JPH03282057A - Oil pressure control device for continuously variable transmission for vehicle - Google Patents

Abstract

PURPOSE:To increase the coefficient of the increase of an engine rotation speed along with the increase of a car speed by providing a change gear ratio pressure generating means to generate a change gear ratio pressure which is changes in response to the change gear ratio of a continuously variable transmission and the rate of change of which is increased with the increase in an acceleration operation amount. CONSTITUTION:A change gear speed detecting valve 130 is formed such that when a throttle valve opening thetath is decreased, a rotation arm 137 is rotated counterclockwise and a displacement transmission ratio (distance ratio a/b) is decreased, and when the valve opening thetath is increased, the arm 137 is rotated clockwise and a displacement ratio (distance ratio a/b) is increased. For example, since, when a detecting rod 132 is pressed in, a flow rate of working oil fed through an orifice 138 by means of a first line oil passage 98 and discharged by a spool valve 136 is decreased, a working oil pressure on the downstream side of the orifice 138 is increased. A working oil pressure is a change gear ratio pressure Pgamma and is increased, as shown in the diagram, with the decrease in a change gear ratio gamma.

Description

DETAILED DESCRIPTION OF THE INVENTION Field of the Invention The present invention relates to a hydraulic control device for controlling the gear ratio of a continuously variable transmission for a vehicle.

Conventional technology In a vehicle equipped with a continuously variable transmission in which the gear ratio can be changed steplessly, a valve for controlling the gear ratio is equipped with an input shaft rotational speed pressure (pitot pressure) that represents the input shaft rotational speed. A hydraulic control device for a continuously variable transmission for a vehicle is configured to control the gear ratio of the continuously variable transmission by applying a thrust toward the speed increasing side based on the throttle valve opening and a thrust toward the decelerating side based on the throttle valve opening. Are known. For example, JP-A-55-65
This is the hydraulic control device described in Japanese Patent No. 755.

According to this, a hydraulic gear ratio control system is basically realized. However, according to this hydraulic gear ratio control system, the engine rotation speed remains constant regardless of changes in the gear ratio of the continuously variable transmission. There were drawbacks such as the rotation speed being too low.

On the other hand, for example, as described in Japanese Unexamined Patent Publication No. 61-48658, thrust toward the deceleration side is generated using line oil pressure that is controlled in relation to the throttle valve opening and the gear ratio. A speed change control valve equipped with a mechanism has been proposed. According to this, the engine rotation speed is increased as the gear ratio changes toward the deceleration side and as the vehicle speed increases.

Problems to be Solved by the Invention By the way, during normal driving of a vehicle, when the amount of acceleration operation of the accelerator pedal (throttle valve opening) is small, the engine rotational speed increases due to a change in the gear ratio to the speed increasing side (increase in vehicle speed). However, when the amount of operation of the accelerator pedal is large, it is necessary to increase the amount of increase in the engine rotational speed as the gear ratio changes to the speed increasing side in order to maintain the driving feel. Become. However, in the conventional hydraulic control device as described above, in a state where the throttle valve opening is large compared to a state where the throttle valve opening is small, the gear ratio changes to the speed increasing side, that is, the engine speed increases as the vehicle speed increases. This has the disadvantage that the amount of increase in rotational speed is small, giving a sense of discomfort to the driver.

That is, conventionally, a pitot tube is used to detect the input shaft rotational speed of a continuously variable transmission, and the pitot pressure output from this pitot tube is used for the speed change control valve.
Normally, this pitot pressure has a property that the rate of change increases as the input shaft rotational speed goes from a low region to a high region, and is changed in a multidimensional manner. In addition, with the speed change control valve in the conventional hydraulic control system, regardless of the throttle valve opening,
The range of change in pitot pressure for operating the valve element of the speed change control valve in a balanced manner is constant with respect to the range of change in gear ratio pressure. The range of change in the input shaft rotational speed obtained in response to the constant range of change in pitot pressure is large in a low pitot pressure region and small in a high pressure region, according to the characteristics of the pitot pressure. Therefore, when the vehicle is running, compared to a low throttle valve opening state where the input shaft rotational speed is relatively low, in a high throttle valve opening state where the input shaft rotational speed is relatively high, the gear ratio This means that the amount of increase in E of the engine rotational speed that accompanies a change to the acceleration side (increase in vehicle speed) becomes smaller.

The present invention has been made against the background of the above circumstances,
The purpose of this is to change the gear ratio toward the speed increasing side, that is, to increase the amount of increase in engine rotational speed as the vehicle speed increases, when the amount of acceleration operation is large compared to the condition where the amount of acceleration operation is small. An object of the present invention is to provide a hydraulic control device for a continuously variable transmission for a vehicle.

Means for Solving the Problems The gist of the present invention for achieving the object is to provide a vehicle equipped with a continuously variable transmission in which the gear ratio can be changed steplessly. A hydraulic control device for a continuously variable transmission for a vehicle for controlling (71)
(b) an input shaft rotational speed pressure having a magnitude corresponding to the input shaft rotational speed of the continuously variable transmission; and (C) a gear ratio pressure that changes in accordance with the gear ratio of the continuously variable transmission and whose rate of change increases as the acceleration operation amount increases. and (d) a valve for changing the gear ratio of the continuously variable transmission to an increasing side or a decelerating side, the gear ratio being changed to the decelerating side according to the required output pressure and the gear ratio pressure. and a speed change control valve that controls the gear ratio of the continuously variable transmission based on the thrust toward the speed increasing side and the thrust toward the speed increasing side due to the input shaft rotational speed pressure.

Operation and Effect of the Invention With this arrangement, the gear ratio pressure generated by the gear ratio pressure generating means changes in accordance with the gear ratio of the continuously variable transmission, and as the acceleration operation amount increases. Since the rate of change increases, the speed change control valve changes the continuously variable speed based on the thrust toward the deceleration side due to the gear ratio pressure and the required output pressure, and the thrust toward the speed increase side due to the input shaft rotational speed pressure. By controlling the gear ratio of the machine, when the throttle valve opening is large compared to when the throttle valve opening is small, the engine rotational speed is reduced due to a change in the gear ratio to the speed increasing side (increase in vehicle speed). The rise will be larger. Therefore, compared to driving with a low throttle opening, when driving with a high throttle opening, the change in the gear ratio to the speed increasing side, that is, an increase in the input shaft rotational speed, and an increase in the vehicle speed increases the engine speed. This increases the amount of increase in rotational speed and provides a good driving feel.

EXAMPLE Hereinafter, an example of the present invention will be described in detail based on the drawings.

In FIG. 2, the power of the engine 10 is transmitted through a fluid coupling 12 with a lock amplifier clutch and a belt type continuously variable transmission (hereinafter referred to as
The power is transmitted to drive wheels 24 connected to a drive shaft 22 via a CVT (CVT) 14, a forward/reverse switching device 16, a reduction gear device 18, and a differential gear device 20.

