JP2964757B2 - Hydraulic control device for automatic transmission for vehicles - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a hydraulic control device for an automatic transmission for a vehicle, and more particularly to a control for lowering an engine torque and an engagement of a frictional engagement device in relation to a shift speed change by a hydraulic frictional engagement device. The present invention relates to a hydraulic control device of a type that implements change control for alleviating the rise of characteristics.

[0002]

2. Description of the Related Art There is known a vehicle having a fluid transmission with a direct coupling clutch and an automatic transmission in which a gear can be switched by engaging and disengaging a hydraulic friction engagement device. And in such a vehicle hydraulic control device,
In order to reduce a shift shock generated in connection with engagement of the hydraulic friction engagement device, an engine torque reduction control unit that reduces engine torque in accordance with a running state of the vehicle when changing gears is provided. In addition, in connection with the engine torque reduction control, for example, when the engine torque is restored, it has been proposed to change the engagement characteristics of the hydraulic friction engagement device to a gentler rise than usual.
For example, this is described in Japanese Patent Application Laid-Open No. 3-84259. According to such a hydraulic control device, the reduction of the engine torque and the change of the engagement characteristics of the engagement friction device are performed in parallel with the shift operation during the shift operation. There is the advantage of being lowered.

[0003]

In the conventional automatic transmission for a vehicle as described above, it is conceivable that the direct coupling clutch is released during the change of the shift speed in order to further suppress the shift shock. For this purpose, the hydraulic control device of the automatic transmission outputs a switching signal pressure for switching to any one of the engagement mode, the rapid release mode, the first release mode, and the second release mode of the direct coupling clutch according to the traveling state of the vehicle. In the case where the switching electromagnetic solenoid valve device is provided, if the traveling state of the vehicle is in the engagement region of the direct-coupled clutch, the switching from the engagement mode to the first or second release mode is performed by the switching signal pressure when the hydraulic engagement friction device is switched. It is necessary to switch to However, in this case, in order to control the switching of the direct coupling clutch and the change of the engagement characteristic, respectively, the switching electromagnetic solenoid valve device for controlling the direct coupling clutch and the engagement characteristic of the hydraulic friction engagement device are changed. And the electromagnetic switching valves must be provided independently of each other, which complicates the hydraulic circuit and increases the cost of the hydraulic control device.

[0004] The present invention has been made in view of the above circumstances, and an object thereof is to appropriately apply a shift shock without complicating the hydraulic circuit or making the hydraulic control device expensive. It is an object of the present invention to provide a hydraulic control device for an automatic transmission for a vehicle which can be reduced.

[0005]

SUMMARY OF THE INVENTION In order to achieve the above object, the gist of the present invention is, as shown in FIG. 1 for explaining the gist of the invention, a fluid transmission with a direct-coupled clutch, and a hydraulic friction device. In a vehicle including an automatic transmission in which a shift speed is switched by engagement and disengagement of an engagement device, an engine torque reduction control unit that reduces engine torque during a shift speed change; Engagement mode, normally used to release the direct coupling clutch
First release mode, release the directly connected clutch
A second release mode used during the change of gear stage for
And the directly connected clutch thereof in the first release mode and the
Quick release mode for releasing faster than 2 release mode
And a switching electromagnetic solenoid valve device that outputs a switching signal pressure for switching to any of the gear positions.
A hydraulic control device of a type in which the direct coupling clutch is released and the engagement characteristics of the hydraulic friction engagement device are changed to a more gradual rise than usual, and (a) engine torque reduction control by the engine torque reduction control means. (B) during the period in which it is determined that the engine torque reduction control is being executed by the engine torque reduction determination means, the switching electromagnetic solenoid valve device performs the second release. Second release mode switching command means for outputting a switching signal pressure for setting the mode, and (c) a first position and a second position for making engagement characteristics of the hydraulic friction engagement device gentler than the first position. And a switching valve for switching to the second release mode from the switching electromagnetic solenoid valve device. Lies in including, an engagement characteristic switching valve the valve element is positioned in a second location.

[0006]

In this manner, during the period in which the engine torque reduction determining means determines that the engine torque reduction control is being executed, the second release mode switching instruction means switches the electromagnetic solenoid valve device for switching from the second release mode to the second release mode. A switching signal pressure for setting the mode is output. While the switching signal pressure is being output, the valve element of the engagement characteristic switching valve is located at the second position where the engagement characteristic of the hydraulic friction engagement device gradually rises from the first position. .

[0007]

According to the present invention, the engagement characteristic for changing the engagement characteristic of the hydraulic friction engagement device in response to the switching signal pressure for switching to the second release mode from the switching electromagnetic solenoid valve device. Since the switching valve is provided, the direct coupling clutch is released in synchronization with the start of the engine torque reduction control executed during the shift speed change, and the engagement characteristics of the hydraulic friction engagement device are gradually increased. Is changed to. As described above, the switching electromagnetic solenoid valve device for controlling the direct coupling clutch can be used also for changing the engagement characteristics, so that it is not necessary to separately provide a new electromagnetic switching valve for changing the engagement characteristics. The shift shock can be suitably reduced by an inexpensive hydraulic control device having a simple circuit.

[0008]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS An embodiment of the present invention will be described below in detail with reference to the drawings.

FIG. 2 is a skeleton diagram of a lateral transaxle for an FF vehicle to which the control device according to one embodiment of the present invention is applied, and FIG. 3 is a block diagram showing a configuration example of the control device. 2, the power of the engine 10 is supplied by a fluid coupling 12 with a lock-up clutch, a forward / reverse switching device 14, a belt-type continuously variable transmission (hereinafter, referred to as CVT).
The transmission is transmitted to wheels 26 connected to a drive shaft 24 via a transmission 16, an auxiliary transmission 18, a reduction gear device 20, and a differential gear device 22.

The fluid coupling 12 is connected to the engine 1
Pump wheel 30 connected to crankshaft 28
A turbine impeller 32 rotated by oil from the pump impeller 30, an output shaft 34 connected to the turbine impeller 32 so as not to rotate relative thereto, and an output shaft 34 via a damper 36. A lock-up clutch 38 is provided. The pump impeller 30 includes a hydraulic pump 40.
Are connected to generate hydraulic pressure for operating the hydraulic actuator of each section. In the above-described fluid coupling 12, when hydraulic oil is supplied to the release-side oil chamber 46 and hydraulic oil in the engagement-side oil chamber 48 is discharged, the lock-up clutch 38 is released. When hydraulic oil is supplied to the chamber 48 and hydraulic oil in the release-side oil chamber 46 is discharged, the lock-up clutch 3
8, the crankshaft 28 and the output shaft 34 are directly connected.

The forward / reverse switching device 14 is a double pinion type planetary gear device that can be selectively switched to a forward gear or a reverse gear in accordance with an operation position of a shift lever 142 described later. It is arranged on the side opposite to the coupling 12. The output shaft 34 of the fluid coupling 12 penetrates the axis of the input shaft 58 of the CVT 16 and protrudes to the opposite side. A ring gear 52 provided concentrically with 50, a pair of planet gears 54 and 56 meshing with one and the other of the sun gear 50 and the ring gear 52, and rotatably supporting the planet gears 54 and 56. And a carrier 60 which is connected to the input shaft 58 of the CVT 16 so as not to rotate relatively. A multi-plate forward clutch C1 is provided between the sun gear 50 and the carrier 60, and a multi-plate reverse brake B1 is provided between the ring gear 52 and the housing 64. The engagement is controlled by a forward hydraulic actuator 42 and a reverse hydraulic actuator 44, respectively. When the forward clutch C1 is engaged in a state where the reverse brake B1 is released, the output shaft 34 and the carrier 60 are connected to each other so as not to rotate relatively, and the input shaft 58 is rotated integrally with the output shaft 34, and When the clutch C1 is disengaged and the reverse brake B1 is engaged, the rotation of the ring gear 52 is prevented, so that the carrier 6
0 Further, the gear ratio γ _FR (= the rotation speed of the output shaft 34 / the rotation speed of the input shaft 58) = − 1+ (the number of teeth of the ring gear 52) in the direction in which the input shaft 58 is opposite to the output shaft 34, that is, in the direction in which the vehicle moves backward. It is decelerated rotation in Z number of teeth Z _S of _R / sun gear 50).

The CVT 16 has the input shaft 58 and an output shaft 70 parallel to the input shaft 58.
8. A drive-side variable pulley 72 and a driven-side variable pulley 74 are provided on the output shaft 70, and a transmission belt 76 is wound between the variable pulleys 72 and 74. The variable pulleys 72 and 74 are connected to the input shaft 5
8 and a fixed rotating body 7 fixed to the output shaft 70, respectively.
8 and 80, and movable rotating bodies 82 and 84 provided on the input shaft 58 and the output shaft 70 so as to be movable in the axial direction and not to rotate relatively around the axes, respectively. By being moved in the axial direction by the hydraulic actuators 86 and 88 disposed on the back side, the V-groove width, that is, the hanging diameter (effective diameter) of the transmission belt 76 is changed, and the CVT 16 is moved.
Gear ratio gamma _CVT (= rotational speed N _out of the rotational speed N _in / output shaft 70 of the input shaft 58) is adapted to be changed.

