C1 control valve 90. Since the reverse hydraulic pressure PR is not output from the manual valve 82, the S1B1 control valve 92 is in the non-reverse state. The modulator pressure PLPM is input as the original pressure to the D1 solenoid valve SLG as the output pressure Psbv of the S1B1 control valve 92, and the SLG output pressure Pslg is output from the D1 solenoid valve SLG. The SLG output pressure Pslg is supplied to the hydraulic actuator 56 via the C1 control valve 90 and the S1B1 control valve 92.

  During medium-to-high speed forward traveling of the vehicle 10, the power transmission device 16 enters the forward belt traveling mode. At this time, the engagement clutch D1 is controlled to be engaged according to the second clutch C2 and the vehicle speed V. Specifically, the SL2 output pressure Psl2 is output from the C2 solenoid valve SL2 using the forward hydraulic pressure PD as a source pressure. The SL2 output pressure Psl2 is supplied to the hydraulic actuator C2a via the sequence valve 88.

  When the vehicle 10 is traveling backward, the power transmission device 16 is in a reverse gear traveling mode. At this time, control is performed so that only the first brake B1 and the meshing clutch D1 are engaged. Specifically, when the reverse hydraulic pressure PR is output from the manual valve 82, the S1B1 control valve 92 is switched to the reverse state. The reverse hydraulic pressure PR is supplied as the original pressure to the D1 solenoid valve SLG as the output pressure Psbv of the S1B1 control valve 92, and the SLG output pressure Pslg is output from the D1 solenoid valve SLG. The SLG output pressure Pslg is supplied to the hydraulic actuator B1a via the C1 control valve 90 and the S1B1 control valve 92. Further, the modulator pressure PLPM is supplied to the hydraulic actuator 56 via the S1B1 control valve 92.

  Next, an operation when an abnormality occurs in any of the primary solenoid valve SLP, the lockup solenoid valve SLU, and the D1 solenoid valve SLG will be described.

  A case where an off-failure abnormality has occurred in the primary solenoid valve SLP at the forward travel operation position will be described. In this case, when the maximum SLP output pressure Pslp is always output from the primary solenoid valve SLP and the maximum primary control pressure Ppri is always output from the primary pressure control valve 84, the continuously variable transmission 24 is used. Running becomes difficult. However, the maximum SLP output pressure Pslp output from the primary solenoid valve SLP switches the sequence valve 88 to the fail-safe state. As a result, the modulator pressure PLPM is input as the original pressure to the D1 solenoid valve SLG as the output pressure Psbv of the S1B1 control valve 92 in the non-reverse state, and the SLG output pressure Pslg is output from the D1 solenoid valve SLG. The SLG output pressure Pslg is input to the primary pressure control valve 84 as the output pressure Pseq1 of the sequence valve 88 in the fail-safe state. Therefore, the primary control pressure Ppri output from the primary pressure control valve 84 can be controlled by controlling the SLG output pressure Pslg output from the D1 solenoid valve SLG, and a fail-safe function capable of ensuring safety even when an abnormality occurs. Is demonstrated.

  A case where an on-fail abnormality has occurred in the primary solenoid valve SLP at the forward travel operation position will be described. In this case, if the SLP output pressure Pslp is not output from the primary solenoid valve SLP and the minimum primary control pressure Ppri is always output from the primary pressure control valve 84, the speed ratio γcvt of the continuously variable transmission 24 is maximized. There is a risk of sudden deceleration. When the on-fail abnormality of the primary solenoid valve SLP is detected, the electronic control unit 100 stops the output of the SL2 output pressure Psl2 from the C2 solenoid valve SL2, and stops the supply of the SL2 output pressure Psl2 to the hydraulic actuator C2a. To do. As a result, the second clutch C2 is disengaged and the neutral range is set to avoid sudden deceleration. After the vehicle speed V has stabilized, the electronic control unit 100 can engage the first clutch C1 with the SL1 output pressure Psl1 output from the C1 solenoid valve SL1, and can travel forward in the gear travel mode. Is demonstrated.

  A case where an on-fail abnormality has occurred in the lockup solenoid valve SLU at the forward travel operation position will be described. In this case, the lockup clutch LU cannot be released. When detecting an on-fail abnormality of the lockup solenoid valve SLU, the electronic control unit 100 outputs the maximum SLP output pressure Pslp from the primary solenoid valve SLP, and switches the sequence valve 88 to the fail-safe state. As a result, the modulator pressure PLPM is input as the original pressure to the D1 solenoid valve SLG as the output pressure Psbv of the S1B1 control valve 92 in the non-reverse state, and the SLG output pressure Pslg is output from the D1 solenoid valve SLG. The SLG output pressure Pslg is input to the LU pressure control valve 94 as the output pressure Pseq1 of the sequence valve 88 in the fail safe state, whereby the lockup clutch LU is released and the fail safe function is exhibited.

  A case where an on-fail abnormality has occurred in the D1 solenoid valve SLG at the forward travel operation position will be described. In this case, if the SLG output pressure Pslg is always output from the D1 solenoid valve SLG, the meshing clutch D1 remains engaged. When detecting an on-fail abnormality of the D1 solenoid valve SLG, the electronic control unit 100 outputs the maximum SLP output pressure Pslp from the primary solenoid valve SLP and switches the sequence valve 88 to the fail-safe state. As a result, the output pressure Pseq2 of the sequence valve 88 set as the drain pressure and the C2 control pressure Pc2 set as the forward hydraulic pressure PD are input to the C1 control valve 90, and the C1 control valve 90 is switched to the tie-up prevention state. The C2 control pressure Pc2, which is the forward hydraulic pressure PD, is supplied to the hydraulic actuator C2a. By switching the C1 control valve 90 to the tie-up prevention state, the input of the SLG output pressure Pslg is cut off and is not output to any, and the hydraulic actuator C1a is set to the drain pressure. The output pressure Pc1v1 of the C1 control valve 90 is cut off from the SLG output pressure Pslg, and the reverse hydraulic pressure PR is output to the output pressure Pc1v1, but the reverse hydraulic pressure PR communicates with the discharge oil passage EX at the forward travel operation position. Therefore, since the output pressure Pc1v1 communicated with the exhaust oil passage EX is supplied to the hydraulic actuator 56 via the non-reverse S1B1 control valve 92, the meshing clutch D1 is switched from the engaged state to the released state, and the fail-safe function Is demonstrated.