The fluid coupling 12 connects a pump impeller 28 connected to the crankshaft 26 of the engine IO, and an input shaft 3 of the CVT 14.
a turbine impeller 32 fixed at zero and rotated by oil from the pump impeller 28; a lock-up clutch 36 fixed to the input shaft 30 via a damper 34;
An engagement-side oil chamber 33 connected to an engagement-side oil passage 292 described below and a release-side oil chamber 3 connected to a release-side oil passage 294 described below.
5. The inside of the fluid coupling 12 is always filled with hydraulic oil, and for example, when the vehicle speed (■) exceeds a predetermined value, or when the difference in rotational speed between the pump impeller 28 and the turbine impeller 32 becomes less than a predetermined value, the engaging side oil is filled. As hydraulic oil is supplied to the chamber 33 and hydraulic oil flows out from the disengagement side oil chamber 35, the lock-up clutch 36 is engaged and the crankshaft 26 and the input shaft 30 are directly connected.
On the other hand, when the vehicle speed V in L falls below a predetermined value or the rotational speed difference exceeds a predetermined value, hydraulic oil is supplied to the disengagement side oil chamber 35 and hydraulic oil flows out from the engagement side oil chamber 33. As a result, the lock-up clutch 36 is released.

The CVT 14 has variable pulleys 40 and 42 of approximately the same diameter provided on its input shaft 30 and output shaft 38, respectively;
The transmission belt 44 is wound around the variable pulleys 40 and 42.

The variable pulleys 40 and 42 are fixed rotating bodies 46 and 48 fixed to the input shaft 30 and the output shaft 38, respectively, and are provided on the input shaft 30 and the output shaft 38 so as to be movable in the axial direction but not relatively rotatable around the axes. The movable rotating bodies 50 and 52 are
is moved by the primary side hydraulic cylinder 54 and the secondary side hydraulic cylinder 56 that function as hydraulic actuators, thereby changing the V groove width, that is, the hanging diameter (effective diameter) of the transmission belt 44, and changing the gear ratio r (
= rotational speed N+- of input shaft 30/rotational speed N of output shaft 38. ut) is now being changed. Since variable pulleys 40 and 42 have substantially the same diameter, I-distribution pressure cylinders 54 and 56 have similar pressure receiving areas. Typically, the pressure in the driven side of hydraulic cylinders 54 and 56 is controlled to control the tension in transmission belt 44.

The forward/reverse switching bag Y116 is a well-known double pinion type planetary gear mechanism, and includes a pair of planetary gears 62 and 64 that are rotatably supported by a carrier 60 fixed to its output shaft 58 and mesh with each other. A sun gear 66 that is fixed to the input shaft (output shaft of the CVT 14) of the forward switching device 16 and meshes with the planetary gear 62 on the inner peripheral side, a ring gear 68 that meshes with the planetary gear 64 on the outer peripheral side, and stops the rotation of the ring gear 68. Reverse brake 70 for
and the input shaft 3 of the carrier 60 and the forward/reverse switching device 16.
8. The reverse brake 70 and the forward clutch 72 are hydraulically actuated friction engagement devices, and when they are not engaged, the forward/reverse switching device i16 is placed in a neutral state and power transmission is cut off. . However, when the forward clutch 72 is engaged, the output shaft 3 of the CVT 14
8 and the output shaft 58 of the forward/reverse switching device 16 are directly connected to transmit power in the forward direction of the vehicle. Also, when the reverse brake 70 is engaged, the output shaft 38 of the CVTI 4 and the forward/reverse switching bag? ! ! Since the direction of rotation is reversed between the output shaft 58 and the output shaft 58 of No. 16, power in the backward direction of the vehicle is transmitted. −
As shown in section F, the forward/reverse switching device 16 is placed in the neutral state when the shift lever 252 (described later) is operated to the N range, so it
It functions as a clutch device that is rotated in series with the power transmission path from the drive shaft 24 to the drive shaft 24 and releases the power transmission path.

Note that annular members 78 and 8 are provided on the outer periphery of the fixed rotary bodies 46 and 48 of the CVT 14 to form annular grooves 74 and 76 that are open to the inner periphery so as to be filled with oil.
0 is provided in the annular grooves 74 and 76, and an input shaft rotational speed pressure PNis+ representing the rotational speed of the input shaft 30 and the output shaft 38 and an output shaft rotational speed pressure P N
Pitot tubes 82 and 84 for generating 611& are inserted, respectively. 3 and 4 show examples of output pressure characteristics of the pitot tubes 82 and 84, where the input shaft rotational speed pressure PNi. and output shaft rotational speed pressure P
If the angle is 8°□, the rotational speed of the input shaft 30 and the output shaft 38 increases and the stop becomes higher. In this example, E
The pitot tube 82 functions as a manual shaft rotation speed pressure generating means.

Further, a vehicle speed detection pump 86 for generating vehicle speed pressure Pv is connected to the shaft of the reduction gear device 18, and is driven to rotate at a speed corresponding to the vehicle speed. As shown in FIG. 6a, the vehicle speed detection pump 86
The discharged oil is made to flow back through throttles 88 and 90. Figure 5 shows the vehicle speed detection pump 86.
2 shows an example of the output pressure characteristics, and the vehicle speed pressure Pv increases as the vehicle speed V increases.

6a, 6b and 6c show in detail the hydraulic control circuit 89 of FIG. 2 for controlling the vehicle power transmission system.

The oil pump 92 constitutes the hydraulic pressure source of this hydraulic control circuit, and is integrally connected with the pump impeller 28 of the fluid coupling 12 so that the crankshaft 26
It is designed to be rotationally driven by. The oil pump 92 sucks the hydraulic oil that has returned into the oil tank (not shown) through the strainer 94 and returns it to the oil passage 9.
The hydraulic oil returned via 6 is sucked into the first line oil passage 9.
In this embodiment, the hydraulic oil in the first line oil passage 98 is returned to the oil passage 96 and the clutch pressure oil passage 10 by the overflow (relief) type first pressure regulating valve 15B.
By leaking to 0, the first line oil pressure pH in the first line oil passage 98 is regulated.

Note that the first line oil passage 98 is provided with a relief valve 99 that provides relief when the discharge amount from the oil pump 92 becomes excessive.

First, the first throttle pressure P1, the second throttle pressure PL
A signal pressure generation mechanism for generating bt-, gear ratio pressure P, . . . clutch pressure Pct will be explained.

The first throttle pressure P is a signal having a magnitude corresponding to the output torque of the engine 10, and is generated by the first throttle detection valve 102.

The first throttle detection valve 102 includes a cam 104 that is rotated together with a throttle valve (not shown), and a cam 104 that is rotated together with a throttle valve (not shown).
A plunger 106 that engages with the cam surface of No. 4 and is driven in the axial direction in relation to the rotation angle of the cam 104, a thrust in the valve opening direction applied from the plunger 106 via a spring 108, and a feed bank. The first line oil pressure pH is reduced by being positioned at a position where the first throttle pressure PLhl, which is the pressure, and the thrust in the valve closing direction by the spring 110 are balanced, and the first line pressure PLhl, which is the pressure, is positioned at a position where the thrust in the valve closing direction is balanced. 1 throttle pressure P is provided.

FIG. 7 shows the relationship between the first throttle pressure Pthl and the actual opening degree θ of the throttle valve. It is formed in consideration of the output of the engine IO so that the pressure Pthl is generated. The first throttle pressure Pthl is guided by a first throttle pressure oil passage 113.