The sub-transmission 18 is an automatic transmission composed of a single pinion type planetary gear set, and has an output shaft 7.
A sun gear 90 rotatably arranged around the axis 0
A ring gear 92 connected to the output shaft 70 so as to be relatively non-rotatable; a planet gear 94 meshed with the sun gear 90 and the ring gear 92; a second output shaft 96 that rotatably supports the planet gear 94 And a carrier 98 which is connected to the motor 98 so as not to rotate relatively. A multi-plate high speed clutch C2 is provided between the sun gear 90 and the carrier 98, and a one-way clutch 102 and a multi-plate low speed brake B2 are provided between the sun gear 90 and the housing 64. They are provided in series. High speed clutch C2 and low speed brake B2
Are controlled by a high speed hydraulic actuator 106 and a low speed hydraulic actuator 108, respectively. In the low gear stage established by engaging the low speed brake B2,
The one-way clutch 102 prevents the sun gear 90 from rotating in the opposite direction to the ring gear 92 in the positive torque drive state, but allows the sun gear 90 to rotate in the same direction as the ring gear 92 in the negative torque drive (engine brake) state. Wheel 2
The power transmission path for transmitting the torque of No. 6 to the engine 10 is released. Therefore, the high speed clutch C
2 is released and the low speed gear brake B2 is engaged, the low gear is established. In this state,
When the output shaft 70 of the CVT 16 is rotated in a direction for moving the vehicle forward, the carrier 98 and the second output shaft 96 move in the same direction as the rotation direction of the output shaft 70 to change the gear ratio γ _AT (= the rotation speed N of the output shaft 70). _out / rotation speed N of second output shaft 96
_o ) = 1 + (number of teeth Z _{S of} sun gear 90 / ring gear 92)
With the number of teeth Z _R ). Conversely, the low-speed gear brake B2 is released and the high-speed gear clutch C2
Is engaged, a high-speed gear is established. In this state, the sun gear 90 and the carrier 98 are connected to each other so that they cannot rotate relative to each other, so that the planetary gear device can be rotated integrally, and the second output shaft 96 is connected to the output shaft 70 at the speed ratio γ _AT = 1. Rotated in the same direction. When the vehicle is moving forward, the shift speed can be switched by engaging the high speed clutch C2 while keeping the low speed brake B2 engaged.

A first gear 110 is provided on the second output shaft 96, and a second gear 1 is provided on the intermediate shaft 112.
14 is engaged. The intermediate shaft 112 is rotatably disposed around an axis c parallel to the axis b of the second output shaft 96, and the large-diameter gear 11 of the differential gear device 22.
And a third gear 118 meshed with the second gear 6. Second
The gear 114 has a larger diameter than the first gear 110 and the third gear 1
18 has a smaller diameter than the second gear 114,
The reduction gear device 20 is constituted by the gear 110, the second gear 114, and the third gear 118. The differential gear device 22 is supported rotatably about an axis orthogonal to the drive shaft 24 and rotates with the large-diameter gear 116 integrally with the pair of differential small gears 120. A pair of differential gears 12 connected to a drive shaft 24;
2 is provided. Therefore, the power transmitted from the reduction gear device 20 is equally distributed to the left and right drive shafts 24 in the differential gear device 22, and then the left and right front wheels (drive wheels).
26.

Referring to FIG. 3, the engine 10 includes a well-known ignition device including an ignition coil 27a, a distributor 27b having a shaft (not shown) connected to a crankshaft 28, and a spark plug (not shown). An air flow sensor 129 and a throttle sensor 130 provided in an intake pipe (not shown) of the engine 10 output signals indicating an intake air amount Q of the engine and a throttle valve opening _θth , respectively. Also, the crank angle sensor 1 provided in the distributor 27b
35 outputs a signal representing the rotation angle G of the crankshaft 28. Further, the input shaft rotation sensor 136, the first output shaft rotation sensor 138, and the second output shaft rotation sensor 139 provided in the housing 64 are configured to control the rotation speed N _in of the input shaft 58 of the CVT 16 and the rotation speed N _{out of the} output shaft 70. and it outputs a signal representative of the rotational speed N _o of the second output shaft 96.
A vehicle speed sensor 140 provided on the housing 64 for detecting rotation of the drive shaft 24, that is, the rotation of the front wheel 26.
Outputs a signal corresponding to the vehicle speed SPD. Furthermore, the operation position sensor 144 is operating position of the shift lever 142 P _s
Is output.

Among the above signals, signals indicating the throttle valve opening θ _th , the input shaft rotation speed N _in , and the output shaft rotation speed N _out are supplied to the transmission electronic control unit 132. Further, the transmission electronic control unit 132 as a signal representative of the rotational speed N _e of the signal engine 10 representing the crankshaft rotation angle G from the crank angle sensor 135
Is supplied to Further, the engine electronic control unit 13
3 supplies a signal SF _ETD indicating that the engine torque reduction control by the engine torque reduction control is being executed. The transmission electronic control unit 132 includes a CPU 146, a RAM 148, a ROM 15
0, and a so-called microcomputer including an interface (not shown) and the like.
Uses the temporary storage function of the RAM 148 and
The input signal is processed in accordance with the program stored in M150 to control the gear ratio of the CVT 16, control the engagement of the lock-up clutch 38 of the fluid coupling 12, control the speed change of the auxiliary transmission 18, and control the engagement of the auxiliary transmission 18. The first solenoid valve 152, the second solenoid valve 1
54, the third solenoid valve 156, the fourth solenoid valve 158, the fifth solenoid valve 160, and the sixth solenoid valve 162 are driven.

[0017] The engine electronic control unit 133, the intake air quantity Q of the signal from the sensors, the throttle valve opening theta _th, the crankshaft rotation angle G, and the signal representative of the second output shaft speed N _o Is supplied. The transmission electronic control unit 132 outputs a signal _SFsc indicating that a gear shift in the positive torque drive state of the auxiliary transmission 18, that is, a gear shift command at power-on, has been issued. It is supplied to the device 133. The engine electronic control unit 133 is a microcomputer including a CPU, a ROM, a RAM, and an interface (not shown), similar to the transmission electronic control unit 132, and has a crankshaft rotation angle G and an intake The basic ignition timing of the engine 10 is determined based on the air amount Q, and an ignition signal is supplied to the ignition coil 27a based on the basic ignition timing to control the ignition timing of the engine 10.
On the other hand, during the engine torque reduction control executed at the time of the gear change with the engagement of the high speed gear clutch C2, the torque reduction determined in accordance with the type of gear shift of the vehicle from the relationship stored in the ROM in advance. The basic ignition timing is retarded by a predetermined amount so that the amount is obtained, and the ignition coil 27a is controlled in accordance with the corrected ignition timing to which such correction has been applied.

FIG. 4 shows the relationship between the operating state of the forward clutch C1, the reverse brake B1, the high speed clutch C2 and the low speed brake B2, which is controlled in relation to the operating position of the shift lever 142, and the shift speed. Is shown.

The hydraulic control circuit 170 shown in FIG. 3 is constituted as shown in, for example, FIGS. 5, 6, 7 and 8 separately. 5 to 8, the hydraulic pump 40 constitutes a hydraulic source of a hydraulic control circuit 170.
The hydraulic pump 40 sucks in the hydraulic oil returned to the oil tank (not shown) through the strainer 174, and sucks the hydraulic oil returned through the return oil line 176 and sends it to the primary oil line 178. The operating oil in the primary oil passage 178 is supplied to the return oil passage 176 by the relief type primary pressure regulating valve 180.
And, the primary line oil pressure _Pr1 in the primary oil passage 178 is adjusted by leaking to the clutch pressure oil passage 182. _{Reference numeral} 184 denotes a relief valve for preventing the primary line pressure _Pr1 from being excessively boosted.

The primary pressure regulating valve 180 is a spool valve 1
90, spool valve 1 via spring seat 192
A return spring 194 that applies a biasing force to the valve 90 in the valve closing direction, a first plunger 196 that contacts the spool valve element 190, and a second plunger 198 that contacts the first plunger 196 and has the same diameter as the first plunger 196. The spool valve 190 is connected to the port 200 communicating with the primary oil passage 178.
a and a drain port 200b communicating with the return oil passage 176 and a clutch 200c communicating with the clutch pressure oil passage 182. The spool valve 190 is the first
Land 202 and second land 204 having a larger diameter
And an oil chamber 206 for applying a feedback pressure to an end face of the first land 202. An oil chamber 208 between the first land 202 and the second land 204 is open to the atmosphere. In the oil chamber 206, a primary line oil pressure _Pr1 as a feedback pressure is applied via a throttle 210, and the spool valve element 190 is urged in the valve opening direction. A first plunger 196 and a second plunger 198 provided coaxially with the spool valve 190.
Between the hydraulic pressure P in the primary hydraulic actuator 86
_in the oil chamber 212 is provided for guiding through the branch oil passage 210, and still end face of the second plunger 198 and the oil chamber 214 is provided for guiding the tension control pressure P _belt. The pressure receiving area of the first land 202 of the spool valve 190 is A ₁ , the cross-sectional areas of the first plunger 196 and the second plunger 198 are A ₃ , and the return spring 194
, The spool valve 190 is equilibrated at the position where the expression 1 is established, and the primary line hydraulic pressure P
_r1 is regulated. And this primary line oil pressure _Pr1 is
Throttle valve opening detection valve 2 for adjusting throttle pressure P _th
20, a valve pressure regulating valve 222 for regulating the valve pressure _Pv supplied to each solenoid valve, a tension control pressure regulating valve 224 for regulating the tension control pressure, and a member for engaging and operating the forward clutch C1 and the reverse brake B1. The combined operating oil pressure _Pbc is supplied to the engagement operating oil pressure adjusting valve 226 for adjusting the pressure.

