JPH1194062A - Hydraulic controller of toroidal-type continuously variable transmission - Google Patents

Abstract

PROBLEM TO BE SOLVED: To run a vehicle while restraining shifting shock in case of a sudden shifting caused by damage of a hydraulic system. SOLUTION: A hydraulic controller is provided with a trunnion pinched between input and output discs 28, 29 and for supporting power rollers 30, 31 so as to be freely tilted and rolled, a hydraulic cylinder for driving the trunnion in the axial direction, a step motor 152 and a shift control valve for controlling oil pressure to the hydraulic cylinder on the basis of line pressure so as to obtain the target gear ratio determined according to the operating state of a vehicle, and a line pressure solenoid valve 528 for controlling line pressure to the shift control valve, and a set value of line pressure is decreased when the actual shifting speed exceeds a specified value.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to an improvement in a hydraulic control device for a toroidal type continuously variable transmission employed in a vehicle or the like.

[0002]

2. Description of the Related Art As a hydraulic control device of a toroidal type continuously variable transmission used for a vehicle, for example, Japanese Patent Application Laid-Open No. 8-233 is disclosed.
No. 083 is known.

In a toroidal-type continuously variable transmission, gear shift control and torque transmission are performed by hydraulic pressure applied to a hydraulic servo cylinder supporting a trunnion, so that a normal hydraulic pressure can be obtained when the hydraulic piping system is damaged. Otherwise, the shift control becomes impossible.

If normal hydraulic pressure cannot be obtained due to damage to the hydraulic system and the target gear ratio cannot be maintained, the gear ratio mode is forcibly switched to change the gear ratio to the maximum Lo gear ratio (= maximum reduction gear ratio). It is set to enable traveling.

[0005]

However, in the above conventional example, it is determined that the target gear ratio cannot be maintained even if the control valve of the hydraulic circuit is stuck or a control program malfunctions, and the gear ratio is forcibly increased. Since it is set to Lo, there is a possibility that a sudden shift occurs during normal traveling and a large shift shock occurs.

Accordingly, the present invention has been made in view of the above-described problems, and suppresses a shift shock when a sudden shift occurs due to damage to a hydraulic system, malfunction of control means, and the like.
Enables the vehicle to run.

[0007]

According to a first aspect of the present invention, there is provided a trunnion which is held by an input / output disk and supports a power roller so as to be tiltable, a hydraulic cylinder which drives the trunnion in an axial direction, and operation of a vehicle. A shift control unit that controls a hydraulic pressure to the hydraulic cylinder based on a line pressure so as to have a target gear ratio determined according to a state; and a line pressure control unit that supplies a predetermined line pressure to the shift control unit. An actual shift speed detecting means for detecting an actual shift speed, and a set value of the line pressure control means when the detected actual shift speed exceeds a predetermined value. Line pressure lowering means for lowering the pressure.

According to a second aspect of the present invention, there is provided a trunnion which is supported by an input / output disk and tiltably supports a power roller, a hydraulic cylinder which drives the trunnion in an axial direction, and a driving state in accordance with a driving state of the vehicle. An actual transmission speed detection unit for detecting an actual transmission speed, in a hydraulic control device for a toroidal-type continuously variable transmission, the transmission control unit comprising: a transmission control unit that controls a hydraulic pressure to the hydraulic cylinder so as to have the determined target transmission ratio. A pressure difference reducing means for reducing or eliminating the differential pressure across the piston constituting the hydraulic cylinder when the detected actual gear speed exceeds a predetermined value.

In a third aspect based on the second aspect, the hydraulic cylinder includes a speed-increasing oil chamber and a deceleration-side oil chamber defined by a piston. The front-rear differential pressure is reduced by connecting the high-speed oil chamber to the deceleration-side oil chamber or the deceleration-side oil chamber to the speed-increasing oil chamber.

In a fourth aspect based on the second aspect, the hydraulic cylinder includes a speed-increasing oil chamber and a deceleration-side oil chamber defined by a piston, and the front-rear differential pressure reducing means increases the pressure difference. The line pressure is supplied to the speed-side oil chamber and the deceleration-side oil chamber, respectively, so that the front-rear differential pressure is reduced.

In a fifth aspect based on the second aspect, the hydraulic cylinder includes a speed-increasing oil chamber and a deceleration-side oil chamber defined by a piston, and the front-rear differential pressure reducing means increases the pressure difference. By connecting the high-speed oil chamber and the deceleration-side oil chamber to the tank, the pressure difference between the front and rear is reduced.