The second throttle pressure P2 is a signal whose magnitude corresponds to the actual opening degree θ of the throttle valve provided in the engine IO or the operation amount of the accelerator pedal (not shown), and functions as a required output pressure generating means. 2nd throttle detection valve 1
14. 2nd throttle detection valve 1
14 includes a cam 116 that is rotated together with a throttle valve (not shown), a plunger 11B that engages with the cam surface of this cam 116 and is driven in the axial direction in relation to the rotation angle of this cam 116, and a spring 120. The thrust force in the valve opening direction is applied from the plunger 118 via the second throttle pressure P Lh!, which is the feedback pressure. By being positioned at a position where the thrust in the valve closing direction by the spring 122 is balanced, the first line oil pressure PI is reduced, and the second throttle pressure PI2 has a magnitude corresponding to the throttle valve opening θ. The spool valve 124 generates a spool valve. FIG. 8 shows the relationship between the second throttle pressure p thz and the actual opening θ of the throttle valve, and the cam curve of the cam 116 shows the speed ratio control that provides suitable fuel efficiency and running performance. A second throttle pressure P2 is generated to enable this.

The second throttle pressure P2 is the second throttle pressure oil passage 1.
26.

The gear ratio pressure P is a signal representing the gear ratio γ of the CVTI 4, and the gear ratio detection valve 1 functions as a gear ratio pressure generating means.
30. As shown in detail in FIG.
A spring 134 transmits an urging force corresponding to the position of the detection rod 132, and a first line hydraulic pressure PI receives a thrust in the valve closing direction from the spring 134.

The spool valve element 136 is provided with a spool valve element 136 that changes the discharge flow rate to the drain by receiving a thrust force in the valve opening direction from the spool valve element 136 and being positioned at a position where both thrust forces are balanced. The detection rod 132 has a pin 133 attached to one end of the variable fulcrum link 131.
The other end of the variable fulcrum link 131 is made to slide into an annular groove 135 formed on the outer periphery of the movable rotary body 50, and the detection rod 132 is rotatably connected to the movable rotary body 50. 50 is pushed in as it moves toward the speed increasing side. The variable fulcrum link 131 is, for example, a fulcrum pin 1 provided at the tip of a rotating arm 137 that is fixed to the rotating shaft of a cam 104, 116, which will be described later.
39 is fitted into an arc-shaped elongated hole 141, and when the throttle valve opening θ becomes small, the rotating arm 137 is rotated counterclockwise to change the displacement transmission ratio (the change from the fulcrum pin 139). However, when the throttle valve opening degree θ increases, the L rotary arm 137 is rotated clockwise to reduce the displacement transmission ratio (distance ratio a / b ) from the fulcrum pin 139. b) is made larger.

For example, when the detection rod 132 is pushed in, the flow rate of hydraulic oil supplied from the first line oil passage 98 through the orifice 138 and discharged to the drain by the spool valve 136 is reduced, so that the operation downstream of the orifice 138 is reduced. Oil pressure is increased. This working oil pressure is the gear ratio pressure P,
As shown in FIG. 9, it is increased as the gear ratio T decreases (changes toward speed increasing side).

In addition, a limit valve 1 is connected to the gear ratio pressure oil passage 140 that is connected to the output boat of the gear ratio detection valve 130 and guides the gear ratio pressure P.
42 are provided. As a result, the gear ratio pressure P is
In order to make the second line hydraulic pressure Plz have a polygonal line characteristic close to an ideal curve for tension control of the transmission belt 44, it is limited according to the magnitude of the first throttle pressure P Lh+. The limit valve 142 controls the first throttle pressure P
The plunger 144 generates a thrust in the valve closing direction by being actuated by the plunger 144, and the spring 146 generates a thrust in the valve closing direction. and a spool valve element 14B which is positioned in a position where both thrusts are balanced in response to the thrust of Drain oil to port 15
0 to limit pressure rise north of that point. Therefore, the limit value of gear ratio pressure P is,
The higher the first throttle pressure P Lhl is, the higher the first throttle pressure P Lhl is.

Furthermore, as the throttle valve opening θth increases, the displacement transmission ratio (=a/b) increases, so as shown in FIG. Pressure P increases.

Next, the first line oil pressure pH, the second line oil pressure P1t, and the third line oil pressure PI! 3 and the pressure regulating mechanism for the clutch oil pressure PCL will be explained.

The first pressure regulating valve 158 includes a spool valve element 160, a spring seat 162, a return spring 164, and a plunger 166, respectively. Spool valve 16
0 opens and closes between the boat 170a that communicates with the first line oil passage 98 and the boat 170b that communicates with the return oil passage 96 or the boat 170C that communicates with the clutch pressure oil passage 100; The first line oil pressure P11 as feedback pressure is throttled to the end face of 172 174
A chamber 176 is provided for acting through the
The spool valve element 160 is biased in the valve opening direction by this first line oil pressure pH. Spool bento 1
A chamber 182 to which the first throttle pressure PLhl is introduced is provided between the first land 17B and the second land 180 of the plunger 166, which are provided coaxially with the plunger 60, and
A chamber 184 into which the second line oil pressure P12 is guided is provided on the end surface side of the first land 17B. Since the biasing force of the return spring 164 is applied to the spool valve element 160 in the valve closing direction via the spring seat 162,
The pressure receiving area of the first land 172 of the spool valve 160 is A
I, the cross-sectional area of the first land 178 of the plunger 166 is A
3. If the cross-sectional area of the second land 180 is Aa, and the biasing force of the return spring 164 is W, then the spool valve 160
is balanced at a position where the following equation (1) holds true, and the first line oil pressure P11 is regulated. Note that a chamber 186 between the first land 172 and the second land 173 of the spool valve 160 is open to the drain and is at atmospheric pressure.

11- [PI3 ・A3+P1(A4-A3) +W) /
AI...
is controlled to be a predetermined pressure higher than the second line oil pressure P12 according to the second line oil pressure P12, while the characteristics of the second line oil pressure P12 are vertically translated in parallel according to changes in the first throttle pressure Pthl. As a result, even if the downstream side hydraulic cylinder 54 and the secondary side hydraulic cylinder 56 have the same pressure receiving area, not only a sufficient shift range can be obtained, but also sufficient shift responsiveness can be obtained. Line oil pressure P1
By controlling 1 to a necessary and sufficient value, the power...
loss is reduced. In addition, the input side hydraulic cylinder pressure P! is applied to the F chamber 184. The reason why R is not applied is to prevent the first line oil pressure pH from increasing divergently to the maximum value during sudden speed increases and the like.

Next, the second pressure regulating valve 1 generates the second line oil pressure P12.
The configuration of 90 will be explained. The second pressure regulating valve 190 includes a spool valve 194 that opens and closes between the first line oil passage 98 and the second line oil passage 192, a spring seat 196, a return spring 198, and a plunger 200. At the shaft end of the spool valve element 194, a first land 202, a second land 204, and a third land 206, each having a larger diameter in this order, are formed in this order. 2nd land 204 and 3rd land 206
2nd line hydraulic pressure PI! .
2, the valve is biased toward the valve closing direction. Further, on the end surface side of the first land 202 of the spool valve element 194, there is a chamber 216 into which the gear ratio pressure P is introduced via the throttle 214.
is provided, and the spool valve 194 has a gear ratio pressure P.

The valve is biased toward the valve closing direction.