[0021]

[Number 1] P _r1 = _{[(P in or P belt) ·} A 3 + W / A 1 ]

[0022] In the primary regulator valve 180, the hydraulic pressure P _in the tension control pressure P of the primary side hydraulic actuator 86
_{When the belt} pressure is higher than _belt (P _belt = oil pressure P _out in the secondary hydraulic actuator 88 in the steady state), the first plunger 196 and the second plunger 198 are separated from each other and the hydraulic pressure in the primary hydraulic actuator is increased. thrust by P _in acts in the closing direction of the spool 190, but if the primary side hydraulic actuator in the hydraulic P _in is less than the tension control pressure P _belt, the second plunger 1 to the first plunger 196
Since the abutment 98 contacts, the thrust by the tension control pressure P _belt acting on the end face of the second plunger 198 acts in the valve closing direction of the spool valve element 190. As a result, the primary line hydraulic pressure _Pr1 is adjusted to a value proportional to the higher of the primary hydraulic pressure _{in the} hydraulic actuator _Pin and the tension control pressure _Pbelt, and power loss for generating the hydraulic pressure is reduced. It is made as small as possible.

The throttle valve opening detection valve 220 is engaged with a throttle cam 230 which is linked with a throttle valve which is rotated in response to operation of an accelerator pedal (not shown), and is engaged with a cam surface of the throttle cam 230. A plunger 232 whose axial position is changed in relation to the rotation angle of 230, a spool valve 234 for adjusting the throttle pressure P _th , and a spring 236 for urging the spool valve 234 in the valve opening direction. Have.
The spool valve element 234 is configured to close based on the thrust in the valve opening direction applied from the plunger 232 via the spring 238, the thrust in the valve closing direction of the spring 236, and the throttle pressure P _th acting as feedback pressure. by the direction of the thrust is brought into a position so as to balance, and reduces the primary line pressure P _r1, to generate a larger throttle pressure P _th with the throttle valve opening theta _th.

The valve pressure regulating valve 222 is positioned such that the thrust in the valve opening direction applied from the spring 240 and the thrust in the valve closing direction generated based on the valve pressure _Pv acting as feedback pressure are balanced. It includes a spool 242 that is, regardless of the variation of the primary line pressure P _r1 as the original pressure, to generate a constant valve pressure P _v it to reduced pressure. The valve pressure _Pv is supplied to the third solenoid valve 156, the fourth solenoid valve 158, and the fifth solenoid valve 160, respectively. The third solenoid valve 156 and the fourth solenoid valve 158 include an input port to which the valve pressure _Pv is supplied, a drain port,
An output port, wherein the spherical valve element closes the drain port and communicates between the input port and the output port, and the spherical valve element closes the input port and communicates between the drain port and the output port. It is a 3 port 2 position valve that can be switched. The fifth solenoid valve 160
Has a spool valve element 246 biased in the valve closing direction by the spring 244 and the feedback pressure, and a linear solenoid 248 biasing the spool valve element 246 in the valve opening direction by a thrust corresponding to the exciting current. It is configured to generate a signal pressure P _lin that increases according to the exciting current.

[0025] The tension control pressure regulator valve 224, the primary line pressure P _r1 primary oil passage 178 leads to a spool 262 that opens and closes between the tension control pressure oil passage 260 for guiding the tension control pressure P _belt, spring seat A return spring 266 for applying an urging force in the valve opening direction to the spool valve 262 via the H.264, and a plunger 268 for applying an urging force in the valve opening direction in contact with the spool valve 262 are provided. Further, a first land 270 and a second land 272 whose diameters increase in order are formed at the shaft end of the spool valve element 262 in order. Between the first land 270 and the second land 272, there is provided an oil chamber 276 into which a tension control pressure P _belt as a feedback pressure is introduced through a throttle 274. The signal pressure P output from the fifth solenoid valve 160 is provided on the end face side of the first land 270 of the spool valve element 262.
An oil chamber 278 in which _lin acts is provided, and the spool valve element 262 is urged in the valve closing direction based on the speed ratio γ _cvt . The plunger 268 includes a third land 2 having a smaller diameter in order from the spool valve element 262 side.
80 and a fourth land 282 are provided. An oil chamber 284 for applying the throttle pressure P _th is provided on the end face side of the fourth land 282, and the spool valve element 26
2 is urged in the valve opening direction by the throttle pressure _Pth . Further, between the third land 280 and the fourth land 282, an oil chamber 286 to which a signal pressure P _sol4 output from the fourth solenoid valve 158 is applied is provided, and this signal pressure P _sol4 is generated. In this case, the spool valve element 262 is urged in the valve opening direction and the tension control pressure P
_{The belt} is designed to be raised to a predetermined pressure. Therefore, the pressure receiving area of the first land 270 is A ₄ , the cross sectional area of the second land 272 is A ₅ , the cross sectional area of the third land 280 is A ₆ , the pressure receiving area of the fourth land 282 is A ₇ , and the return is performed. Assuming that the urging force of the spring 266 is W, the spool valve element 262 is balanced at a position where the equation 2 is satisfied. The tension control pressure P _belt is basically adjusted based on the throttle valve opening θ _th corresponding to the output torque of the engine 10 and the speed ratio γ _cvt , so that the tension of the transmission belt 76, that is, the clamping pressure is required. In addition, the power transmission belt 76 is controlled to a sufficient value, power loss is reduced, and durability of the transmission belt 76 is enhanced.

[0026]

P _belt = [A ₇ · P _in + (A ₆ -A ₇ ) P _sol4 + W -A ₄ · P _lin ] / (A ₅ -A ₄ )

The engagement hydraulic pressure regulating valve 226 opens and closes a spool valve 292 for opening and closing between a primary oil passage 178 and an engagement hydraulic oil passage 290 and a spool valve 292 via a spring seat 294. Spring 29 biasing in the direction
6 and a plunger 298 that contacts the spool valve 292
And The first land 300 and the second land 302 having the larger diameter in order from the end thereof are provided on the spool valve element 292.
An oil chamber 304 is provided between the first land 300 and the second land 302 so that the engagement operating oil pressure _Pbc acts as a feedback pressure.
The plunger 298 has a spool valve 292
A third land 306 and a fourth land 308 having a smaller diameter are provided sequentially from the side, and a manual valve is provided between the third land 306 and the fourth land 308 when the shift lever 142 is operated to the R range. an oil chamber 312 which R range pressure P _R is supplied to output from 310 is provided. Further, an oil chamber 314 for receiving the throttle pressure P _th acting on the end face of the fourth land 308.
Is provided. Therefore, the spool valve element 292
Is the throttle pressure P _th or the throttle pressures P _th and R
And the valve opening direction of the thrust by the valve opening direction of the thrust and the spring 296 based on the range pressure P _R, the closing direction of the thrust based on the feedback pressure is actuated so as to balance,
Generating an engaging hydraulic pressure P _bc having a magnitude corresponding to the throttle pressure P _th. Also, when the R range pressure P _R is supplied, it increases its engagement hydraulic pressure P _bc only predetermined pressure. Thus, the engagement hydraulic pressure P _bc, as well is increased in accordance with the output torque of the throttle pressure P _th, that is, to the engine 10, the shift lever 142 is increased further by a predetermined pressure therefrom when operated to the R-range, the forward Clutch C1, reverse brake B1, high speed clutch C
2, or the low-speed stage brake B2 is engaged with a necessary and sufficient thrust. The engagement hydraulic pressure P
_bc is also supplied to the first solenoid valve 152, the second solenoid valve 154, and the sixth solenoid valve 162. The sixth solenoid valve 162 is configured similarly to the third solenoid valve 156 and the fourth solenoid valve 158 described above, while the first solenoid valve 152 and the second solenoid valve 154 are throttled when in the off state. 318 and 32
This is a two-port two-position valve that opens the drain downstream from 0 to the drain, but makes the downstream from throttles 318 and 320 _engage hydraulic pressure _Pbc , respectively, when in the ON state.

The first solenoid valve 152 is provided with a shift direction switching valve 330 for switching the direction in which the speed ratio of the CVT 16 changes.
, And the second solenoid valve 154 controls the shift speed control valve 332 for controlling the speed ratio change speed of the CVT 16. The shift direction switching valve 330 includes a first input port 334 that communicates with the primary oil passage 178, a second input port 338 that communicates with the primary oil passage 178 through a medium throttle 336, a relatively small throttle 340, and a relatively large throttle. A first output port 344 communicating with the hydraulic actuator 86 on the primary side via the throttle 342, and an input port 3 of the shift speed control valve 332;
The second output port 348 and the drain port 3 communicating with 46
50, between the first input port 334 and the first output port 344 in the off position and the second input port 338.
And the second output port 348, respectively, but in the ON position, the spool valve 352 communicates between the second output port 348 and the drain port 350.
And a spring 354 for urging the spool valve element 352 toward the off position. Therefore, when the first solenoid valve 152 is turned off, the spool valve 3
Numeral 52 is located at the off position, hydraulic fluid is supplied into the hydraulic actuator 86 on the primary side, and the CVT 16 is changed in the speed increasing direction. Conversely, when the first solenoid valve 152 is turned on, the spool valve element 352 is positioned at the on position, the hydraulic oil in the hydraulic actuator 86 on the primary side is discharged from the drain port 350 and the CVT 1
6 is changed in the deceleration direction.