  A case where an off-failure abnormality has occurred in the D1 solenoid valve SLG at the reverse travel operation position will be described. In this case, if the SLG output pressure Pslg is not output from the D1 solenoid valve SLG, the first brake B1 is released and the neutral state is established. When the electronic control device 100 detects an off-failure abnormality of the D1 solenoid valve SLG, the electronic control device 100 outputs the maximum SLP output pressure Pslp from the primary solenoid valve SLP and switches the sequence valve 88 to the fail-safe state. As a result, the output pressure Pseq2 of the sequence valve 88 set as the drain pressure and the C2 control pressure Pc2 set as the forward hydraulic pressure PD are input to the C1 control valve 90, and the C1 control valve 90 is switched to the tie-up prevention state. The reverse hydraulic pressure PR is supplied as the output pressure Pc1v1 of the C1 control valve 90 to the hydraulic actuator B1a of the first brake B1 via the S1B1 control valve 92. The modulator pressure PLPM is supplied to the hydraulic actuator 56 of the meshing clutch D1 via the S1B1 control valve 92. As a result, the first brake B1 and the meshing clutch D1 are engaged to allow reverse travel, and the fail-safe function is exhibited.

Here, in the sequence valve 88, the first hydraulic oil chamber 88c1 to which the modulator pressure PLPM (MPa) is input has a pressure receiving area AL1 (mm ² ), and the SL2 output pressure Psl2 (MPa) is input. The second hydraulic oil chamber 88c2 has a pressure receiving area AL2 (mm ² ), and the third hydraulic oil chamber 88c3 to which the SLP output pressure Pslp (MPa) is input has a pressure receiving area AL3 (mm ² ). To do. The magnitude of the urging force that causes the spring 88sp of the sequence valve 88 to move the spool valve element 88sv toward the normal position is defined as a spring load SP (N). Therefore, the thrust DFn (N) that directs the spool valve 88sv of the sequence valve 88 to the normal position is expressed by the following expression (1), and the thrust DFf (N ) Is represented by the following formula (2). When the thrust DFn is larger than the thrust DFf, the sequence valve 88 is switched to the normal state, and when the thrust DFn is smaller than the thrust DFf, the sequence valve 88 is switched to the fail-safe state.
DFn = SP + PLPM × AL1 + Psl2 × AL2 (1)
DFf = Pslp × AL3 (2)

  In the above formulas (1) and (2), the spring load SP is constant without depending on the modulator pressure PLPM. The SLP output pressure Pslp is based on the modulator pressure PLPM, and the SL2 output pressure Psl2 is based on the forward hydraulic pressure PD, but the forward hydraulic pressure PD is the same pressure as the modulator pressure PLPM. Each of Psl2 increases / decreases as the modulator pressure PLPM increases / decreases.

  FIG. 4 is a diagram illustrating the switching operation of the sequence valve 88 when the second clutch C2 is in the released state in the hydraulic control circuit 46 of FIG. As described above, the modulator pressure PLPM increases as the engine rotational speed Ne increases. However, when the engine rotational speed Ne is higher than the predetermined saturation rotational speed Neh, the modulator pressure PLPM is saturated. It becomes a constant value. In the present embodiment, the modulator pressure PLPM is saturated at 1.8 (MPa) when the engine speed Ne is in a high speed state where the engine speed Ne is equal to or higher than a predetermined saturation speed Neh. When the second clutch C2 is in the disengaged state, the SL2 output pressure Psl2 is 0 (MPa) regardless of the modulator pressure PLPM, so the thrust by the SL2 output pressure Psl2 is zero. The thrust toward the normal state due to the modulator pressure PLPM is proportional to the modulator pressure PLPM. The thrust in the normal state direction by the spring 88sp is constant regardless of the modulator pressure PLPM. The thrust toward the fail-safe state due to the SLP output pressure Pslp increases / decreases as the modulator pressure PLPM increases / decreases, and is proportional, for example, as shown in FIG. In FIG. 4, the thrust by the SLP output pressure Pslp, that is, the thrust DFf, shows the case where the maximum SLP output pressure Pslp by the primary solenoid valve SLP is output. As shown in FIG. 4, when the engine speed Ne is in a low speed state where the engine speed Ne is less than a predetermined saturation speed Neh, the thrust valve DFf becomes larger than the thrust force DFn and the sequence valve 88 is unintentionally in a fail-safe state. There is an erroneous switching area. In FIG. 4, the hatched portion is an erroneous switching area. In the case of the present embodiment, when the modulator pressure PLPM exceeds 0.8 (MPa) and the maximum SLP output pressure Pslp is output, erroneous switching occurs in the sequence valve 88.

  When erroneous switching occurs in the sequence valve 88, for example, the following situation occurs and drivability deteriorates. If the forward hydraulic pressure PD is supplied to the hydraulic actuator C2a of the second clutch C2 via the sequence valve 88, the second clutch C2 may be suddenly engaged. Further, since the C1 control pressure Pc1 supplied to the hydraulic actuator of the first clutch C1 becomes the drain pressure, the first clutch C1 may be in a released state. Further, the SLG output pressure Pslg output from the D1 solenoid valve SLG is input to the primary pressure control valve 84 and the LU pressure control valve 94 as the output pressure Pseq1 of the sequence valve 88, whereby the primary control pressure Ppri is reduced. The lockup clutch LU may be released.

  FIG. 5 is a diagram for explaining the switching operation of the sequence valve 88 when the second clutch C2 is engaged in the hydraulic control circuit 46 of FIG. FIG. 5 is substantially the same as FIG. 4 except that the thrust in the normal state direction by the SL2 output pressure Psl2 is applied. Therefore, it demonstrates centering on a different part and abbreviate | omits description suitably about the part which is common in FIG. When the second clutch C2 is in the engaged state, the SL2 output pressure Psl2 is proportional to the modulator pressure PLPM, for example, but the SL2 output pressure Psl2 is related to the saturation of the second clutch C2 in order to suppress fuel consumption deterioration. It is controlled not to become larger than the combined pressure. In this embodiment, the SL2 output pressure Psl2 becomes the saturation engagement pressure of the second clutch C2 when the modulator pressure PLPM is 1.3 (MPa), and the SL2 output pressure Psl2 does not increase even when the modulator pressure PLPM increases further. Saturation engagement pressure is maintained. As shown in FIG. 5, there is no erroneous switching region where the thrust DFf is larger than the thrust DFn. Therefore, even if the maximum SLP output pressure Pslp is output by the primary solenoid valve SLP, there is no erroneous switching in which the sequence valve 88 unintentionally enters the fail-safe state. In FIG. 5, the relationship between the thrust DFn and the modulator pressure PLPM in FIG. 4 is indicated by a one-dot chain line for reference.

  FIG. 6 is an example of a method for setting the upper limit guard pressure Pslpg of the SLP output pressure Pslp in accordance with the operating state of the second clutch C2 in the hydraulic control circuit 46 of FIG.