Further, between the first land 202 and the second land 204 of the spool valve element 194, the output shaft rotational speed pressure P8°,
A chamber 218 is provided in which the spool valve 1 is guided.
94 is biased in the valve closing direction by the output shaft rotational speed pressure P.sub.lout. On the other hand, the second pressure regulating valve 19
0, the biasing force of the return spring 198 in the valve opening direction is applied to the spool valve element 194 via the spring seat 4196. Further, a chamber 208 for applying a first throttle pressure PLhl is provided on the end face side of the plunger 200, and the spool valve element 194 is urged in the valve opening direction by this first throttle pressure P th+. ing.

Therefore, if the pressure-receiving area of the first land 202 is A5, the cross-sectional area of the second land 204 is A4, the cross-sectional area of the third run F206 is A1, the pressure-receiving area of the plunger 200 is A8, and the biasing force of the return spring 198 is W, then The spool valve element 194 is balanced at a position where the following equation (2) holds true. That is, by moving the spool valve 194 according to equation (2), the first line oil pressure Pf
, is reduced and the second line oil pressure Pf is generated. Since the second line oil passage 192 is a relatively closed system, the second pressure regulating valve 190 lowers the second line oil pressure P12 to the tenth oil pressure by reducing the first line oil pressure PiI, which is a relatively high oil pressure. It is generated as shown by the solid line in the figure.

P l 2= ((As -Pth+ +W-8S' Pr(Ah As) PNouL) /(
(8, -8,) ・ ・ ・ ・ (2) As is clear from the above equation (2), the characteristics of the second line oil pressure P12 are limited by the transmission gear ratio pressure P and the control valve 142. 44, while the characteristics of the second line oil pressure Pit are vertically translated in response to changes in the first throttle pressure P.

Further, the second line oil pressure P12 also increases or decreases in response to changes in the output shaft rotational speed pressure P Haul. In this way, when the output shaft rotational speed pressure P Moat increases, the characteristics of the second line oil pressure P12 are moved in the decreasing direction, so it is preferable to increase the belt clamping force by the centrifugal oil pressure generated in the secondary hydraulic cylinder 56. will be compensated for. Note that an oil passage 322 that guides the hydraulic oil flowing out from the second pressure regulating valve 190 is connected to the input side of an oil cooler 314, which will be described later, and the second line oil pressure PI22 is smaller than the Tara pressure even if it is at its lowest. It is designed not to happen. Thereby, even if the second line oil pressure P12 is set to the minimum for centrifugal oil pressure compensation, a gas cavity is prevented from being formed in the secondary side hydraulic cylinder 56.

The third pressure regulating valve 220 generates an optimum third line oil pressure Pl for operating the reverse brake 70 and the forward clutch 72 of the forward/reverse switching device 16.

This third pressure regulating valve 220 includes a spool valve 222 that opens and closes between the first line oil passage 98 and the third line oil passage 221;
Spring seat 224, return spring 226,
and a plunger 228. Spool bento 2
Between the first land 230 and the second land 232 of No. 22, the third line oil pressure P13 is applied as feedback pressure to the restrictor 2.
A chamber 236 is provided through which the spool valve element 222 is urged in the valve closing direction by the third line oil pressure Pis. Furthermore, a chamber 240 is provided on the first land 230 side of the spool valve element 222 to which a gear ratio pressure P is introduced via a throttle 238.
The spool valve element 222 is biased in the valve closing direction by the gear ratio pressure P. Inside the third pressure regulating valve 220, a biasing force in the valve opening direction of a return spring 226 is applied to the spool valve element 222 via a spring seat 224. Further, a chamber 242 for applying a first throttle pressure P th+ is provided on the end face of the plunger 228, so that the spool valve element 222 is urged in the valve opening direction by this first throttle pressure P th+. It has become.

Further, a shift lever 25 is located between the first land 244 of the plunger 228 and the second land 246 having a larger diameter.
A chamber 248 is provided for guiding the R range pressure (=third line oil pressure PR1) output from the manual valve 250 when the manual valve 250 is operated to the R (reverse) range. Therefore, the third line oil pressure Pf3 is determined by the gear ratio pressure P and the throttle pressure Pub from a formula similar to the formula (1) above.
The pressure is regulated to an optimum value based on the characteristics shown in FIG. This optimal value is the forward clutch 7
2 or a necessary and sufficient value to ensure that torque can be transmitted without slipping in the reverse brake 70. Furthermore, when traveling in reverse, the third line hydraulic pressure Pf is guided into the chamber 248, so the force that urges the spool valve element 222 in the valve opening direction is increased, and the third line hydraulic pressure Pf3 is approximately doubled. The value is greatly increased. In other words, the characteristics shown in FIG. 11 are translated upward. Thereby, in the forward clutch 72 and the reverse brake 70, suitable torque capacities can be obtained respectively during forward movement and reverse movement.

The third line oil pressure PR1 regulated as described above is supplied to the manual valve 250. The manual valve 250 includes a spool valve 254 that is mechanically connected to and can be moved by a shift lever 252 of the vehicle.
When operating in a forward range such as the L (low) or D (drive) range, the third line oil pressure Pffi,
A DL range pressure equal to 1 is supplied from the output port 256 to the forward clutch 72, and at the same time, oil is allowed to drain from the reverse brake 70 to the drain. On the contrary, when the R range is operated, R range pressure equal to the third line oil pressure P13 is supplied from the output port 258 to the reverse brake 70, and at the same time, oil is allowed to drain from the forward clutch 72.
When the vehicle is operated to the N (neutral) range, oil is allowed to drain from both the forward clutch 72 and the reverse brake 70. Further, the manual valve 250 is provided with an output port 260 that outputs PRL range pressure when the shift lever 252 is operated to the R or P (parking) range. In addition, the accumulator 26
2 and 264 are connected to the forward clutch 72 and the reverse brake 70, respectively, in order to gradually increase the hydraulic pressure and smoothly advance the frictional engagement. In addition, the shift timing valve 266 initially opens the throttle 268 to increase the flow rate, as shown on the left side of the center line of the shift timing valve 266 in FIG. 6C, but the hydraulic pressure in the hydraulic cylinder of the forward clutch 72 Aperture 26 depending on height
8 to adjust the transient inflow flow rate as shown to the right of the centerline.

The hydraulic oil discharged from the port 170a of the first pressure regulating valve 15B to the clutch pressure oil passage 100 is transferred to the lock-up clutch 3 of the fluid coupling 12 by the clutch pressure regulating valve 270.
The lock-up clutch oil pressure Pct is adjusted to a pressure suitable for operating the lock-up clutch 6. That is, the clutch pressure regulating valve 270 includes a spool valve element 272 that receives clutch hydraulic pressure PcL as feedback pressure and is biased in the valve opening direction, and a spring 274 that biases the spool valve element 272 in the valve closing direction. The spool valve includes a chamber 276 to which a first throttle pressure P is supplied, and a plunger 278 that receives the first throttle pressure P of the chamber 276 and urges the spool valve element 272 in the valve closing direction. The valve element 272 is operated so that the thrust in the valve opening direction based on the feedback pressure is balanced with the thrust in the valve closing direction from the spring 274 and the plunger 278, thereby causing the hydraulic oil in the clutch pressure oil passage 100 to flow out. , the clutch oil pressure P corresponding to the output torque of the engine 10
cl is generated. The hydraulic oil in the clutch pressure oil passage 100 is passed through a throttle 280 etc. to the CVT as lubricating oil.
14 and the forward/reverse switching device 16. Further, the hydraulic oil discharged from the clutch pressure regulating valve 270 is returned to the return oil passage 96.