The shift speed control valve 332 communicates between the input port 346, an output port 356 communicating with the primary hydraulic actuator 86, and between the input port 346 and the output port 356 when in the on position, and in the off position. , A spool valve element 358 that shuts off and a spring 360 that urges the spool valve element 358 toward the off position. Therefore, the second solenoid valve 1
When the valve 54 is turned off, the spool valve 358 shuts off the connection between the input port 346 and the output port 356. Therefore, when the first solenoid valve 152 is in the on state, the mode becomes the slow deceleration mode, and the first solenoid valve 152 Is in the off state, the mode is the slow acceleration mode. When the second solenoid valve 154 is turned on, the spool valve element 358 is connected to the input port 34.
6 and the output port 356, the first
When the solenoid valve 152 is in the off state, the mode is the rapid acceleration mode, and when the first solenoid valve 152 is in the on state, the mode is the rapid deceleration mode. FIG. 9 shows the relationship between the combination of the operating states of the first solenoid valve 152 and the second solenoid valve 154 and the shift mode of the CVT 16 obtained by the combination.

The manual valve 310 has a spool valve element 364 interlocked with the shift lever 142, a first port 366, a second port 368, and a third port 370. , the engaging hydraulic pressure P _bc whose pressure regulated by the engaging hydraulic pressure regulating valve 226 is supplied as a source pressure. From the first port 366,
The shift lever 142 is D range, S range, the forward range pressure P _F is output when it is operated to the forward range such as L range, from the third port 370, the shift lever 14
2 is the reverse range pressure P _R is output when it is operated to the R range.

[0031] The forward range pressure P _F output from the first port 366, through aperture 374, or aperture 37
6 and the shift timing valve 378 are supplied to the forward hydraulic actuator 42. The spool valve element 380 of the shift timing valve 378 moves against the spring 382 in response to an increase in the hydraulic pressure in the hydraulic actuator 42 for forward movement, and suppresses the inflow flow rate. When the shift lever 142 is operated to a range other than the forward range, the hydraulic oil in the forward hydraulic actuator 42 is released from the check valve 38.
4 and drain through the manual valve 310 quickly.

Further, when the shift lever 142 is operated to the R range, the reverse range pressure P _R that is output from the third port 370, the oil chamber 31 of the engagement hydraulic pressure regulating valve 226
2 and the reverse inhibit valve 37
2 and the throttle 386 are supplied to the reverse hydraulic actuator 44. Conversely, when the shift lever 142 is operated to a range other than the R range, the hydraulic oil in the reverse hydraulic actuator 44 is quickly discharged to the drain through the check valve 387, the reverse inhibit valve 372, and the manual valve 310. , reverse range P _R is set to atmospheric pressure.

The reverse inhibit valve 372 has a first land 388, a second land 390 having a diameter larger than the first land 388, and a third land 392 having the same diameter as the first land 388.
A spool valve 394 that opens and closes between the third port 370 and the reverse hydraulic actuator 44 by 0, and a spring 39 that urges the spool valve 394 in the valve opening direction.
6 and a plunger 398 that contacts the spool valve element 394 in order to bias the spool valve element 394 in the valve opening direction. Also,
The plunger 398 has the same cross-sectional area as the cross-sectional area difference between the first land 388 and the second land 390. A signal pressure P _sol3 generated when the third solenoid valve 156 is in the ON state (the engagement state of the lock-up clutch 38) is applied to the end surface of the first land 388, and the first land 388 and the third land 388 are connected to the first land 388. reverse range pressure P _R is allowed to act between the second land 390. Further, the engagement operating oil pressure _Pbc is constantly applied to the end face of the plunger 398, and the spool valve element 394 and the plunger 398
The hydraulic pressure in the reverse hydraulic actuator 44 is made to act between them. Therefore, the spool by thrust and brake engagement pressure P _bc be urged by reverse range pressure P _R to spool 394 in the valve opening direction valve member 394
When the shift lever 142 is operated to the R range (reverse range), the spool valve element 394 is moved to the valve-opening position by the urging force of the spring 396. When it is positioned and the signal pressure P _sol3 is applied, the spool valve element 394 is positioned in the valve closing direction, that is, in the inhibit position.

The forward hydraulic actuator 42 and the reverse hydraulic actuator 44 have a throttle pressure P _th
Accumulator 402 which is acted as back pressure
And 404 are connected respectively, and as the transmission torque increases, the forward hydraulic actuators 42
In addition, the rise of the hydraulic pressure in the reverse hydraulic actuator 44 is moderated, and the engagement of the forward clutch C1 and the reverse brake B1 is smoothened.

The high-speed clutch C2 and the low-speed brake B2 of the auxiliary transmission 18 are switched by a sixth solenoid valve 162 to a C2 control valve 410 and a B2 control valve 412.
It can be switched by. The C2 control valve 410 is connected to an output port 414 communicating with the high-speed stage hydraulic actuator 106 via the engagement pressure oil passage 105 and a port 4 to which the engagement operating oil pressure _Pbc is supplied via a throttle 415.
A spool valve 420 that selectively communicates with a drain port 418 for draining the hydraulic oil 16 through a throttle 417, a spring 422 that biases the spool valve 420 toward the engagement side position, , This spring 4
22 and the signal pressure P _sol6 from the sixth solenoid valve 162
And a throttle 425 for applying hydraulic pressure in the low-speed stage hydraulic actuator 108 to an oil chamber 424 for receiving oil pressure and an end surface of the spool valve 420 opposite to the spring 422 side.
And an oil chamber 426 connected to the low-speed stage hydraulic actuator 108 through the oil chamber 426. Therefore, in the C2 control valve 410, when both the oil chamber 424 and the oil chamber 426 are at the atmospheric pressure, or when the signal pressure P _sol6 and the low-speed stage hydraulic actuator 108 are applied to the oil chamber 424 and the oil chamber 426, respectively.
When the internal hydraulic pressure is supplied, the spool valve 420 is positioned on the engagement side, and the high-speed stage hydraulic actuator 106 engages the high-speed stage clutch C2. However, when the oil pressure in the low-speed stage hydraulic actuator 108 is supplied to the oil chamber 426 while the oil chamber 424 is at atmospheric pressure, the spool valve element 420 moves to the non-engagement side position (the off side position in FIG. 8). After being positioned, the hydraulic oil in the high speed gear hydraulic actuator 106 is drained, and the high speed gear clutch C2 is released.

The B2 control valve 412 operates at the engagement operating pressure P
first port _{430 bC} is supplied, a second port 432 which throttle pressure P _th is supplied, a drain port 434, a third port 436, the forward range pressure P _f is supplied that is connected to the low speed stage hydraulic actuator 108 4th port 4
38, the fifth connected to the back pressure chamber of the accumulator 428
Port 440 and the fifth port 440 to the first port 4
30 or the second port 432, and the spool valve 442 for selectively switching the third port 436 to the drain port 434 or the fourth port 438, and the spool valve 442 to the engagement side position. Spring 444 biasing toward the
An oil chamber 446 that receives the oil pressure in the forward hydraulic actuator 42 and receives the oil pressure to _{apply a} signal pressure P _sol6 to the end face of the spool valve element 442 opposite to the spring 444. It has. Therefore, in the B2 control valve 412, when both the oil chamber 446 and the oil chamber 448 are at the atmospheric pressure,
46 and oil chamber 448 are provided with forward hydraulic actuator 42.
When the internal hydraulic pressure and the signal pressure P _sol6 are supplied, respectively, the spool valve 420 is positioned on the engagement side, and the low-speed gear brake actuator 108 engages the low-speed gear brake B2. However, when the signal pressure P _sol6 is supplied to the oil chamber 448 while the oil chamber 446 is at atmospheric pressure, the spool valve element 442 is positioned at the non-engagement side position, and The hydraulic oil is drained, and the low-speed gear brake B2 is released.

When the shift lever 142 of the vehicle is operated in the R (reverse) range or the N (neutral) range, as shown in FIG. 4, the sixth solenoid valve 162 is held in the ON state and the C2 control valve is turned on. Since the signal pressure P _sol6 in the oil chamber 424 of the 410 is applied, the spool valve 420 of the C2 control valve 410 is located at the engagement position (on position in FIG. 8), and the high-speed _gear clutch C2 is engaged. with provoking, since the forward range pressure P _F is not outputted from the manual valve 310, B2 because the low speed stage brake B2 regardless the operating position of the control valve 412 is in the released has been directly coupled in the auxiliary transmission 18, Power transmission is interrupted only in the forward / reverse switching device 14 in the power transmission path. As a result, even when the shift lever 142 is operated to the N range while the vehicle is traveling, the CVT 16 is rotated, so that the gear ratio control of the CVT 16 can be performed. Further, since the sub-transmission 18 is in the direct connection state in the N range and the R range as described above, when the shift lever 142 is operated from the N range to the R range or the D range,
Since only one friction engagement device, which is the forward clutch C1 or the reverse brake B1, needs to be engaged, the forward gear or the reverse gear can be established without requiring complicated timing control. FIG. 10 shows the high-speed gear hydraulic actuator 10 when the shift lever 142 is operated to the R range or the N range as described above.
6 shows the hydraulic pressure (C2 pressure) and the hydraulic pressure (B2 pressure) of the low-speed stage hydraulic actuator 108. Incidentally, when the shift lever 142 is operated to the R range, the R range pressure P _R as engaging force corresponding to the transmission torque of the reverse gear is obtained an oil chamber of the engagement hydraulic pressure regulating valve 226 31
2, the engagement operating pressure _Pbc is increased by a predetermined pressure, so that the C2 pressure is higher than in the N range.