  As shown in FIG. 6, in the high rotation state where the engine rotation speed Ne is equal to or higher than the predetermined saturation rotation speed Neh and the modulator pressure PLPM is saturated, regardless of the operation state of the second clutch C2. The first predetermined value Pslpg1 (MPa) is uniformly set as the upper limit guard pressure Pslpg so that erroneous switching does not occur in the sequence valve 88 even if the SL2 output pressure Psl2 is 0 (MPa). The upper limit guard pressure Pslpg is not controlled. For example, in this embodiment, the modulator pressure PLPM can ensure the saturation modulator pressure PLPMmax (= 1.8 (MPa)) because the engine rotation speed Ne in FIG. 4 is in a high rotation state where the engine rotation speed Ne is equal to or higher than the saturation rotation speed Neh. In some cases, the thrust in the fail-safe state direction by the upper limit guard pressure Pslpg is greater than the thrust in the normal state direction by the total value DFs of the spring load SP and the thrust force by the modulator pressure PLPM at the saturation modulator pressure PLPMmax. Is set to be small. That is, the first predetermined value Pslpg1 is the sum of the spring load pressure SPld (= spring load SP / pressure receiving area AL3) and the area ratio modulator pressure PLPMr (MPa) (= modulator pressure PLPM × pressure receiving area AL1 / pressure receiving area AL3). Set to When the engine rotational speed Ne is equal to or higher than the saturation rotational speed Neh and the modulator pressure PLPM is the saturated modulator pressure PLPMmax, the area ratio modulator pressure PLPMr has a pressure receiving area AL1 / pressure receiving area AL3 that is a ratio of the pressure receiving area to the saturated modulator pressure PLPMmax. It will be multiplied. Thereby, regardless of the operating state of the second clutch C2, erroneous switching that causes the sequence valve 88 to unintentionally enter the fail-safe state does not occur. When pressure receiving area AL1 and pressure receiving area AL3 are the same, area ratio modulator pressure PLPMr is the same as saturation modulator pressure PLPMmax.

  As shown in FIG. 6, in a low rotation state where the engine rotation speed Ne is less than a predetermined saturation rotation speed Neh and the modulator pressure PLPM is not saturated, depending on the operation state of the second clutch C2. An upper limit guard pressure Pslpg for the SLP output pressure Pslp is set.

  When the operation state of the second clutch C2 is a complete release and release transition, the upper limit guard pressure Pslpg is calculated from the spring load pressure SPld (= spring load SP / pressure receiving area AL3) by the spring 88sp and the PLPM pressure map. Set to the sum of the hydraulic pressure. Thus, the thrust in the fail-safe state direction by the SLP output pressure Pslp limited by the upper limit guard pressure Pslpg is greater than the thrust in the normal state direction by the sum of the spring load pressure SPld and the oil pressure calculated by the PLPM pressure map. Is made smaller. The PLPM pressure map shows the relationship between the engine rotational speed Ne and the hydraulic oil temperature Tho and the area ratio modulator pressure PLPMr obtained by multiplying the modulator pressure PLPM by the pressure receiving area AL1 / the pressure receiving area AL3. It is a representation. As the engine rotation speed Ne is lower and the hydraulic oil temperature Th is higher, the modulator pressure PLPM decreases due to the deterioration of the hydraulic oil flow rate balance. Therefore, by calculating the area ratio modulator pressure PLPMr from the PLPM pressure map defined in advance using the engine speed Ne and the hydraulic oil temperature Tho as parameters, erroneous switching of the sequence valve 88 does not occur even when the flow rate balance deteriorates. It is a thing.

  When the operating state of the second clutch C2 is in an engagement transition and before the engagement of the clutch-to-clutch switching control is started, the upper limit guard pressure Pslpg is the spring load pressure SPld (= spring load SP / pressure receiving area) by the spring 88sp. AL3) and the hydraulic pressure calculated by the PLPM pressure map. When the operation state of the second clutch C2 is an engagement transition and the clutch-to-clutch switching control is started after the start of engagement, the upper limit guard pressure Pslpg is equal to the spring load pressure SPld (= spring load SP / pressure receiving area) by the spring 88sp. AL3) and the hydraulic pressure calculated by the PLPM pressure map are added to an effective clutch pack end pressure Ppeeff (described later) of the second clutch C2. Here, the clutch pack end pressure Ppe is a hydraulic pressure before and after the hydraulic friction engagement device starts to have an engagement torque, for example, a hydraulic pressure just before the engagement torque is started. The clutch pack end pressure Ppe is measured in advance by measuring an SL2 output pressure Psl2 that is an engagement pressure of the second clutch C2 at an inertia phase start time tiner described later using the engine speed Ne and the hydraulic oil temperature Tho as parameters. You can ask for it. The effective clutch pack end pressure Ppeeff is obtained by multiplying the clutch pack end pressure Ppe by a pressure receiving area AL2 / pressure receiving area AL3 which is a ratio of the pressure receiving area. By preparing in advance a clutch pack end pressure map representing the relationship between the engine rotational speed Ne and the hydraulic oil temperature Tho and the effective clutch pack end pressure Ppeeff, the engine rotational speed Ne and the hydraulic oil temperature Tho. The effective clutch pack end pressure Ppeeff can be calculated using as a parameter. When the operating state of the second clutch C2 is garage engagement, the upper limit guard pressure Pslpg is the oil pressure calculated by the spring load pressure SPld (= spring load SP / pressure receiving area AL3) by the spring 88sp and the PLPM pressure map. Is set to the sum of

  When the operating state of the second clutch C2 is completely engaged, the SLP output pressure Pslp is set to the second value so that erroneous switching does not occur in the sequence valve 88 even if the SL2 output pressure Psl2 is 0 (MPa). The predetermined value Pslpg2 (MPa) is uniformly set as the upper limit guard pressure Pslpg, and is controlled so as not to exceed the upper limit guard pressure Pslpg. For example, in the present embodiment, the second predetermined value Pslpg2 can be the same value as the first predetermined value Pslppg1 described above.