The clutch oil pressure Pcl regulated as described above is selectively supplied to the engagement side oil passage 292 and the release side oil passage 294 of the fluid coupling 12 by the lockup control valve 290,
The lock-up clutch 36 is configured to be in an engaged state or a released state. That is, the lock-up control valve 290 connects the clutch pressure oil passage 100 to the engagement side oil passage 2.
92 and the release side oil passage 294, a spring 298 that biases the spool valve 296 toward the release side, and a PRL range pressure (-3rd Line oil pressure P
f, ) is introduced, and the chamber 300 is drained in other ranges, and the spool valve 2 receives the PRL range pressure.
A plunger 302 that urges the valve 96 toward the release side, and a first valve provided between the spool valve 296 and the plunger 302
A chamber 304 that receives the signal pressure P S01+ and a chamber 30 that receives the DL range pressure and urges the spool valve element 296 toward the engagement side.
6. The first signal pressure P. 1. is generated by setting the downstream side of the throttle 310 to the same pressure as the DL range pressure (-third line oil pressure Pf3) when the normally closed first solenoid valve 308 is in the off state. When the first solenoid valve 308 is in the on state, the downstream side of the throttle 310 is drained and the first signal pressure P,
. ■ is made to disappear.

Therefore, in a state where the DL range pressure is being generated, when the first solenoid valve 308 is turned on, the first signal pressure PSO11 cannot be generated, and the spool valve 296 is positioned to the engagement side. It will be done. As a result, the clutch oil pressure Pct is supplied to the engagement side oil chamber 33 via the lockup control valve 290 and the engagement side oil passage 292, and at the same time, the hydraulic oil in the release side oil chamber 35 is supplied to the engagement side oil chamber 33 via the lockup control valve 290 and the engagement side oil passage 292. Drained through lock-up control valve 290. Conversely, when the first solenoid valve 308 is turned off, the first signal pressure P. 11 is generated, the spool valve 29
6 is positioned on the release side. As a result, the clutch oil pressure Pcl is supplied to the release side oil chamber 35 via the lockup control valve 290 and the release side oil passage 294, and at the same time,
The hydraulic oil in the maintenance side oil chamber 33 is drained through the maintenance side oil passage 292, the lockup control valve 290, the throttle 312, and the oil cooler 314.

When the PRL range pressure is being generated, the spool valve element 296 is preferentially positioned toward the release side, so that the lock-up clutch 36 is maintained in the released state. As a result, even when the first solenoid valve 308 is stuck in the on side or when the first solenoid valve 308 is constantly energized due to a failure in the electric circuit,
When the shift lever 252 is operated to the L range, the lock-up clutch 36 is released, allowing the vehicle to restart.

A Tara pressure control valve 316 is provided between the throttle 312 and the oil cooler 314 to maintain the input pressure of the oil cooler 314 constant.

This Tara pressure control valve 316 is provided with a spool valve element 318 for releasing to the drain and a spring 320 that biases this spool valve element 31B in the valve closing direction, and the thrust in the valve opening direction due to cooler pressure is provided. When the thrust force of the spring 320 in the valve closing direction is exceeded, the hydraulic oil on the input side of the oil cooler 314 is released to the drain.

Next, the switching valve 328, first shift control valve 330, second shift control valve 332, and associated valves related to shift control will be explained.

The switching valve 328 connects the primary oil passage 334 and the secondary oil passage 336, which are output oil passages of the first speed change control valve 330, to the primary hydraulic cylinder 54 and the secondary hydraulic cylinder 56, respectively. A neutral connecting the traveling position and the second outgoing oil passage 338 and the second outgoing oil passage 340, which are the output oil paths of the second speed change control valve 332, to the primary hydraulic cylinder 54 and the secondary hydraulic cylinder 56, respectively. spool valve 342 and spool valve 34 positioned at
2 towards its neutral position, and the spool valve 342 receives the DL range pressure.
A plunger 346 is provided between the plunger 346 and the spool valve element 342 and receives R range pressure for urging the spool valve element 342 toward the travel position. 348. This causes the shift lever 252 to shift to D, L, R.
When the range is being operated, the first shift control valve 330
is enabled and operated to the N range, the second
Shift control valve 332 is enabled. Switching valve 3 in Figure 6C
The left side of the center line at 28 shows the spool valve 342 in the neutral position, and the right side of the center line shows the spool valve 342 in the running position (
D, L range position).

As shown in detail in FIG. 12, when the first speed change control valve 330 is located at the deceleration side position, it connects the first oil passage 334 to the drain oil passage 335, and at the same time connects the first oil passage 334 to the drain oil passage 335. When the passage 336 is connected to the first line oil passage 98 and positioned at the speed increasing side position, the first line oil passage 334 is connected to the first line oil passage 98.
A spool valve element 350 connects the next oil passage 336 to the first line oil passage 98 and at the same time connects the next oil passage 336 to the second line oil passage 192. An upper neutral spring 352 that once biases, a lower neutral spring 354 that biases the spool valve element 350 toward the deceleration side, and a spring receiver 356 that receives the end of the gain change spring 360 are integrally provided. , when the first signal pressure P5゜■ acts, the return spring 35
8 is moved in the direction away from the spool valve element 350, but when the first signal pressure Pi6Ll is not acting, it is moved toward the spool valve element 350 side according to the urging force of the return spring 358. A piston 362 which enables the biasing operation of the gain change spring 360 by sandwiching it between the spool valve element 350 and the spring receiving part 356, and a lower neutral spring 354 side (small diameter side) of the spool valve element 350. ) a second throttle pressure P th2 and a corrected gear ratio pressure P,
, °, whichever is higher, a chamber 364;
The input shaft rotational speed pressure P H4n is provided at the end of the upper neutral spring 352 side (large diameter side) of the spool valve element 350.
and a chamber 368 provided between the large diameter part and the small diameter part of the spool valve element 350 to receive the gear ratio pressure Pr.

The left side of the center line of the first speed change control valve 330 in FIG. 6b shows the spool valve element 350 in the speed increase side position, and the right side of the center line shows the spool valve element 350 in the deceleration side position. A V-shaped notch 369 is formed in the land of the spool valve 350 to smoothly change the opening degree of the spool valve 350.

The second throttle pressure P2 supplied to the first shift control valve 330 passes through the throttle 370 and the accumulator 37.
2, the rise is moderated. The accumulator 372 includes a stepped piston 371 having a large diameter portion and a small diameter portion, and a pressure accumulation chamber 373 provided on the end face side of the large diameter portion of the stepped piston 371 and connected to the first speed change control valve 330 side of the throttle 370. , a back pressure chamber 375 provided on the outer peripheral side of the small diameter portion of the stepped piston 371 and connected to the second throttle detection valve 114 side of the throttle 370 , and a spring that biases the stepped piston 371 toward the pressure accumulation chamber 373 side. 377. Since the second throttle pressure P2 which does not pass through the throttle 370 is supplied to the back pressure chamber 375, as shown in FIG. Relaxation starts from the value. In FIG. 13, the diagonal solid line shows the part where the rise is relaxed by the accumulator 372, the solid line shown on the upper side shows the case where the amount of depression of the accelerator pedal is large and the rise width is large, and the lower part shows the case where the rise is large. The solid line shown is
This shows a case where the accelerator pedal's Xli included 1 is not so large and the rising width is small. Due to the second throttle pressure P2 whose rise is moderated by the accumulator 372, the change in the gear ratio T is made gradual even during sudden acceleration operations, and the proportion of torque consumed to increase the rotation of the engine 10 is reduced. In addition, the proportion consumed as driving force of the vehicle is increased, and the feeling of acceleration of the vehicle is improved.