However, immediately after the shift lever 142 is operated from the N range to the D (forward) range, the shift lever 142 is driven through a high gear stage with a small gear ratio and then a low gear stage is established to gradually change the drive torque. The control is started, and during the squat control period, as shown in FIG. 4, the high-speed gear crack C2 and the low-speed gear brake B2
Are maintained in the same state as the previous N range. That is, the ON state of the sixth solenoid valve 162 is maintained by the electronic control unit 132 and the spool valve 42 of the C2 control valve 410 is turned on.
0 is located at the engagement position (on position in FIG. 8), and the high speed clutch C2 is continuously engaged. In this state, the forward range pressure P
_F is output to supply hydraulic oil into the forward hydraulic actuator 42 for engaging the forward clutch C1 and into the oil chamber 446 of the B2 control valve 412, and the hydraulic pressure (C1 pressure) in the forward hydraulic actuator 42 is reduced. Although it rises slowly due to the action of 374 and 376 and the accumulator 402, until it rises, the signal pressure P _sol6 is _applied in the oil chamber 448 of the B2 control valve 412, and the spool valve element 442 is disengaged ( 8, the state in which the low-speed gear hydraulic actuator 108 for operating the low-speed gear brake B2 is drained is maintained. Therefore, in the auxiliary transmission 18, only the high-speed gear clutch C2 is provided. It will be in the engaged state.

When the shift lever 142 is operated from the N range to the D range, the forward clutch C
1 is completed, for example, when the rotation of the output shaft 34, which has been rotating until then, is stopped, or when, for example, about 0.7 seconds have elapsed since the start of the squat control, the squat control is ended. In order to perform this, the sixth solenoid valve 162 is switched to the off state by the electronic control unit 132. As a result, the oil chamber 448 of the B2 control valve 412 to which the signal pressure P _sol6 has been _applied is brought to the atmospheric pressure, and the spool valve element 442 is engaged (the off position in FIG. 8).
Since being is located, the forward range pressure P _F to the low speed stage hydraulic actuator 108 for actuating the low speed stage brake B2 is supplied, the low-speed brake B2 is engaged. At the same time, the oil chamber 424 of the C2 control valve 410 to which the signal pressure P _sol6 has been applied is set to the atmospheric pressure, while the pressure of the hydraulic oil in the low-speed stage hydraulic actuator 108 is controlled through the throttle 425 to the C2 control. Since the spool valve element 420 is also acted on by the oil chamber 426 of the valve 410, the spool valve element 420 is positioned at its disengaged position (off position in FIG. 8) against the urging force of the spring 422, and the high-speed clutch C2
Is released. That is, when the squat control temporarily executed in association with the operation of the shift lever 142 from the N range to the D range ends, the low speed gear is established in the auxiliary transmission 18 as described above. Thus, the driving force at the time of starting the vehicle can be obtained. FIG.
Indicates the C1, C2, and B2 pressures that change in response to the shift lever 142 being operated from the N range to the D range. As described above, the sixth solenoid valve 1
Even if 62 is not switched from the on state to the off state,
When the C1 pressure for engaging the forward clutch C1 increases in association with the operation of the shift lever 142 from the N range to the D range, the spool valve element 420 of the B2 control valve 412 moves to the off position in FIG. As a result, the low-speed gear brake B2 is engaged.

Here, when the shift lever 142 is operated from the N range to the D range and the low gear is established after the squat control period, the pressure of the hydraulic oil in the low gear hydraulic actuator 108 increases. As shown in FIG. 11, the spool valve element 420 of the C2 control valve 410 is moved to the non-engaged position based on the engagement pressure of the low-speed gear brake B2 (２B2 pressure). ) Reaches at least a value α that generates a thrust corresponding to the urging force of the spring 422 (for example, 2.5 to 3 kg), and then the release of the high speed gear clutch C2 is started. Accordingly, the period in which both the low-speed gear brake B2 and the high-speed gear clutch C2 are temporarily overlap-engaged is minimized and surely provided. The shift shock generated in connection with the shift to the gear is properly prevented, and the engine 10 is reliably prevented from blowing up.

In the running state in which the low gear is established as described above, for example, Japanese Patent Application Laid-Open No. 61-241561
As described in Japanese Patent Application Laid-Open Publication No. H11-260, when the traveling state of the vehicle falls within a switching permission region from a low gear position to a high gear position in the shift diagram shown in FIG.
2, the sixth solenoid valve 162 is switched from the off state to the on state. As a result, in the C2 control valve 410, when the signal pressure P _sol6 is _applied to the oil chamber 424 that has been at the atmospheric pressure, the spool valve element 420 is positioned at the engagement position (the ON position in FIG. 8). Therefore, the engagement operating oil pressure _Pbc is passed through the throttle 415 so that the high-speed stage hydraulic actuator 1
06 to engage the high speed clutch C2. At the same time, the B2 control valve 412, which was previously at atmospheric pressure,
The signal pressure P _sol6 is also applied to the oil chamber 448 of the B2 control valve 412.
Since it has reached _{F, the} spool valve element 442 is held at the engagement position by the urging force of the spring 444. That is, in the auxiliary transmission 18, the low-speed gear brake B2
In this state, the high-speed gear clutch C2 is engaged, and the high-speed gear is established. In this embodiment, when the running state of the vehicle at the time of the upshift is a positive torque driving state accompanied by engine torque reduction control, as described later, the engagement characteristic due to the back pressure of the accumulator 520 not being applied. Change control is performed.

When a kick-down operation of depressing an accelerator pedal (not shown) is performed while the vehicle is running at a high gear, the electronic control unit 132 switches the sixth solenoid valve 162 from the on state to the off state. . As a result, the signal pressure P _sol6 acting on the oil chamber 448 of the B2 control valve 412 is set to the atmospheric pressure, but the operating valve in the forward hydraulic actuator 42 causes the spool valve element 442 to be already in the engaged position. The engagement position is maintained as it is. At the same time, since the oil chamber 424 of the C2 control valve 410 to which the signal pressure P _sol6 has been applied is set to the atmospheric pressure, the spool valve element 420
Is disengaged position (off position in FIG. 8) against the urging force of the spring 422 by a thrust based on the pressure in the low speed gear hydraulic actuator 108 (forward range pressure P _F ) acting on the oil chamber 426. ), The hydraulic oil in the high-speed gear hydraulic actuator 106 is discharged through the throttle 417 and the drain port 418, thereby releasing the high-speed gear clutch C2 and establishing a low gear. Further, as described in Japanese Patent Application Laid-Open No. 61-241561, the same applies to the case where the actual gear ratio γ _CVT falls below a predetermined value γ _o while the vehicle is running. Then, the downshift of the subtransmission 18 is performed. In this embodiment, at the time of the downshift, a release characteristic change control of the high-speed gear clutch C2, which will be described later, is executed. FIG. 14 shows changes in the C1, C2, and B2 pressures when the sixth solenoid valve 162 is switched from the on state to the off state in order to obtain a low gear in the traveling state in the high gear. I have.

Even during the squat control period immediately after the operation from the N range to the D range, if the accelerator pedal is depressed, the electronic control unit 132 immediately turns the sixth solenoid valve 162 off from the on state. State. As a result, the signal pressure P _sol6 acting on the oil chamber 448 of the B2 control valve 412 is set to the atmospheric pressure, and the spool valve 4
Since the reference numeral 42 is moved from the previously disengaged position to the engaged position, hydraulic oil is supplied into the low-speed gear hydraulic actuator 108. At the same time, since the oil chamber 424 of the C2 control valve 410 to which the signal pressure P _sol6 has been applied is set to the atmospheric pressure, the spool valve 420 is moved to the low speed stage hydraulic actuator 108 which is operated in the oil chamber 426. The disengagement position (off position in FIG. 8) of the spring 422 against the urging force of the spring 422 by the thrust based on the internal pressure (forward range pressure P _F ), the operation within the high-speed gear hydraulic actuator 106 Oil is throttled 417 and drain port 4
The clutch C for the high-speed gear
2 is released, and the low gear stage is established. FIG.
Represents changes in the C1, C2, and B2 pressures when the accelerator pedal is depressed during the squat control period.

Further, when the shift lever 142 is operated from the D range to the R range in the low gear stage of the auxiliary transmission 18, such as during low-speed forward running immediately before stopping, the manual valve 310 engages the shift lever 142 until then. At the same time as the forward clutch C1 being disengaged is released and the reverse brake B1 is started to be engaged by the reverse hydraulic actuator 42, the sixth solenoid valve 162 is switched from the off state to the on state as shown in FIG. As a result, the engagement of the high speed clutch C2 which has been in the non-engaged state is started. At this time, the forward hydraulic actuator 42
The internal pressure and the oil chamber 446 of the B2 control valve 412 are exhausted by the manual valve 310, and the signal pressure P _sol6 from the sixth solenoid valve 162 is applied to the oil chamber 448 of the B2 control valve 412. spool 442 of 412 so is moved to the on position, the back pressure of the accumulator 520 is switched to the high engagement hydraulic pressure P _bc than P _th. As a result, as shown by a broken line in FIG. 15, the transient hydraulic pressure in the high-speed gear hydraulic actuator 106 when the shift lever 142 is operated from the D range to the R range is shown by a solid line due to the action of the accumulator 520. The speed is increased as compared with the case of normal gear change, and the engagement capacity of the high speed clutch C2 is higher than that of the reverse brake B1.