  FIG. 7 is a time chart when the upper limit guard pressure Pslpg of the SLP output pressure Pslp corresponding to the operating state of the second clutch C2 is set in the hydraulic control circuit 46 of FIG. From time t0 to time t1, the operating state of the second clutch C2 is completely released, and for example, the vehicle 10 is in a low-speed forward traveling state. From time t1 to time t2, the operating state of the second clutch C2 is an engagement transition. For example, the vehicle 10 is in a transition state from low speed forward traveling to medium high speed forward traveling. At this time, clutch-to-clutch switching control for releasing the first clutch C1 and engaging the second clutch C2 is executed, and the engagement of the second clutch C2 is started at the inertia phase start time tiner. And As described above with reference to FIG. 6, the upper limit guard pressure Pslpg in the period from the inertia phase start time tiner to the time t2 after the second clutch C2 starts to be engaged is before the second clutch C2 starts to be engaged. Is higher by the effective clutch pack end pressure Ppeeff of the second clutch C2 than the period from the time t1 to the inertia phase start time tiner. If this is after the inertia phase start time tiner, the thrust force that brings the sequence valve 88 to the normal state increases by the amount obtained by multiplying the clutch pack end pressure Ppe by the pressure receiving area AL2. Therefore, even if the upper limit guard pressure Pslpg is increased by the effective clutch pack end pressure Ppeeff, the thrust that makes the sequence valve 88 fail-safe is increased only by the same amount as the thrust that makes the normal state, and the sequence valve 88 This is because there is no possibility of erroneous switching. Here, the inertia phase is an input that is an input rotating member in a shift period by switching from the first power transmission path PT1 to the second power transmission path PT2, that is, switching from the gear travel mode to the belt travel mode. This is a period during which the rotation speed of the shaft 22 changes. From time t2 to time t3, the operating state of the second clutch C2 is completely engaged, and for example, the vehicle 10 is in a middle-high speed forward traveling state. From time t3 to time t4, the operating state of the second clutch C2 is a disengagement transition. For example, the vehicle 10 is in a transition state from a medium-high speed forward travel to a low speed forward travel. After time t4, the operating state of the second clutch C2 is completely released. The C1 control valve Pc1 is regulated by controlling the C1 solenoid valve SL1 so that the SL1 output pressure Psl1 of the C1 solenoid valve SL1 becomes the C1 command oil pressure Pc1req which is the command pressure of the electronic control device 100. As a result, the connection / disconnection state of the first clutch C1 is controlled. The C2 solenoid valve SL2 is controlled so that the SL2 output pressure Psl2 of the C2 solenoid valve SL2 becomes the C2 command oil pressure Pc2req that is the command pressure of the electronic control device 100, thereby adjusting the C2 control pressure Pc2. As a result, the connection / disconnection state of the second clutch C2 is controlled. Further, the upper limit guard pressure Pslpg is set according to the operating state of the second clutch C2, the engine rotational speed Ne, and the hydraulic oil temperature Th as described with reference to FIG.

  Here, the control function of the electronic control device 100 for various controls in the vehicle 10 and the main part of the control system will be described with reference to FIG. The electronic control device 100 includes a switching prevention determination unit 102, a rotation speed determination unit 104, a first operation state determination unit 106, a second operation state determination unit 108, a guard pressure setting unit 110, and a guard pressure execution unit 120.

  The switching prevention determination unit 102 determines whether or not a request for preventing erroneous switching to the fail safe state of the sequence valve 88 is made. For example, when the shift operation of the upshift of the continuously variable transmission 24 and the engagement operation of the second clutch C2 overlap, the supply of hydraulic oil to the hydraulic actuator 58c of the primary pulley 58 and the second clutch C2 This is a state where the engagement flow overlaps and the flow rate balance of the hydraulic oil tends to deteriorate. Further, when the shift lever 98 is operated from the N operation position to the D operation position and the first clutch C1 is switched from the released state to the engaged state, the accumulator 96 pressure accumulation operation and the first clutch C1 are engaged. The combined operation overlaps and the flow rate balance of the hydraulic oil is likely to deteriorate. As described above, when the flow rate balance of the hydraulic oil is likely to deteriorate, the switching prevention determination unit 102 determines that a request for preventing erroneous switching to the fail-safe state of the sequence valve 88 is made. The switching prevention determination unit 102 outputs a command signal to the rotational speed determination unit 104 when it is determined that a request for prevention of erroneous switching is made, and the command signal is transmitted to the guard pressure setting unit when it is determined that a request for prevention of erroneous switching is not made. Output to 110

  When the command signal is input from the switching prevention determination unit 102, the rotation speed determination unit 104 is based on the engine rotation speed Ne input from the rotation speed sensor 64 and the oil temperature Tho of hydraulic oil input from the oil temperature sensor 78. Thus, it is determined whether or not the engine rotational speed Ne is equal to or higher than a predetermined rotational speed, for example, a predetermined saturation rotational speed Neh, with reference to a preset map using the oil temperature Tho of the hydraulic oil as a parameter. When the engine rotation speed Ne is equal to or higher than a predetermined saturation rotation speed Neh, the modulator pressure PLPM is saturated. The saturation rotation speed Neh corresponds to the “predetermined rotation speed” in the present invention. When the rotational speed determination unit 104 determines that the engine rotational speed Ne is equal to or higher than the saturation rotational speed Neh determined by the map, the rotational speed determination unit 104 outputs a command signal to the guard pressure setting unit 110, and the engine rotational speed Ne is saturated by the map. If it is determined that the rotational speed is less than Neh, a command signal is output to the first operating state determination unit 106.

  When the command signal is input from the rotation speed determination unit 104, the first operation state determination unit 106 determines the operation state of the second clutch C2 (full release, engagement transient, release transient, and complete engagement states). judge. The first operating state determining unit 106 outputs a command signal to the second operating state determining unit 108 after determining that the operating state of the second clutch C2 is an engagement transition. The first operating state determining unit 106 outputs a command signal to the guard pressure setting unit 110 after determining that the operating state of the second clutch C2 is any of complete release, release transition, and complete engagement.

  Here, the control function of the 1st operation state determination part 106 is demonstrated in detail based on FIG. FIG. 8 shows the relationship between the operating state of the second clutch C2 represented by four states of complete release, complete engagement, release transition, and engagement transition, and transition conditions for determining switching of the operation state. FIG.

  The first operating state determination unit 106 performs complete release, complete engagement, and release based on the state of hydraulic control for the second clutch C2 and the state of a differential rotational speed ΔNc2 (rpm) described later in the second clutch C2. It is determined whether or not the transition condition of the operating state of the second clutch C2 represented by the four states of transition and engagement transition is satisfied. The first operating state determination unit 106 determines the operating state of the second clutch C2 based on the determination result of whether or not the transition condition is satisfied. The first operating state determination unit 106 acquires the state of hydraulic control for the second clutch C2 based on the hydraulic control command signal Scbd. The first operating state determination unit 106 calculates a differential rotational speed ΔNc2 (= Nsec−Nout) in the second clutch C2 based on the secondary rotational speed Nsec and the output shaft rotational speed Nout.

  The state of the hydraulic control for the second clutch C2 is the tendency of the hydraulic pressure supplied to the hydraulic actuator C2a of the second clutch C2 being raised or lowered, and / or the second clutch C2. This is the state of the command pressure in the hydraulic control for. The state of the differential rotation speed ΔNc2 is the actual state of how the second clutch C2 is actually operating.