In addition, the corrected gear ratio pressure P,'' is adjusted by a pressure reducing valve 374 using the PRL range pressure as the source pressure, as shown in FIG. There is one in which the gear ratio γ is increased with an increase in the gear ratio pressure P, and the increase in the gear ratio γ is restricted in the speed increase side region from the above-mentioned constant value T0.Pressure reducing valve 3
74 is the corrected gear ratio pressure P,, ° as a feedback pressure.
A spring 378 is sandwiched between the spool valve element 376, which receives thrust in the valve closing direction, and the spring 378, which applies a thrust force in the valve opening direction to the spool valve element 376. The plunger 380 increases the urging force of the spring 378 as P increases, but the plunger 380 becomes immovable toward the spool valve element 376 by abutting against the valve body at a predetermined position. The flat portion shown by the solid line in FIG. 14 is a state in which the movement of the plunger 380 is stopped, and the gear change corresponding to the gear ratio pressure P of the magnitude that brings the plunger 380 into contact with the valve body is shown. The ratio γ is the above constant value γ. It is.

The or valve 382 is a ball valve 384 that is moved by the differential pressure between the second throttle pressure P2 and the corrected gear ratio pressure P1° and closes the boat to which the lower pressure is supplied.
The higher of the second throttle pressure P2 and the corrected gear ratio pressure Pr° is applied to the first gear change control valve 3.
30 chambers 364. That is, the corrected gear ratio pressure P
While r° changes according to the gear ratio γ, the second throttle pressure P Llt2 changes according to the throttle valve opening θ, so in a region where the second throttle pressure P Llt2 is small,
The corrected gear ratio pressure Pr ° of the part located above the second throttle pressure P in FIG. 15 is the second throttle pressure P Lk
! Instead, it is supplied to the first speed change control well 330.

Therefore, in the first shift control well 330, for example, when the accelerator pedal is depressed while driving in the L, or R range and the second throttle pressure P2 is increased, the spool valve 350 is The thrust in the direction of moving the engine toward the deceleration side is increased, and the rotational speed of the engine 10 is increased.
The rotational speed of the input shaft 30 and the input shaft rotational speed pressure P N
IT+ is increased. When the input shaft rotational speed pressure P Nin is increased in this manner, the thrust force in the direction of moving the spool valve element 350 toward the speed increasing side is increased. Further, as described above, when the accelerator pedal is depressed, the output torque of the engine IO is increased, so the speed of the vehicle increases in accordance with the amount of increase in the output torque, and the spool valve 350
Thrust in the direction of moving the motor toward the speed increasing side is sequentially applied. The spool valve element 350 is positioned at a position where the thrust in the deceleration direction and the acceleration direction are balanced, and the gear ratio T is controlled.
As shown in FIG. 6, the input shaft rotational speed N=n is increased. As shown in FIG. 9, the gear ratio pressure P is set to increase as the throttle valve opening θ increases. Therefore, at a constant throttle valve opening θ, the gear ratio T is at its maximum. Gear ratio γ. Minimum gear ratio γ from 1X
14, the range of increase in the rotational speed of the human power shaft ΔN i
n increases as the accelerator pedal operation amount increases, in other words, as the throttle valve opening degree θ increases. Due to the control operation of the first speed change control valve 330, the rotational speed N of the input shaft 30 is increased.

is controlled along an optimal curve that takes fuel efficiency and drivability into consideration.

As a result, the first shift control valve 330 substantially operates at a target input shaft rotational speed (optimum value taking fuel efficiency and drivability into consideration) determined based on the second throttle pressure P2 and vehicle speed pressure PVLi#. ) N, fi'' and the input shaft rotational speed pressure P Min representing the actual input shaft rotational speed N, ゎ.

When L or R range is selected, pressure reducing valve 3
The source pressure is supplied to 74 to correct the gear ratio pressure P.

is generated, and this corrected gear ratio pressure P.

and the second throttle pressure P th2, whichever is higher, is supplied to the chamber 364 of the first shift control valve 330. That is, in a region where the throttle valve opening θ is small, the corrected gear ratio pressure P, except when the gear ratio γ is near the maximum value,
,° are supplied to the chamber 364. Therefore, in FIG. 16, the throttle valve opening θ is small and the gear ratio T
A characteristic is obtained in which the area where is near the maximum value is blank.

For example, the curve corresponding to throttle valve opening θth from 0% to 60% is removed, and the throttle valve opening θth is 70%.
is represented by the curve corresponding to This allows L or R
Sufficient engine braking action can be obtained during range driving.

Further, in the first shift control valve 330, a first signal pressure P5° for releasing the lock-up clutch 36;
is being supplied, the piston 362 is moved against the urging force of the spring 358, so the end of the gain change spring 360 is separated from the spring receiver 356 or the spool valve 350, and the 17th The spool valve 350 is moved according to the characteristics shown by the solid line in the figure.

However, when the first signal pressure P soz is not supplied, the piston 362 is moved according to the biasing force of the spring 358, and the gain change spring 360 is activated. The gain change spring 360 is compressed only in the region where the spool valve element 350 moves from the neutral position toward the deceleration side, thereby biasing the spool valve element 350 toward the speed increasing side. In the region on the deceleration side, the spool valve 350 is also moved according to the characteristics shown by the one-dot chain line in FIG. As a general property of the CVT 14, the lower the circumferential speed of the transmission belt 44 (''=, vehicle speed or input shaft rotational speed N, 7), the longer the shift response time becomes, and the speed change response time becomes longer to a certain extent when the actual gear ratio T is on the deceleration side. However, the influence of the circumferential speed of the transmission belt 44 is the greatest.For this reason, in this embodiment, the low vehicle speed region where the circumferential speed of the transmission belt 44 is low is generally defined as a vehicle speed region of about 30 km/h or less. The determination is made using the first signal pressure P, which is generated to turn the lock-up clutch 36 off.In other words, the low vehicle speed region where the lock-up clutch 36 is turned off is determined. Now, with the characteristics shown by the solid line in Figure 17, the spool valve 3 is
50 is driven, while the lock-up clutch 3
In the high vehicle speed range where the switch 6 is on, the spool valve 350 is driven with the characteristics shown by the dashed line in FIG. 17, so the gain of the shift control in the high vehicle speed range is reduced. As a result, a suitable gain is obtained in the low vehicle speed region where the shift control response is insensitive, and in the high vehicle speed region where the shift control response is too sensitive, the gain is set lower than the gain in the low vehicle speed region. The shift control response is suitably maintained over the entire range including the vehicle speed range. In addition, in the above high vehicle speed region,
Although the gain is made small only on the deceleration side of the spool valve element 350, since the responsiveness is inherently low on the speed increase side, there is no problem even if the gain is made small only on the deceleration side.