The accumulator 520 and the accumulator back pressure switching valve 521 control the time required for engagement so that the engagement and disengagement of the high speed clutch C2 becomes smooth. The accumulator 520 is connected to the first port 414 of the C2 control valve 410 to supply the engagement operating oil pressure _Pbc in the engagement pressure oil passage 105 to the accumulator port 5.
22 and the back-pressure oil passage 524 in accordance with the operating position of the B2 control valve 412 according to the engagement operating oil pressure _Pbc and the throttle pressure _Pth.
And a back pressure port 526 for guiding atmospheric pressure to the back pressure chamber 525;
A piston 532 that changes the volume in the pressure accumulator 530 by being retracted to the end face side opposite to the pressure accumulator port 522 in response to the urging force of No. 8; The rising characteristic of the engagement hydraulic pressure in the engagement pressure oil passage 105 can be changed according to the magnitude of the back pressure.

The accumulator back pressure switching valve 521 has a first position for supplying back pressure to the back pressure chamber 525 of the accumulator 520 and a clutch for the high speed gear from the first position by opening the back pressure chamber 525 to the atmosphere. By selectively placing the clutch C2 in the second position in which the engagement characteristic of the clutch C2 is moderate, the engagement of the high-speed clutch C2 can be performed during a normal shift and a shift accompanied by engine torque reduction control. It changes the characteristics and corresponds to the engagement characteristic switching valve of the present invention. This accumulator back pressure switching valve 521 is
A drain port 534; an input port 536 connected to the fifth port 440 of the B2 control valve 412 via the first back pressure oil passage 524a; an output port 538 connected to the second back pressure oil passage 524b; It has a spool valve element 540 for selectively connecting the output port 538 to the input port 536 and the drain port 534, and a return spring 542 for urging the spool valve element 540 toward the first position. The spool valve element 540 is provided with a first land 544, a second land 546 having a diameter larger than the first land 544, and a third land 548 having the same diameter as the second land 546 from the lower end surface side. . The signal pressure P of the fourth solenoid valve 158 is applied to the end surface of the first land 544.
_An oil chamber 550 for receiving _sol4 is provided, and an oil chamber 552 for receiving the signal pressure P _sol3 of the third solenoid valve 156 is provided between the first land 544 and the second land 546, respectively. _Only when the signal pressures P _sol4 and P _sol3 are _{applied to} both the oil chambers 550 and 552, the spool valve 540 is positioned at the second position shown on the right side of FIG. 8 against the return spring 542. As a result, the drain port 534 and the output port 538 communicate with each other and the back pressure oil passage 524 is shut off.
_When at least one of _sol4 and P _sol3 is not actuated, the spool valve element 540 is located at the first position shown on the left side of FIG.
And the output port 538 communicate with one back pressure oil passage 524.
a and 524b are formed.

Therefore, when at least one of the fourth solenoid valve 158 and the third solenoid valve 156 is turned off, the fifth port 440 of the B2 control valve 412 and the accumulator back pressure switching are performed by the back pressure oil passage 524. The back pressure P is sequentially passed through the input port 536 and the output port 538 of the valve 521.
_bc or P _th is the back pressure port 52 of the accumulator 520
6, the engagement operating oil pressure P _bc ′ in the engagement pressure oil passage 105 is adjusted to a value corresponding to the back pressure applied to the piston 532. However, when both the fourth solenoid valve 158 and the third solenoid valve 156 are turned on, the hydraulic oil in the back pressure chamber 525 drains through the back pressure port 526, the output port 538, and the drain port 534 sequentially. Therefore, the piston 532 is quickly retracted to the position where the volume of the pressure accumulating chamber 530 is maximized against the spring 528, that is, the position shown on the left side of FIG. 8, and the engagement operating oil pressure P _bc ′ rises. Is alleviated. When the piston 532 is set to the left position in FIG. 8, the engagement hydraulic pressure P _bc ′ in the engagement pressure oil passage 105 thereafter rises to the engagement hydraulic pressure P _bc upstream of the throttle 415. In the embodiment, the _urging force of the spring 528 is set so that the above-described increase in the engagement operating oil pressure P _bc ′ in the engagement pressure oil passage 105 does not occur until the engagement of the high-speed clutch C2 is completed. ing.

Next, the clutch oil pressure P _cl in the clutch pressure oil passage 182 is adapted to be pressure regulated according to the throttle pressure P _th by a clutch pressure regulating valve 450. The clutch oil pressure P _cl is related to the engagement pressure of the lock-up clutch 38, but the hydraulic oil in the clutch pressure oil passage 182 passes through the throttle 453 to the sliding portion of the transmission belt 76, the bearing portion, and the planetary gear. Meshing part, differential gear device 2
2 and the like as lubricating oil. In addition, the hydraulic oil discharged from the relief port 455 of the clutch pressure regulating valve 450 is also pressure-fed as lubricating oil. The clutch pressure regulating valve 450 includes a spool valve element 452 for allowing the hydraulic oil in the clutch pressure oil path 182 to escape to the return oil path 176, a spring 454 for urging the spool valve element 452 in the valve closing direction, A plunger 456 that receives the throttle pressure P _th and transmits a thrust in the valve closing direction based on the throttle pressure P _th to the spool valve element 452, and a clutch oil pressure P as a feedback pressure to apply a thrust in the valve opening direction to the spool valve element 452.
_An oil chamber 458 for receiving _cl is provided.

In order to control engagement and disengagement of the lock-up clutch 38, a third solenoid valve 156 and a fourth solenoid valve 15 corresponding to the switching solenoid solenoid valve device of the present invention.
8, a lock-up relay valve 460 switched by the third solenoid valve 156 and a lock-up control valve 462 switched by the fourth solenoid valve 158 are provided. The lock-up relay valve 460 is connected to the drain port 46.
4. First port 468, second port 470, third port 4 to which clutch oil pressure _Pcl is supplied via check valve 466
72, a fourth port 474, a fifth port 476, a drain port 478, a spool valve element 480 for switching between these ports, and the spool valve element 480 connected to the oil chamber 4
And a spring 484 biasing toward the 82 side. For this reason, the signal pressure P _sol3 from the third solenoid valve 156 is
82, the spool valve element 480 is positioned toward the oil chamber 482, so that the first port 468 and the second port 470, and the third port 472 and the fourth port 47
The fourth and fifth ports 476 and the drain port 478 communicate with each other. Conversely, when the signal pressure P _sol3 from the third solenoid valve 156 is supplied to the oil chamber 482, the spool valve element 480 is positioned toward the oil chamber 484, so that the drain port 464 and the second port 470, 1 port 468
And the third port 472, the fourth port 474 and the fifth port 4
76 are communicated with each other.

The lock-up control valve 462 includes a check valve 46
First port 4 to which clutch hydraulic pressure _Pcl is supplied via port 6
90, second port 470 of lock-up relay valve 460
Port 492 connected to the third port 49 connected to the fifth port 476 of the lock-up relay valve 460.
4. Fourth port 496 connected to third port 472 of lock-up relay valve 460, fluid coupling 1
A fifth port 498 connected to the second release-side oil chamber 46, a sixth port 500 connected to the engagement-side oil chamber 48 of the fluid coupling 12, and a spool valve 502 for switching between these ports. And a spring 506 for urging the spool valve 502 toward the oil chamber 504. Therefore, when the signal pressure P _sol4 from the fourth solenoid valve 158 is not supplied to the oil chamber 504, the spool valve element 50
2 is located on the oil chamber 504 side, so the second port 4
92 and the fifth port 498, and the fourth port 496 and the sixth port 500, respectively. Conversely, when the signal pressure P _sol4 from the fourth solenoid valve 158 is supplied to the oil chamber 504, the spool valve element 502 is positioned toward the spring 506, so that the first port 490 and the fifth port 4
98, the third port 494 and the sixth port 500 are communicated with each other.

Therefore, the fourth solenoid valve 158 and the third solenoid valve 158
When both of the solenoid valves 156 are turned off, the clutch oil pressure P _cl is applied to the release side oil chamber 4 via the first port 468, the second port 470, the second port 492, and the fifth port 498 sequentially.
6 and at the same time, the hydraulic oil in the engagement-side oil chamber 48 is drained through the sixth port 500, the fourth port 496, the third port 472, the fourth port 474, and the oil cooler 510 in that order, and locked. The up clutch 38 is released. That is, the first release mode is set. At this time, part of the hydraulic oil discharged from the engagement side oil chamber 48 is also drained through the oil cooler 510. This oil cooler 51
On the upstream side of 0, a cooler bypass valve 512 for draining without passing through the oil cooler 510 when the pressure of the hydraulic oil discharged from the engagement side oil chamber 48 exceeds a predetermined value is provided. When both the fourth solenoid valve 158 and the third solenoid valve 156 are turned on, the clutch oil pressure P _cl is applied to the release-side oil chamber 46 via the first port 490 and the fifth port 498, and at the same time. The hydraulic oil in the engagement side oil chamber 48 is supplied to the sixth port 500, the third port 494, the fifth port 476, the fourth port 474, and the oil cooler 5.
After being drained through 10, the lock-up clutch 38 is released. That is, the second release mode is set. Thus, in this embodiment, there are two modes for releasing the lock-up clutch 38. Of these, usually the third
The former first release mode in which both the solenoid valve 156 and the fourth solenoid valve 158 are turned off is used, and the third solenoid valve 15
The second release mode, in which the sixth and fourth solenoid valves 158 are both turned on, is selected when, for example, an engagement characteristic change control is performed to make the engagement characteristic of the high speed clutch C2 a gradual rise. Is done. The control for changing the engagement characteristics will be described later in detail.