  The hydraulic control for the second clutch C2 described above is an engagement control for engaging the second clutch C2, or a release control for releasing the second clutch C2. In this embodiment, the hydraulic control for engaging the second clutch C2 is called C2 engagement hydraulic control, and the hydraulic control for releasing the second clutch C2 is called C2 release hydraulic control. In the present embodiment, the hydraulic control command signal Scbd, which is the command pressure in the hydraulic control for the second clutch C2, is referred to as the C2 command hydraulic pressure.

  The C2 engagement hydraulic pressure control is assumed to be C2 engagement hydraulic pressure control in switching from the state where the first power transmission path PT1 is formed to the state where the second power transmission path PT2 is formed. Control to switch from the state in which the first power transmission path PT1 is formed to the state in which the second power transmission path PT2 is formed is a clutch toe that releases the first clutch C1 and engages the second clutch C2. This is switching control for switching the driving mode from the gear driving mode to the belt driving mode, which is executed in the clutch switching control. Alternatively, the C2 engagement hydraulic control is assumed to be C2 engagement hydraulic control in control for switching from the neutral state of the power transmission device 16 to the state in which the second power transmission path PT2 is formed. The control to switch the power transmission device 16 from the neutral state to the state in which the second power transmission path PT2 is formed is performed in accordance with the garage operation in which the shift lever 98 is operated from the N operation position to the D operation position in the belt travel mode. This is garage engagement control for engaging the second clutch C2.

  The C2 release hydraulic control is assumed to be C2 release hydraulic control in control for switching from the state in which the second power transmission path PT2 is formed to the state in which the first power transmission path PT1 is formed. The control to switch from the state where the second power transmission path PT2 is formed to the state where the first power transmission path PT1 is formed is a clutch toe that releases the second clutch C2 and engages the first clutch C1. This is switching control for switching the running mode from the belt running mode to the gear running mode, which is executed in the clutch switching control. Alternatively, the C2 release hydraulic control is assumed to be C2 release hydraulic control in control for switching from the state in which the second power transmission path PT2 is formed to the neutral state of the power transmission device 16. The control for switching from the state in which the second power transmission path PT2 is formed to the neutral state of the power transmission device 16 is performed in accordance with the garage operation in which the shift lever 98 is operated from the D operation position to the N operation position in the belt traveling mode. This is garage release control for releasing the second clutch C2.

  When [Condition 1] as the transition condition is satisfied when the second clutch C2 is in a completely released state or a released transition state, it is determined that the second clutch C2 has been switched to the engaged transition state. Is done. In the process of engaging the second clutch C2 with the C2 control pressure Pc2, the actual pressure (hereinafter referred to as C2 actual pressure) of the C2 control pressure Pc2 with respect to the command hydraulic pressure to the C2 control pressure Pc2 (hereinafter referred to as C2 instruction hydraulic pressure). Response delay occurs in the output of. Therefore, the relationship of “C2 command hydraulic pressure> C2 actual pressure” can be guaranteed during the response delay period.

  Therefore, [Condition 1] is that the C2 engagement hydraulic pressure control is operating, that is, the hydraulic control command signal Scbd for the C2 engagement hydraulic pressure control is output.

  The first operating state determination unit 106 determines that the second clutch C2 is in the engagement transition state when [Condition 1] is satisfied when the second clutch C2 is in the completely released state or the release transition state. It is determined that the state has been switched.

  When [Condition 2] as a transition condition is satisfied when the second clutch C2 is in a fully engaged state, a disengagement transition state or an engagement transition state, the second clutch C2 is completely disengaged. It is determined that the state has been switched. In addition to the fact that the C2 release hydraulic pressure control is operated, if the C2 command hydraulic pressure in the C2 release hydraulic pressure control is equal to or lower than the value for completely releasing the second clutch C2, and the differential rotational speed ΔNc2 is increased, the second Even if it is determined that the clutch C2 is switched to the fully released state, a problem hardly occurs. However, this determination method is effective in a region where the input shaft rotational speed Nin (= primary rotational speed Npri = turbine rotational speed Nt) that is the input rotational speed of the continuously variable transmission 24 is high. In a region where the input shaft rotational speed Nin is low, there is a possibility that the differential rotational speed ΔNc2 becomes difficult to detect. For example, in the switching control from the belt travel mode to the gear travel mode when the vehicle 10 is stopped, the first clutch C1 is engaged after the second clutch C2 is released and the input shaft rotational speed Nin cannot be increased. The differential rotational speed ΔNc2 cannot be detected. In the region where the input shaft rotational speed Nin is low, the second clutch C2 is switched to the fully released state because the C2 command hydraulic pressure in the C2 release hydraulic pressure control has decreased to a value that can guarantee the complete release of the second clutch C2. It is determined that Further, in addition to the above-described determination method, it may be determined that the second clutch C2 has been switched to the fully released state when the neutral state of the power transmission device 16 is established. For example, when the shift lever 98 is operated to the N operation position, the original pressure of the clutch hydraulic pressure supplied to the second clutch C2 is not supplied from the manual valve 82, and the clutch hydraulic pressure of the second clutch C2 is reduced. As a result, the second clutch C2 is released. When the C2 release hydraulic control is activated by operating the shift lever 98 to the N operation position, for example, the transition condition when the above-described C2 release hydraulic control is activated is used.

  Therefore, [Condition 2] is that the C2 release hydraulic pressure control is operating, that is, the hydraulic control command signal Scbd for C2 release hydraulic pressure control is output, and the input shaft rotational speed Nin is predetermined. Any of the conditions that the rotational speed Ninf (rpm) or higher, the C2 command hydraulic pressure in the C2 release hydraulic pressure control is equal to or lower than the first predetermined command pressure, and the differential rotational speed ΔNc2 is greater than the first predetermined differential rotational speed ΔNc2f1 (rpm) That is, the predetermined time TM1 or more is established. Alternatively, in the [Condition 2], in addition to the C2 release oil pressure control being operated, the input shaft rotation speed Nin is less than the predetermined rotation speed Ninf, and the C2 release oil pressure in the C2 release oil pressure control is the second predetermined instruction pressure. Each of the following conditions is established for a predetermined time TM2 or more. Alternatively, [Condition 2] is that the neutral state of the power transmission device 16 is confirmed, for example, that a predetermined time TM3 or more has elapsed since the shift lever 98 was operated to the N operation position.