Note that the drain oil passage 335 is provided with a drain prevention valve 400 whose opening and closing are controlled by a second signal pressure P sat □ generated in connection with the operation of the second electromagnetic valve 4IO. The drain prevention valve 400 prevents overspeeding of the engine 10, prevents excessive engine braking during ABS control, and closes the primary hydraulic cylinder 54 when the vehicle stops before the gear ratio γ reaches its maximum. It was established for this purpose. This drain prevention valve 400 includes a spool valve 402 that opens and closes between the drain oil passage 335 and the drain.
, a spring 404 that biases this spool valve element 402 in the valve opening direction, and a third line oil pressure PR1 that constantly acts on the spool valve element 402 to bias it in the valve closing direction.
and a second signal pressure P5 for acting on the spool valve element 402 to urge it in the valve opening direction.
A chamber 408 for receiving oL2 is provided. The second solenoid valve 410 closes the drain prevention valve 400 by opening and closing the downstream side of the throttle 412 in its ON state and eliminates the second signal pressure P5otz, but in its OFF state, the second solenoid valve 410 closes the drain prevention valve 400 downstream of the throttle 412. 3rd line hydraulic PA on the side.

A second signal pressure P sol□ equal to is generated to open the drain prevention valve 400.

As shown in detail in FIG. 18, when the second shift control valve 332 is located at the deceleration side position, the second shift control valve 332
The second secondary oil passage 340 is connected to the first line oil passage 98 at the same time as the next oil passage 338 is connected to the drain boat 41B. 338 to the first line oil passage 98, and at the same time
A spool valve element 420 that connects the secondary oil passage 340 to the second line oil passage 192, and an upper neutral spring 422 that biases the spool valve element 420 toward the speed increasing side (lower side in the figure).
, a lower neutral spring 424 that urges the spool valve element 420 toward the deceleration side (upper side in the figure), a plunger 426 inserted between the upper neutral spring 422 and the spool valve element 420, and a spool valve element The output shaft rotation speed pressure P is provided at the lower neutral spring 424 side end of the child 420.
A chamber 428 for receiving Nout and a plunger 426
The vehicle speed pressure P is provided on the upper neutral spring 422 side.

The receiving chamber 430 and the large diameter part of the plunger 426 (
A chamber 432 is provided between the spool valve element 420 side) and the small diameter portion (upper neutral spring 422 side) to receive and receive the DL range pressure.

Note that the land of the spool valve 420 is also formed with a square notch 434 for smoothly changing the opening degree of the spool valve 420.

For this reason, as in coasting, the shift lever 252 is in the N position.
When the forward/reverse switching device 16 is in a neutral state (non-power transmission state) by being operated in the range, the second shift control valve 332 generates a thrust toward the speed increasing side based on the vehicle speed pressure Pv and an output shaft. Rotational speed pressure P No
By positioning the spool valve 420 so that the thrust toward the deceleration side based on ut is balanced, the CVT1
The gear ratio γ of 4 is controlled so that the rotational speed difference between the output shaft 38 of the CVT 14, which is the input shaft of the forward/reverse switching device 16, and the output shaft 5 of the forward/reverse switching device 16 is eliminated. However, when the shift lever 252 is operated to the D or L range, the DL range pressure is supplied to the chamber 432, and the plunger 426 and the spool valve element 420 are forcibly positioned to the speed increasing side. As a result, the spool valve element 420 is prevented from moving regardless of the effects of the vehicle speed pressure Pv and the output shaft rotational speed pressure P Nout, and the spool valve element 420 is prevented from sliding into the hydraulic oil due to the sliding movement of the spool valve element 420 when unnecessary. In addition to suppressing the generation of powder, the spool valve element 420 is prevented from sticking in the deceleration side position, and over-speeding of the engine 10 when operated to the D or L range in the deceleration side position is suitably prevented. It has become so.

Note that the secondary hydraulic cylinder 56 and the second line oil passage 192
A bypass oil passage 440 including a one-way valve 43G and a throttle 438 in parallel is provided between the two. This one-way valve 436 allows flow from the second line oil passage 192 toward the secondary hydraulic cylinder 56, but blocks flow in the opposite direction.

Returning to FIG. 2, the engine 10 includes an engine rotation speed sensor 444 that detects the engine rotation speed N8 and supplies it to the electronic control unit 450, and an engine rotation speed sensor 444 that detects the throttle valve opening θ and supplies it to the electronic control unit 450. A throttle sensor 446 is provided for each. Further, the differential gear device 20 is provided with a vehicle speed sensor 448 that detects the vehicle speed V and supplies the detected vehicle speed V to the electronic control device 450. The electronic control device 450 includes a CPU 452, a ROM 454, and a RAM 456.
, and a so-called microcomputer including an interface circuit (not shown). CPU452 is RAM4
56, the input signal is processed according to a program stored in advance in the ROM 454, and a signal for driving the first solenoid valve 30B and the second solenoid valve 410 is output.

For example, in the lock-up clutch control routine, engagement or disengagement of the lock-up clutch 36 is determined based on the vehicle speed and the throttle opening θ from a diagram stored in the ROM 454 in advance.
When it is determined that the lock-up clutch 36 is engaged, a drive signal to turn on the first solenoid valve 308 is output, but when it is determined that the lock-up clutch 36 is released, the first solenoid valve 308 is turned on.
A drive signal for turning 08 off is output.

In addition, in the deceleration shift prevention routine, it is determined whether the actual engine rotation speed N exceeds a pre-stored overspeed judgment reference value, and if it is determined that it has exceeded it, the drain oil passage 335 is closed. Therefore, a drive signal that turns on the second solenoid valve 410 is output. As a result, engine 1
0 over-rotation is prevented. Also, when the drive wheels stop, in other words, when the vehicle speed becomes zero, it is determined whether the gear ratio T of the CVT 14 has reached its maximum value γ, □, and the gear ratio γ reaches its maximum value. If it is determined that γ0.8 has not been reached, the second
A drive signal that turns on the solenoid valve 410 is output.

As a result, when the vehicle stops, the switching valve 328,
The hydraulic oil in the primary hydraulic cylinder 54 is prevented from flowing out through the first speed change control valve 330 and the drain oil passage 335 in this order. In addition, if the vehicle is equipped with an ABS control device, it is determined whether or not ABS control is being performed so that recovery of the rotational speed of the drive wheels 24 is not inhibited by excessive engine braking during ABS control. and ABS
If it is determined that the control is in progress, a drive signal is output to turn on the second solenoid valve 410 in order to prevent further speed change to the deceleration side.

In the present embodiment, the first gear change control well 330 controls the gear ratio γ of the CVT 14 based on the second throttle pressure PL), 2, the gear ratio pressure P1, and the input shaft rotational speed pressure PNi11. The second shift control valve 332 operates based on the output shaft rotational speed pressure P Hout and the vehicle speed pressure Pv.
Rotational speed N of the output shaft 38 of the CVTI 4, which is the input shaft of the forward/reverse switching device 16. The gear ratio T is controlled so that the difference in rotational speed between ut and the rotational speed of the output shaft 58 of the forward/reverse switching device 16 becomes small. Then, shift lever 252 is turned to L.
When the operation is performed to the range, the first
Shift control: While the output pressure of the valve 330 is applied to the primary side hydraulic cylinder 54 and the secondary side hydraulic cylinder 56 of the CVT 14, when the shift lever 252 is operated to the N range, the switching valve 328 controls the second gear shift. control valve 332
The output pressure is applied to the primary hydraulic cylinder 54 and secondary hydraulic cylinder 56 of the CVT] 4. For this reason,
This eliminates the need for a throttle sensor, vehicle speed sensor, shift position sensor, computer, electromagnetic valve type gear shift control valve device, etc. for gear ratio control, making the hydraulic control device significantly cheaper.