When the fourth solenoid valve 158 is turned on and the third solenoid valve 156 is turned off, the clutch oil pressure P
_cl is applied to the release-side oil chamber 46 via the first port 490 and the fifth port 498, and at the same time, the engagement-side oil chamber 48
The hydraulic oil inside is drained through the sixth port 500, the third port 494, the fifth port 476, and the drain port 478, and the lock-up clutch 38 is promptly released.
In this case, in addition to the hydraulic oil in the engagement side oil chamber 48 flowing out to the drain without passing through the oil cooler 510, the signal pressure P _sol4 from the fourth solenoid valve 158 is used as the tension control pressure regulating valve. 224 is applied to the oil chamber 286 to control the tension P
with _belt is increased, the throttle valve opening degree that the tension control pressure P _belt is enhanced even primary line pressure P _r1 is applied to the oil chamber 214 of the primary regulator valve 180, to source pressure the primary line pressure P _r1 The throttle pressure P _th output from the detection valve 220 is also increased, and the clutch pressure regulating valve 450
Since the clutch pressure P _cl is increased to be pressure regulated in is the lock-up clutch 38 is rapidly released. Such a rapid release mode of the lock-up clutch 38 is selected when the CVT 16 executes the rapid deceleration shift in connection with the sudden stop of the vehicle.

When the fourth solenoid valve 158 is turned off and the third solenoid valve 156 is turned on, the clutch oil pressure P
_cl is applied to the engagement-side oil chamber 48 via the first port 468, the third port 472, the fourth port 496, and the sixth port 500, and the hydraulic oil in the release-side oil chamber 46 is _supplied to the fifth port 468. 498, second port 492, second port 47
0, the oil is drained through the drain port 464, and the lock-up clutch 38 is engaged. That is, the engagement mode is set. FIG. 16 shows a relationship between a combination of operating states of the third solenoid valve 156 and the fourth solenoid valve 158 and an operating mode of the lock-up clutch 38.

The electronic control unit for transmission 1
32, the engagement control of the lock-up clutch 38, the CV
Speed ratio control of T16, rapid deceleration control of CVT16, tension control pressure control, speed change control of subtransmission 18, subtransmission 1
8 is performed.

In the engagement control of the lock-up clutch 38, for example, it is determined from the relationship previously stored based on the vehicle speed SPD and the throttle valve opening _θth whether the region is the engagement region or the release region. If determined, the
The third solenoid valve 156 is turned on and the fourth solenoid valve 158 is turned on.
Is turned off and the engagement mode is set, and when it is determined that the area is the release area, both the third solenoid valve 156 and the fourth solenoid valve 158 are turned off and the first release mode is set. In the rapid deceleration control of the CVT 16 prior to the sudden stop of the vehicle, the first solenoid valve 152 and the second solenoid valve 154 are both turned on to enter the rapid deceleration shift mode, and at the same time, the third solenoid valve 156 is activated. The lock-up clutch 38 is quickly released by being turned off and the fourth solenoid valve 158 being turned on to be in the rapid release mode.

In the gear ratio control of the CVT 16, for example, the first solenoid valve 152 and the second solenoid valve 154 are driven so that the engine 10 operates along an optimum curve for obtaining fuel efficiency and drivability. γ _CVT is regulated. In the tension control pressure control, for example, the tension control pressure P _{belt applied} to the secondary-side hydraulic actuator 88 is stored in advance so as to obtain a target pressure at which the target pressure becomes small within a range in which the transmission belt 76 does not slip. Fifth from the relationship
The solenoid valve 160 is controlled. In the speed change control of the sub-transmission 18, for example, as described in Japanese Patent Application Laid-Open No. 61-241561 or Japanese Patent Application Laid-Open No. 62-137239, the actual throttle valve opening is determined based on a relationship stored in advance. The gear position to be switched is determined based on _θth and the gear ratio γ _CVT or the vehicle speed SPD. If the determination result is a high gear position, the sixth electromagnetic valve 162 is turned on. 162 is turned off.

Next, the control for changing the engagement characteristics will be described with reference to FIG.
This will be described in detail according to the flowchart of FIG. This routine is repeatedly executed at predetermined time intervals in parallel with other control routines or by interruption or the like. First, in steps not shown, a throttle valve opening theta _th is the state amount of a traveling vehicle, the input shaft rotation speed N _in, the output shaft rotation speed N _out, the signal indicative of the engine rotational speed N _e, and the electronic engine The signal SF _{ETD and the} like from the control device 133 are read.

In step S1, the content of the switching command flag _Fc2 indicating the command to switch from the low gear to the high gear based on the shift diagram of FIG. 12 is "1" by the shift switching control of the subtransmission 18. Is determined. If this determination is denied, this routine is terminated, and the accumulator back pressure switching valve 521 is operated by the third solenoid valve 156 and the fourth solenoid valve 158 controlled according to the engagement control described above, so that the normal first Maintained in position. However, if the determination in step S1 is affirmative, the process proceeds to step S1.
In 2, it is determined whether or not the engine torque reduction control is being executed by the engine electronic control device 133 based on whether or not the content of the engine torque reduction flag F _ETD is “1”.

FIG. 18 shows that the output shaft 34, that is, the rotation speed of the turbine wheel 32 and the operation mode of the lock-up clutch 38 during the shift-up switching from the low gear to the high gear at power-on are the same time axis. Shown above. Figure t ₁ of 18 is a time when the contents of the switching instruction flag F _c2 to a high-speed gear in step S1 is determined to be "1", t ₂ is actually the sixth solenoid valve 1
This is the point in time when the switching signal to the high gear is output to 62.

Here, in the engine electronic control device 133 corresponding to the engine torque reduction control means, the engagement time is shortened and the shift shock is alleviated at the time of gear change with engagement of the hydraulic friction engagement device. The engine torque reduction control is executed as shown in FIG. 4 of Japanese Patent Application Laid-Open No. 3-84250, and a signal SF indicating an engine torque reduction flag F _ETD is provided while the engine torque reduction control is being performed. _{The ETD} is transmitted to the electronic control unit 132 for transmission. The execution period of the engine torque reduction control in this embodiment is t _{3 in} FIG.
From is a period until t _5, substantial engine torque reduction operation period ignition coil 27b is controlled according to the correction ignition timing to reduce the engine torque is from t ₃ in FIG. 19 to t _4, t by the engine torque return command is issued at ₄ time is reduced duration of the engine torque is terminated at t ₅ when the delayed therefrom a predetermined time. The time t ₃ is the time when it is determined that the inertia phase in which the rotation speed of the rotating member changes due to the flow of the hydraulic oil into the high speed gear hydraulic actuator 106 and the generation of frictional force in the high speed gear clutch C2, t
₄ the time and the recovery completion of the termination and the engine torque of the inertia phase is detected based on beta _1, which is set in advance so as to substantially coincide, also t ₅ corresponds respectively to the end of the inertia phase. Further, α ₁ in FIG. 19 is a constant determined in advance in consideration of a detection error or the like in order to detect the start of the inertia phase.

After the command to switch from the low gear to the high gear is issued, the content of the engine torque reduction flag _FETD is switched to "1" indicating execution of the engine torque reduction control. Be executed. In step S3, the second release mode of the lock-up clutch 38 is selected, so that the third solenoid valve 156 and the fourth
The solenoid valves 158 are both switched on. As a result, the accumulator back pressure switching valve 521 sets the normal first mode corresponding to the previous engagement mode or the first release mode.
Since the position is switched from the position to the second position in response to the second release mode, the back pressure chamber 525 of the accumulator 520 is opened to the atmosphere, and the engagement characteristic of the high speed gear clutch C2 is changed as shown in FIG.
As shown by the dashed line 9, the gradual rise is switched.

While the above steps are repeatedly executed, when the engine torque lowering control reaches t ₅ , the engine torque lowering flag F is set at step S2.
_Since the content of the _ETD is determined to be "0", the normal first release mode of the lockup clutch 38 is selected in step S4. As a result, since the third solenoid valve 156 and the fourth solenoid valve 158 are both turned off and the accumulator back pressure switching valve 521 is set to the first position, back pressure is applied to the accumulator 520. The engagement operating oil pressure P _bc ′ in the oil passage 105 is controlled according to the back pressure.

FIG. 20 shows the operation of the turbine speed and the lock-up clutch 38 in the release characteristic change control of the high-speed clutch C2 executed during the downshift from the high gear to the low gear during power-on. The modes are shown on the same time axis. This release characteristic change control is for reducing the shift shock by slowing the release speed of the high speed clutch C2.
This is executed independently of the engine torque reduction control. FIG.
At 0, the switching command to the low gear at the time point t ₁ is issued, when subsequently switching signal to the low gear to the sixth solenoid valve 162 at t ₂ time is outputted, it is at the start of the inertia phase is a second release mode of the lockup clutch 38 at t ₃ accumulator back pressure switching valve 521 is a second position. For this reason, the release characteristic of the high speed gear clutch C2 is switched so as to gradually fall as compared to when the accumulator back pressure control valve 521 is in the first position. At time t ₄ when the inertia phase ends, the lock-up clutch 38 is set to the normal first release mode, and the accumulator back pressure switching valve 521 is switched to the normal first position. Note that α ₂ in FIG. 20 is a constant determined in advance in consideration of a detection error or the like for detecting the start of the inertia phase. The end of the inertia phase at t ₄ is detected based on, for example, satisfaction of Equation 3 continuously for a predetermined number of times in order to increase reliability.