  The predetermined rotational speed Ninf is a lower limit rotational speed in a region of a predetermined input shaft rotational speed Nin that can detect, for example, a differential rotational speed ΔNc2 for use in determining whether or not the second clutch C2 is completely released. . For example, the first predetermined command pressure, the second predetermined command pressure, the first predetermined differential rotational speed ΔNc2f1, the predetermined time TM1, the predetermined time TM2, and the predetermined time TM3 are each in a state in which the second clutch C2 is completely released, for example. This is a predetermined threshold value for determining whether or not. In particular, the second predetermined command pressure is lower than the first predetermined command pressure, and for example, it is possible to ensure that the second clutch C2 is in a completely released state, for example, an upper limit of a predetermined C2 command hydraulic range. Value.

  The first operating state determination unit 106 determines that when any of the above [Condition 2] is satisfied when the second clutch C2 is in a fully engaged state, a disengagement transition state, or an engagement transition state. Determines that the second clutch C2 has been switched to the fully released state.

  When [condition 3] as a transition condition is satisfied when the second clutch C2 is in a fully released state, a released transition state, or an engaged transition state, the second clutch C2 is in a fully engaged state. It is determined that the state has been switched. In addition to the C2 engagement oil pressure control being activated, the C2 command oil pressure in the C2 engagement oil pressure control is greater than or equal to the value necessary for the engagement of the second clutch C2, and the differential rotational speed ΔNc2 is reduced. For example, even if it is determined that the second clutch C2 is switched to the fully engaged state, a problem hardly occurs. That is, when the change direction of the C2 command hydraulic pressure in the hydraulic control for the second clutch C2 is the direction in which the second clutch C2 is engaged, and the actual state of the second clutch C2 can be regarded as a fully engaged state. It is determined that the second clutch C2 has been switched to the fully engaged state. In addition to the above determination method, the hydraulic control for the second clutch C2 is shifted from the C2 engagement hydraulic control to the C2 steady hydraulic control that is a hydraulic control for maintaining the second clutch C2 in a fully engaged state. At this point, it may be determined that the second clutch C2 has been switched to the fully engaged state.

  Therefore, in [Condition 3], in addition to the C2 engagement oil pressure control being operated, the C2 instruction oil pressure in the C2 engagement oil pressure control is greater than or equal to the third predetermined instruction pressure, and the differential rotational speed ΔNc2 is the second predetermined All of the conditions that the rotational speed is smaller than the differential rotational speed ΔNc2f2 (rpm) are satisfied for the predetermined time TM4 or more. Alternatively, [Condition 3] is that the hydraulic control for the second clutch C2 is shifted from the C2 engagement hydraulic control to the C2 steady hydraulic control. The third predetermined command pressure, the second predetermined differential rotation speed ΔNc2f2, and the predetermined time TM4 are predetermined threshold values for determining, for example, that the second clutch C2 is in a fully engaged state.

  When the second clutch C2 is in a fully released state, a release transition state, or an engagement transition state, the first operating state determination unit 106 is in a case where any of the above [Condition 3] is satisfied. Determines that the second clutch C2 has been switched to the fully engaged state.

  When [Condition 4] as a transition condition is established when the second clutch C2 is in a fully engaged state, it is determined that the second clutch C2 has been switched to a disengagement transient state. If the differential rotational speed ΔNc2 is increased in addition to the C2 release hydraulic pressure control being activated, it is difficult to cause a problem even if it is determined that the second clutch C2 is switched to the release transition state. The increase in the differential rotational speed ΔNc2 is conditioned on the failure of the second clutch C2 due to, for example, a failure of the C2 solenoid valve SL2 that regulates the C2 control pressure Pc2 supplied to the hydraulic actuator C2a of the second clutch C2. This is because the actual situation may remain completely engaged.

  Therefore, [Condition 4] is that the condition that the differential rotational speed ΔNc2 is equal to or higher than the third predetermined differential rotational speed ΔNc2f3 (rpm) is satisfied in addition to the C2 release hydraulic pressure control being operated. is there. The third predetermined differential rotation speed ΔNc2f3 and the predetermined time TM5 are predetermined threshold values for determining, for example, that the second clutch C2 is in a disengagement transient state.

  The first operating state determination unit 106 determines that the second clutch C2 is switched to a disengagement transient state when the above [Condition 4] is satisfied when the second clutch C2 is in a fully engaged state. judge.

  If [Condition 5] as a transition condition is established when the second clutch C2 is in the transitional state of engagement, it is determined that the second clutch C2 has been switched to the state of transitional disengagement. Even when it is determined that the second clutch C2 has been switched to the release transition state when the C2 engagement hydraulic control is shifted to the C2 release hydraulic control, a problem hardly occurs.

  Therefore, [Condition 5] is that the C2 release hydraulic pressure control is operating.

  If the [Condition 5] is satisfied when the second clutch C2 is in the engagement transition state, the first operating state determination unit 106 determines that the second clutch C2 has been switched to the release transition state. judge.

  When the command signal is input from the first operation state determination unit 106, the second operation state determination unit 108 determines whether the operation state of the second clutch C2 is based on garage control or the operation state of the second clutch C2 is the clutch. In the toe clutch switching control, it is determined whether it is before the start of engagement, for example, before the start of the inertia phase. When the engagement of the second clutch C2 is started, an upshift is performed by shifting from the gear travel mode to the belt travel mode, and therefore the turbine rotational speed of the torque converter 20, that is, the input shaft rotational speed Nin (= primary rotational speed Npri). ) Decreases. By detecting the decrease in the input shaft rotation speed Nin, the electronic control unit 100 can determine whether the second clutch C2 is before or after the engagement is started. Further, the electronic control unit 100 determines whether or not the garage control is performed by detecting that the accelerator operation amount θacc is zero and the operation position POSsh is the D operation position or the R operation position. it can. After the above determination, the second operating state determination unit 108 outputs a command signal to the guard pressure setting unit 110.

  When the command signal is input from the switching prevention determination unit 102, the guard pressure setting unit 110 sets the upper limit guard pressure Pslpg to a guard invalid value, for example, a value exceeding the maximum SLP output pressure Pslp. When the command signal is input from the rotation speed determination unit 104, the guard pressure setting unit 110 sets the upper limit guard pressure Pslpg to the first predetermined value Pslpg1 described with reference to FIG. When the command signal is input from the first operation state determination unit 106 or the second operation state determination unit 108, the guard pressure setting unit 110 responds to the operation state of the second clutch C2 as described above with reference to FIG. To set the upper guard pressure Pslpg. After setting the upper limit guard pressure Pslpg, the guard pressure setting unit 110 outputs a command signal to the guard pressure execution unit 120.

  When a command signal is input from the guard pressure setting unit 110, the guard pressure execution unit 120 validates the upper limit guard pressure Pslpg set by the guard pressure setting unit 110, and the SLP output pressure Pslp does not exceed the upper limit guard pressure Pslpg. To control.