[0012]

Therefore, in the first invention, normal hydraulic pressure cannot be obtained due to damage to the hydraulic system, malfunction of the control means, etc.
When the target gear ratio cannot be maintained and a sudden gear shift is detected in which the actual gear speed exceeds a predetermined value, the set value of the line pressure control means is reduced, thereby maximizing the hydraulic pressure supplied to the hydraulic cylinder via the gear shift control means. Values will be regulated,
The speed change of the toroidal type continuously variable transmission is performed based on the axial displacement of the trunnion, and the torque transmission capacity of the toroidal type continuously variable transmission is hydraulic pressure against the traction force of the power roller, that is, a hydraulic cylinder. Is determined by the pressure difference between the front and rear of the piston, the torque transmission capacity of the continuously variable transmission is also reduced by reducing the line pressure.
For example, when performing a rapid shift (downshift) to the Lo side,
When the torque transmission capacity of the continuously variable transmission is sufficiently large, the magnitude of the shift shock (deceleration shock) is determined according to the speed of shifting to the Lo side and the magnitude of the inertia moment of the input system. By lowering the line pressure, it is possible to reduce the shift shock that occurs, and furthermore, even if an abnormality occurs in the hydraulic system or the control means that the shift is not possible, the shift based on the reduced line pressure is performed. By transmitting the minimum necessary torque, the vehicle can run and fail-safe of a vehicle equipped with a toroidal-type continuously variable transmission can be secured.

According to a second aspect of the present invention, during a rapid shift in which the actual shift speed exceeds a predetermined value, the differential pressure between the front and rear of the hydraulic cylinder is reduced or the differential pressure is eliminated, and the traction force of the power roller is countered. Oil pressure decreases or becomes zero
Therefore, the torque transmission capacity of the continuously variable transmission decreases sharply, and the differential pressure across the hydraulic cylinder becomes extremely small.
Alternatively, a rapid shift cannot be performed due to the disappearance of the differential pressure, and it is possible to suppress the occurrence of a shift shock at the time of a rapid shift in which an abnormality has occurred in the hydraulic system or the control means.

According to a third aspect of the present invention, one of the speed-increasing oil chamber and the deceleration-side oil chamber defined by the piston is communicated with the other, thereby reducing the differential pressure across the hydraulic cylinder piston. By rapidly reducing the torque transmission capacity of the continuously variable transmission, it is possible to suppress a sudden shift and to suppress the occurrence of a shift shock at the time of a sudden shift in which an abnormality has occurred in the hydraulic system or the control means.

[0015] In the fourth aspect of the present invention, the line pressure is supplied to the speed-increasing oil chamber and the deceleration-side oil chamber defined in the piston, thereby reducing the pressure difference between the front and rear of the hydraulic cylinder. By rapidly reducing the torque transmission capacity of the step transmission, it is possible to suppress a sudden shift and to suppress the occurrence of a shift shock at the time of a sudden shift in which an abnormality has occurred in the hydraulic system or the control means.

According to a fifth aspect of the present invention, the pressure difference between front and rear applied to the piston of the hydraulic cylinder is reduced by connecting the speed-increasing oil chamber and the deceleration-side oil chamber defined in the piston to tanks, respectively. The sudden reduction in the transmission capacity of the transmission
In addition to suppressing a sudden shift, it is possible to suppress the occurrence of a shift shock during a sudden shift in which an abnormality has occurred in the hydraulic system or the control means.

[0017]

DESCRIPTION OF THE PREFERRED EMBODIMENTS One embodiment of the present invention will be described below with reference to the accompanying drawings.

FIGS. 1 to 3 show an example in which the present invention is applied to a double-cavity toroidal type continuously variable transmission 10. FIG. 1 is a schematic configuration diagram of the continuously variable transmission, and FIGS. FIG. 3 shows a circuit diagram of a hydraulic control unit.

In the continuously variable transmission 10, the input shaft 20 is connected to the engine 11 via a torque converter 12 having a lock-up mechanism L / U, while the output shaft 21 is connected to driving wheels (not shown). The transmission mechanism of the toroidal-type continuously variable transmission 10 is configured in the same manner as in the above-described conventional example and the like.
The shift is performed by driving the shift control valve 150 by the shift control valve 150, and the line pressure PL supplied to the shift mechanism is changed by a line pressure solenoid valve 52 that responds to a command from the shift control controller 1.
8.