Here, in this embodiment, as shown in FIG. 9 described above, the gear ratio pressure P is set to increase as the throttle valve opening degree θ increases. When the degree θ is high, the gear ratio T changes from the maximum gear ratio Teamx to the minimum gear ratio T. , fi, the input shaft rotational speed increase width ΔN i n becomes larger as the accelerator pedal operation amount is larger, in other words, as the throttle valve opening θ is larger. As shown in the control characteristics of the speed change control well 330, the characteristics showing the change in the input shaft rotational speed N r n with respect to the vehicle speed V are as follows.
When the throttle valve opening degree θL6 is constant, the rotational speed N in of the input shaft 30 increases as the vehicle speed pressure P Vli# increases, that is, the vehicle speed ■ increases, and in addition,
The amount of increase ΔN. The gear ratio T is controlled such that the larger the throttle valve opening θ is, the larger the gear ratio T becomes. As a result, when the throttle valve opening degree θ is constant, the input shaft rotation speed N.

Compared to the case where the characteristic of no change is used, the engine rotation speed near low vehicle speeds is kept relatively low, noise is suppressed, and the engine rotation speed increases as the vehicle speed increases, and the amount of accelerator pedal operation is reduced. The larger the input shaft rotational speed is, the larger the increase width ΔN i in the input shaft rotational speed becomes, so that a suitable driving feeling can be obtained.

Furthermore, according to this embodiment, as shown in FIG. 9, the corner position of the gear ratio pressure P is formed near the substantially same gear ratio, so as shown in FIG. 10, the second line oil pressure Pj22 Since the bending point position is similarly formed near the same gear ratio, there is an advantage that the second line oil pressure P12 can be brought closer to the ideal curve shown by the broken line. Incidentally, in the hydraulic control device described in Japanese Patent Application No. 1335971 filed earlier by the present applicant, as shown in FIG. 19, as the throttle valve opening θ changes, the gear ratio pressure P1 Since the corner position of P12 shifts to the speed increase side of the gear ratio T, the corner point of the second line oil pressure P12 similarly shifts to the speed increase side of the gear ratio γ as the throttle valve opening θ changes. It is.

Although one embodiment of the present invention has been described above based on the drawings,
The invention has utility in other aspects as well.

For example, in the above-mentioned embodiment, the gear ratio pressure P and the second throttle pressure ptw are used as the oil pressure signal pressure that is the basis of the gear shift control.
or corrected gear ratio pressure Pr °, whichever is higher, was used, but instead of that gear ratio pressure P, the second
Line pressure P! 2 or vehicle speed pressure Pv may be used.

In short, any oil pressure that changes in relation to the gear ratio is sufficient.

The gear ratio pressure described in the claims should be interpreted in this broad sense.

In addition, in the above-mentioned embodiment, vehicle speed pressure P,
A vehicle speed detection pump 86 was used that generates
A well-known governor valve that outputs a vehicle speed pressure P corresponding to the rotational speed of the output shaft 38 of the CVT 14 may be used.

Further, in the above-described embodiment, the rotating arm 137 is connected to the cams 104 and 11 for detecting the throttle valve opening degree θ.
Although it was provided on the rotation shaft of No. 6, it may be provided independently. In short, the gear ratio pressure Pr only needs to be mechanically related to the accelerator pedal and rotated in accordance with the required output value of the vehicle.

Furthermore, in the above embodiment, the first throttle pressure Pth and the second throttle pressure Pth2, which change in accordance with the throttle valve openings θ and h, are used as the required output pressure of the vehicle. A hydraulic signal that varies in relation to the amount of pedal operation may be used.

The above-mentioned embodiment is merely one embodiment of the present invention, and various modifications may be made to the present invention without departing from the spirit thereof.

[Brief explanation of drawings]

FIG. 1 is a diagram illustrating a main part of the embodiment shown in FIG. 2. FIG. FIG. 2 is a block diagram illustrating the configuration of a vehicle power transmission device including an embodiment of the present invention. FIG. 3 is a diagram showing the change characteristics of the input shaft rotational speed pressure generated by the pitot tube of FIG. 2. FIG. FIG. 4 is a diagram showing the change characteristics of the output shaft rotational speed pressure generated by the pitot tube of FIG. 2. FIG. FIG. 5 is a diagram showing the change characteristics of the vehicle speed pressure generated by the vehicle speed detection pump of FIG. 6a. Figure 6a, 6th
FIG. b and FIG. 6c are diagrams showing the hydraulic control circuit of the embodiment of FIG. 2 in detail. FIG. 7 is a diagram showing the change characteristics of the first throttle pressure generated by the first throttle detection valve of FIG. 6b. FIG. 8 is a diagram showing the change characteristics of the second throttle pressure generated by the second throttle detection valve of FIG. 6b. FIG. 9 is a diagram showing the change characteristics of the gear ratio pressure generated by the gear ratio detection valve and the limit valve of FIG. 6a. FIG. 10 is a diagram showing the change characteristics of the second line oil pressure generated by the second pressure regulating valve of FIG. 6a. FIG. 11 is a diagram showing the change characteristics of the third line oil pressure generated by the third pressure regulating valve of FIG. 6b. FIG. 12 is an enlarged view illustrating the first speed change control valve in FIG. 6b. FIG. 13 is a diagram illustrating the second throttle pressure relaxation effect of the accumulator. FIG. 14 is a diagram showing the change characteristics of the corrected gear ratio pressure generated by the pressure reducing valve of FIG. 6b. FIG. 15 is a diagram showing the characteristics obtained by the speed ratio control operation of the corrected speed ratio pressure valve that is applied to the first speed change control valve. FIG. 17 is a diagram showing the characteristics of the first shift control valve shown in FIG. 12. Figure 18 is 6b
FIG. 2 is an enlarged diagram illustrating the second speed change control valve shown in the figure. 1st
Figure 9 shows the change characteristics of the gear ratio pressure in the conventional invention and the second
FIG. 3 is a diagram showing a comparison of control characteristics of pressure regulating valves. 10: Engine 14: Belt type continuously variable transmission (continuously variable transmission) 82: Pitot tube (input shaft rotation speed pressure generation means) 114: Second throttle pressure detection valve (required output pressure generation means)

Claims (1)

[Scope of Claims] A hydraulic control device for a continuously variable transmission for controlling the gear ratio of the continuously variable transmission in a vehicle equipped with the continuously variable transmission whose gear ratio can be changed steplessly. a required output pressure generating means for generating a required output pressure having a magnitude corresponding to the acceleration operation amount of the vehicle; and generating an input shaft rotational speed pressure having a magnitude corresponding to the input shaft rotational speed of the continuously variable transmission. and a gear ratio that generates a gear ratio pressure that changes in accordance with the gear ratio of the continuously variable transmission and whose rate of change increases as the acceleration operation amount increases. pressure generating means; and a valve for changing the gear ratio of the continuously variable transmission to an increasing side or a decelerating side; A hydraulic control device for a continuously variable transmission for a vehicle, comprising: a speed change control valve that controls a gear ratio of the continuously variable transmission based on a thrust toward a speed increasing side due to speed pressure.