  FIG. 9 is a flowchart illustrating a control operation for setting the upper limit guard pressure Pslpg of the SLP output pressure Pslp in accordance with the operating state of the second clutch C2 in the hydraulic control circuit 46 of FIG. The flowchart of FIG. 9 is executed by repeating the start every predetermined time (for example, several ms) in the electronic control device 100, for example.

  First, in step S10 corresponding to the switching prevention determination unit 102, it is determined whether or not a request for preventing erroneous switching to the fail safe position of the sequence valve 88 is made. If the determination in step S10 is negative, step S20 is executed. If the determination in step S10 is affirmative, step S30 is executed.

  In step S20 corresponding to the guard pressure setting unit 110, the upper limit guard pressure Pslpg is set to a guard invalid value. Then step S150 is executed.

  In step S30 corresponding to the rotation speed determination unit 104, it is determined whether or not the engine rotation speed Ne is equal to or higher than a predetermined saturation rotation speed Neh. If the determination in step S30 is affirmative, step S40 is executed. If the determination in step S30 is negative, step S50 is executed.

  In step S40 corresponding to the guard pressure setting unit 110, the upper limit guard pressure Pslpg is set to the first predetermined value Pslpg1. Then step S150 is executed.

  In step S50 corresponding to the first operating state determination unit 106, the operating state of the second clutch C2 (completely released, engaged transient, released transient, fully engaged state) is determined. Then, Step S60 to Step S100 are executed.

  In steps S60 to S100 corresponding to the second operating state determination unit 108 and the guard pressure setting unit 110, the upper limit guard pressure Pslpg is set according to the operating state of the second clutch C2, as described above with reference to FIG. The When the operating state of the second clutch C2 is completely released, the upper limit guard pressure Pslpg is set in step S70 corresponding to the guard pressure setting unit 110, and step S150 is executed. When the operation state of the second clutch C2 is an engagement transition, the upper limit guard pressure Pslpg is set in step S80 corresponding to the second operation state determination unit 108 and the guard pressure setting unit 110, and step S150 is executed. The When the operation state of the second clutch C2 is a release transition, the upper limit guard pressure Pslpg is set in step S90 corresponding to the guard pressure setting unit 110, and step S150 is executed. When the operating state of the second clutch C2 is completely engaged, the upper limit guard pressure Pslpg is set in step S100 corresponding to the guard pressure setting unit 110, and step S150 is executed.

  In step S150 corresponding to the guard pressure execution unit 120, the upper limit guard pressure Pslpg set in any one of steps S20, S40, S70, S80, S90, and S100 is validated, and the SLP output pressure Pslp is set to the upper limit guard. The pressure is controlled so as not to exceed the pressure Pslpg. And return.

  FIG. 10 is an example of a partial flowchart for explaining the control operation for determining the operating state of the second clutch C2 in step S50 of the flowchart of FIG.

  In step S52 corresponding to the first operating state determination unit 106, the hydraulic control state (engagement control / release control, C2 instruction hydraulic pressure) for the second clutch C2, the input shaft rotation speed Nin, the secondary rotation speed Nsec, and the output shaft rotation The speed Nout and the operation position POSsh are acquired, and the differential rotation speed ΔNc2 is calculated. Then step S54 is executed.

  In step S54 corresponding to the first operating state determination unit 106, the operating state of the second clutch C2 is determined. That is, as described with reference to FIG. 8, it is determined whether the operation state of the second clutch C2 is any of the four states represented by complete release, engagement transition, release transient, and complete engagement. Is done. Then, step S50 ends.

  FIG. 11 is an example of a partial flowchart illustrating a control operation related to the setting of the upper limit guard pressure Pslpg during the engagement transition of the second clutch C2 in step S80 of the flowchart of FIG.

  In step S82 corresponding to the second operating state determination unit 108, the operating state of the second clutch C2 is garage control, or the operating state of the second clutch C2 is before the start of engagement in the clutch-to-clutch switching control. It is determined whether or not there is any. If the determination in step S82 is affirmative, step S84 is executed. If the determination in step S82 is negative, step S86 is executed.

  In step S84 corresponding to the guard setting unit 110, the upper limit guard pressure Pslpg is set, and step S150 is executed.

  In step S86 corresponding to the guard setting unit 110, the upper limit guard pressure Pslpg is set, and step S150 is executed.

  According to the vehicle hydraulic control apparatus 150 of the present embodiment, the solenoid valve group SLgr includes the primary solenoid valve SLP, the lockup solenoid valve SLU, and the D1 solenoid valve SLG, and the solenoids included in the solenoid valve group SLgr. If any of the valves is abnormal, a sequence valve 88 is provided that switches the oil passage and exhibits a fail-safe function. The sequence valve 88 is supplied with the modulator pressure PLPM and the SL2 output pressure Psl2 as a thrust to be in a normal state, and is supplied with the SLP output pressure Pslp as a thrust in a fail-safe state. When the engine rotational speed Ne is lower than the predetermined saturation rotational speed Neh, the upper limit guard pressure Pslpg of the SLP output pressure Pslp is not switched to the fail-safe state due to malfunction of the sequence valve 88 according to the operating state of the second clutch C2. Thus, the hydraulic pressure is set to a predetermined value. Thus, since the upper limit guard pressure Pslpg of the SLP output pressure Pslp is provided according to the engine rotational speed Ne of the engine 12 and the operating state of the second clutch C2, the primary of the continuously variable transmission 24 by the SLP output pressure Pslp is provided. The suppression of the influence on the control of the pulley 58 and the suppression of the occurrence of erroneous switching of the sequence valve 88 are compatible. Further, when the control method of the present embodiment is implemented as a method of suppressing the occurrence of erroneous switching of the sequence valve 88, the fuel consumption deteriorates if a control method in which the modulator pressure PLPM is increased by idle-up control of the engine 12 is not adopted. Is avoided.

  According to the vehicle hydraulic control apparatus 150 of the present embodiment, the SLP output pressure that is set according to the operating state of the second clutch C2 when the engine rotational speed Ne of the engine 12 is less than a predetermined saturation rotational speed Neh. The upper limit guard pressure Pslpg of Pslp is determined according to the engine speed Ne of the engine 12 and the oil temperature Tho of the hydraulic oil. Since the modulator pressure PLPM that acts as a thrust to bring the sequence valve 88 into a normal state in this way depends on the engine speed Ne and the oil temperature Tho of the hydraulic oil, the SLP output pressure Pslp that acts as a thrust to bring the fail-safe state to these The upper limit guard pressure Pslpg is set according to the engine speed Ne and the hydraulic oil temperature Th. Thereby, even if the modulator pressure PLPM is reduced, the upper limit guard pressure Pslpg is also lowered accordingly, so that erroneous switching of the sequence valve 88 is suppressed.