[0064]

N _out <N _o × γ _AT1 −β ₂ where γ _ATl is a speed ratio of a low gear, and β ₂ is a constant determined in consideration of a detection error and the like.

As described above, according to the present embodiment, during the period in which it is determined that the engine torque reduction control is being executed in step S2 corresponding to the engine torque reduction determination means, the second release mode switching instruction means is provided. In step S3 corresponding to the third solenoid valve 156 and the fourth solenoid valve 1
When both 58 are turned on, signal pressures P _sol3 and P _sol4 are output. Then, in response to the signal pressures P _sol3 and P _sol4 , the valve element 540 of the accumulator back pressure switching valve 521 moves the engagement pressure oil passage 105 from the first position.
_Is switched to the second position for lowering the engagement operating oil pressure P _bc ′.

In the present embodiment, the third solenoid valve 156 and the fourth solenoid valve 15 correspond to the switching solenoid solenoid valve.
The signal pressure P _SOL3 for switching from 8 to a second release mode
_And an accumulator back pressure switching valve 521 for changing the engagement characteristic of the high speed clutch C2 in response to P _sol4 and P _sol4. The lock-up clutch 38 is released in synchronization with the start of the control, and the engagement characteristic of the high-speed clutch C2 is changed to a gentle rise. As described above, since the third solenoid valve 156 and the fourth solenoid valve 158 for controlling the lock-up clutch 38 can also be used for changing the engagement characteristics, a new electromagnetic switching valve for changing the engagement characteristics is independently provided. It is not necessary to provide, and the shift shock can be reduced favorably by the inexpensive hydraulic control device having a simple hydraulic circuit.

FIG. 19 shows the engagement pressure oil passage 105 at the time of a shift-up shift accompanied by the engagement operation of the high speed clutch C2.
FIG. 6 shows changes over time in the engagement operating oil pressure P _bc ′ and the output torque of the auxiliary transmission 18. In this embodiment, the engine torque reduction control is not performed, and the accumulator 520 is used.
The three cases where the back pressure is applied to the accumulator 520 and the case where the back pressure is applied to the accumulator 520 are displayed in comparison. I of FIG. 19, II, III is a point corresponding to t ₅ of the respective engagement point, i.e. 18, also I, II, the output torque I of III point ', II', when III 'each Of output peak torque. As is apparent from this figure, according to the present embodiment, since the engagement characteristic change control is performed in synchronization with the engine torque reduction control, the output torque of the sub-transmission 18 is reduced, and the engagement characteristic is reduced. Since the actual shifting time required from the start to the completion of the engagement operation is extended slowly, the output peak torque at the time of completion of the engagement is reduced as much as possible. It should be noted that the area indicated by oblique lines in FIG. 19 substantially corresponds to the amount of inertia torque released from the engine, and the areas including I ′ and II ′ are equal.

Although the embodiment of the present invention has been described in detail with reference to the drawings, the present invention can be embodied in other forms.

For example, in the above-described embodiment, the case where the present invention is applied to an automatic transmission having a CVT (belt-type continuously variable transmission) 16 has been described. The transmission may be a so-called traction-type continuously variable transmission in which power is transmitted via rollers. Instead of these, the present invention can be applied to a stepped automatic transmission including a planetary gear transmission mechanism. In short, any automatic transmission for a vehicle having a direct-coupled clutch and capable of switching the gear stage by engaging and disengaging a hydraulic friction engagement device may be used. In the latter case,
Instead of or in addition to the auxiliary transmission 18, the shift shock related to the friction engagement device may be suppressed by engine torque reduction control and engagement characteristic change control.

In the above-described embodiment, the auxiliary transmission 18 is provided downstream of the CVT 16, but the present invention is also applicable to a power transmission device of a type provided upstream of the CVT 16. Further, the auxiliary transmission 18 has two forward gears, but may have three or more gears.

In the above-described embodiment, the fluid coupling 12 with the lock-up clutch is provided.
Instead, a torque converter with a lock-up clutch may be provided.

Although not specifically exemplified, the present invention can be implemented in various modified and improved modes based on the knowledge of those skilled in the art.

[Brief description of the drawings]

FIG. 1 is a diagram illustrating the gist of the present invention.

FIG. 2 is a skeleton diagram of a vehicle power transmission device to which the embodiment of FIG. 3 is applied;

FIG. 3 is a block diagram illustrating a configuration of a control device according to an embodiment of the present invention.

FIG. 4 is a table illustrating a relationship between an operation position of a shift lever and a gear in the power transmission device of FIG. 2;

FIG. 5 is a part of a hydraulic circuit diagram showing a configuration of the hydraulic control device in FIG. 3;

FIG. 6 is a part of a hydraulic circuit diagram showing a configuration of the hydraulic control device in FIG. 3;

FIG. 7 is a part of a hydraulic circuit diagram showing a configuration of the hydraulic control device of FIG. 3;

FIG. 8 is a part of a hydraulic circuit diagram showing a configuration of the hydraulic control device in FIG. 3;

FIG. 9 is a diagram showing a relationship between a control deviation ΔN _in and a shift mode _in feedback control of the continuously variable transmission shown in FIG. 2;

FIG. 10 is a time chart for explaining states of B2 pressure and C2 pressure when operating the R range and the N range in the hydraulic circuit of FIG. 9;

11 is a time chart for explaining changes in C1, B2, and C2 pressures immediately after the operation from the N range to the D range in the hydraulic circuit diagram of FIG. 8;

FIG. 12 is a shift diagram for switching gears of the subtransmission shown in FIG. 2;

FIG. 13 is a time chart illustrating changes in the B2 pressure and the C2 pressure immediately after the sixth solenoid valve is switched from the ON state to the OFF state in the driving state in the D range in the hydraulic circuit diagram of FIG. 8; .

FIG. 14 is a diagram showing the hydraulic circuit diagram of FIG.
6 is a time chart for explaining changes in C1, B2, and C2 pressures when an accelerator pedal is depressed during a squat control period immediately after a range operation.

FIG. 15 is a diagram showing a hydraulic circuit diagram of FIG.
5 is a time chart for explaining changes in C1, B1, B2, and C2 pressures immediately after an operation to a range.

FIG. 16 is a table showing a relationship between a combination of operating states of a third solenoid valve and a fourth solenoid valve and an operation mode of a lock-up clutch.

FIG. 17 is a flowchart illustrating engagement characteristic change control performed at the time of shifting of the automatic transmission for a vehicle in FIG. 2;

18 is a time chart for explaining a change in the turbine wheel rotation speed and an operation of a lock-up clutch in the control of FIG. 17;

19 is a time chart showing the characteristics of the engagement hydraulic pressure and the output torque of the auxiliary transmission in the control of FIG. 17 and in the conventional case in comparison.

FIG. 20 is a flowchart illustrating release characteristic change control that is performed when the vehicle automatic transmission shown in FIG. 2 shifts.

[Explanation of symbols]

Reference Signs List 12 Fluid coupling (fluid transmission) 18 Sub-transmission C2 High speed clutch (hydraulic friction engagement device) 38 Lock-up clutch (direct coupling clutch) 133 Engine electronic control device (engine torque reduction control means) 156 3 solenoid valve, 158 fourth solenoid valve (switching electromagnetic solenoid valve device) 521 accumulator back pressure switching valve (engagement characteristic switching valve) 540 valve step S2 engine torque reduction determining means step S3 second release mode switching command means

──────────────────────────────────────────────────続 き Continued on front page (58) Field surveyed (Int.Cl. ⁶ , DB name) B60K 41/04 F16H 61/14

Claims (1)

(57) [Claims] 1. A vehicle equipped with a fluid transmission with a direct coupling clutch and an automatic transmission in which a gear is switched by engagement and disengagement of a hydraulic friction engagement device while the gear is being changed. Engine torque reduction control means for reducing the engine torque, and an engagement mode of the direct coupling clutch, which is usually used to release the direct coupling clutch
The first release mode is changed to release the direct clutch.
A second release mode used during gear change, and
Setting the coupling clutch in the first release mode and the second release mode
A switching electromagnetic solenoid valve device that outputs a switching signal pressure for switching to any of the rapid release modes for releasing more rapidly than the direct clutch. An oil pressure control device of a type that changes an engagement characteristic of a frictional engagement device to a gentler rise than usual, comprising: an engine torque reduction determining unit that determines execution of engine torque reduction control by the engine torque reduction control unit; A second release mode in which the switching electromagnetic solenoid valve device outputs a switching signal pressure for switching to the second release mode during a period in which the engine torque reduction determining means determines that the engine torque reduction control is being executed; Switching command means, a first position and a second position for making the engagement characteristics of the hydraulic friction engagement device gentler than the first position. An engagement characteristic that the valve element is located at the second position while the switching signal pressure for switching to the second release mode is being output from the electromagnetic solenoid valve device for switching. A hydraulic control device for a vehicle automatic transmission, comprising: a switching valve.