  According to the vehicle hydraulic control apparatus 150 of the present embodiment, the biasing force of the spring 88sp is applied as a thrust that causes the spool valve element 88sv of the sequence valve 88 to be in a normal position. The predetermined hydraulic pressure set as the upper limit guard pressure Pslpg of the SLP output pressure Pslp is determined according to the engine rotational speed Ne and the hydraulic oil temperature Tho, and the spring load pressure SPld of the urging force of the spring 88sp and the modulator pressure. It is summed with PLPM. Even if the modulator pressure PLPM is not increased so much, the sequence valve 88 is in a normal state because of the urging force of the spring 88sp. Therefore, fuel consumption is improved by reducing the modulator pressure PLPM, and erroneous switching of the sequence valve 88 occurs. Suppression of both.

  According to the vehicle hydraulic control apparatus 150 of the present embodiment, when the engine rotational speed Ne is less than the saturation rotational speed Neh and the second clutch C2 is in the engagement transition state, the second clutch C2 is a clutch-to-clutch. The upper limit guard pressure Pslpg of the SLP output pressure Pslp is determined in advance so that the sequence valve 88 does not switch to the fail-safe state according to the engine rotational speed Ne and the hydraulic oil temperature Th when the switching control is started. The effective clutch pack end pressure Ppeeff of the second clutch C2 is added to the sum of the spring load pressure SPld of the urging force of the spring 88sp and the modulator pressure PLPM. Thereby, for example, the upper limit guard pressure Pslpg of the SLP output pressure Pslp is relaxed immediately after the start of engagement in the engagement transition state of the second clutch C2, so that the followability to the target gear ratio of the continuously variable transmission 24 is improved. Suppression of the occurrence of erroneous switching of the sequence valve 88 is made compatible.

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

  In the above-described embodiment, the modulator pressure Ne is determined based on whether or not the engine rotation speed Ne is equal to or higher than a predetermined saturation rotation speed Neh set according to the hydraulic oil temperature Tho in step S30 of the flowchart of FIG. Although it has been determined whether the PLPM is saturated, the present invention is not limited to this. For example, the predetermined saturation rotational speed Neh may be set to a constant value in advance regardless of the oil temperature Tho of the hydraulic oil.

  In the above-described embodiment, it is determined whether or not the engine rotation speed Ne is equal to or higher than the predetermined saturation rotation speed Neh in step S30 of the flowchart of FIG. For example, the upper limit guard pressure Pslpg may be set in steps S50 to S100 even if the determination in step S30 is not performed and the engine 12 is in a high rotation state.

  In the above-described embodiment, the first predetermined value Pslpg1 set as the upper limit guard pressure Pslpg when the engine rotation speed Ne is equal to or higher than the predetermined saturation rotation speed Neh, and the engine rotation speed Ne is less than the predetermined saturation rotation speed Neh. In this case, the second predetermined value Pslpg2 set as the upper limit guard pressure Pslpg when the operating state of the second clutch C2 is fully engaged is the same value, but is not limited thereto. The first predetermined value Pslpg1 and the second predetermined value Pslpg2 may be different values.

  The above are only examples of the present invention, and the present invention can be implemented in variously modified and improved modes based on the knowledge of those skilled in the art without departing from the spirit of the present invention.

10: Vehicle 12: Engine 14: Drive wheel 20: Torque converter 22: Input shaft (input rotating member)
24: continuously variable transmission 28: gear mechanism 30: output shaft (output rotating member)
44: Oil pump 46: Hydraulic control circuit 58: Primary pulley 60: Secondary pulley 80: Modulator valve 84: Primary pressure control valve 86: Secondary pressure control valve 88: Sequence valve 88sp: Spring 90: C1 control valve 92: S1B1 control valve 94: LU pressure control valve 100: Electronic control device 150: Vehicle hydraulic control device B1: First brake C1: First clutch C2: Second clutch LU: Lock-up clutch Ne: Engine rotation speed Neh: Saturation rotation Speed (predetermined rotation speed)
Plu: Lockup control pressure PLPM: Modulator pressure Poff: Off pressure Pon: On pressure Ppeeff: Effective clutch pack end pressure Ppri: Primary control pressure Psec: Secondary control pressure Psl2: SL2 output pressure (first engagement pressure)
Pslp: SLP output pressure (second engagement pressure)
Pslpg: Upper limit guard pressure Psls: SLS output pressure PT1: First power transmission path PT2: Second power transmission path SL1: C1 solenoid valve SL2: C2 solenoid valve SLG: D1 solenoid valve SLgr: Solenoid valve group SLP : Primary solenoid valve SLS: Secondary solenoid valve SLU: Lockup solenoid valve SPld: Spring load pressure (load pressure)
Th: Oil temperature

Claims (1)

An output rotating member connected to an engine, an output rotating member connected to a drive wheel, and the input rotating member and the output rotating member are connected to each other through a gear mechanism, which is a stepped transmission. A first power transmission path for transmitting; a second power transmission path for transmitting the output torque of the engine through the continuously variable transmission through the input rotating member and the output rotating member; and the first power. A first clutch provided in the transmission path and connecting / disconnecting the first power transmission path; and a second clutch provided in the second power transmission path and connecting / disconnecting the second power transmission path; An oil pump that is driven to rotate by the engine, and a modulator that generates a modulator pressure when hydraulic oil is supplied from the oil pump. And a C2 solenoid valve capable of supplying a first engagement pressure for operating the second clutch based on the modulator pressure, and a primary for operating a primary pulley of the continuously variable transmission based on the modulator pressure. When both the solenoid valve group including the primary solenoid valve capable of supplying the second engagement pressure capable of adjusting the control pressure and the solenoid valve included in the solenoid valve group are normal, the first oil The normal state in which the passage is formed and the fail-safe state in which the second oil passage is formed can be switched when any of the solenoid valves included in the solenoid valve group is abnormal, and the normal state is set. A biasing force of a spring is applied as a thrust, and the modulator pressure and the first engagement pressure are supplied. A sequence valve which said second engagement pressure is supplied as a thrust with Rusefu state, a hydraulic control system for a vehicle comprising,
When the engine rotation speed, which is the rotation speed of the engine, is less than a predetermined rotation speed and the second clutch is in an engagement transient state, the second clutch is after the start of clutch-to-clutch switching control engagement. The upper limit guard pressure of the second engagement pressure is determined in advance so that the sequence valve does not switch to the second state according to the engine speed and the oil temperature of the hydraulic oil. The vehicle hydraulic control apparatus, wherein the effective clutch pack end pressure of the second clutch is added to the sum of the load pressure of the urging force and the modulator pressure.

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