The shift control controller 1 is mainly composed of a microcomputer, and has a throttle opening TVO detected by a throttle opening sensor 4 and a continuously variable transmission 10.
The operation of the vehicle is performed based on the output shaft rotation speed No from the output shaft rotation sensor 3 detecting the rotation speed of the output shaft 21 and the rotation speed Nt of the input shaft 20 of the continuously variable transmission 10 detected by the input shaft rotation sensor 2. A target gear ratio according to the state is calculated, and a control amount STP (the number of steps) such that the actual gear ratio of the continuously variable transmission 10 matches the target gear ratio is instructed to the step motor 152, and the vehicle speed VSP The line pressure PL of the hydraulic control unit, which will be described later, is determined based on the operating state such as the input shaft rotation speed Nt and the like.
28 is driven by duty control or the like to control the line pressure PL according to the operating state. In this embodiment, a value obtained by multiplying the output shaft rotation speed No by a predetermined constant is used as the vehicle speed VSP.

Next, the hydraulic control unit shown in FIGS. 2 and 3 will be described.

The hydraulic pump 15 driven by the engine 11
Is regulated to a predetermined line pressure PL by a pressure regulator valve 2 (line pressure control means) provided in a line pressure circuit 534, and is transmitted through a transmission control valve 150 and a forward / reverse switching valve 524. Continuously variable transmission 10
Is driven. The line pressure PL of the line pressure circuit 534 is set by the relief valve 512 so as not to exceed a predetermined upper limit.

The pressure regulator valve 2 regulates the line pressure PL of the line pressure circuit 534 in accordance with a signal pressure PLsol from a line pressure solenoid valve 528 whose duty ratio is controlled by the shift control controller 1. Valve 2 has a signal pressure PLso
The line pressure circuit 534 is changed according to the
By draining, the line pressure PL is adjusted.

The signal pressure PLsol corresponding to the duty ratio of the line pressure solenoid valve 528 is input to the accumulator control valve 514 via the oil passage 601, and the accumulation control valve 514 causes the oil passage 54 to respond to the signal pressure PLsol.
By adjusting the oil pressure of 2, the pressure regulator 5
02 is controlled to set the line pressure PL to a predetermined value.

The pressure oil supplied to the torque converter pressure circuit 535 is supplied to the lock-up control valve 508 via a relief valve 512 for controlling the supplied oil pressure so as not to exceed the withstand pressure upper limit value of the torque converter 12.
The lubricating oil passage 38 of the transmission mechanism is formed downstream of the lock-up control valve 508 via a cooler 530. The lock-up control valve 5
08 is controlled based on the signal pressure from the lock-up solenoid valve 526 that is duty-controlled by the shift control controller 1, and engages or releases the lock-up clutch of the torque converter 12.

On the other hand, the line pressure circuit 534 is supplied to the forward shift control valve 150 and the reverse shift control valve 522, respectively, and the forward / reverse switching valve 524 selects the shift control valve 150 or Reverse speed change control valve 52
2 is supplied to the Hi-side oil passage 40 and the Lo-side oil passage 41 of the transmission mechanism.

As shown in FIGS. 1 and 3, the transmission mechanism of the toroidal type continuously variable transmission 10 includes a half toroidal type first toroidal transmission portion 22 and a second toroidal transmission portion 24. An example is shown in which a pair of input / output disks 28, 32 and 29, 33 is provided, and a pair of power rollers 30, 31 sandwiched between an input disk 28 and an output disk 29 of the first toroidal transmission section 22 are Trunnions 83 and 85 driven in opposite directions by a hydraulic servo cylinder provided at the ends and rotatable about an axis.
Supported by

Hydraulic servo cylinders 87 and 89 for driving the trunnions 83 and 85 include a Hi-side oil chamber 516 (increased-side oil chamber) and a Lo-side oil chamber 518 defined on the left and right sides of the piston in the figure.
(Deceleration-side oil chambers), and the trunnions 83 and 85 are displaced in the axial direction according to the pressure difference between these oil chambers, and the tilt angles (tilt angles) of the power rollers 30 and 31 are changed according to the axial displacement. The gear ratio is continuously changed by changing the gear ratio.

For this reason, the hydraulic servo cylinders 87, 89
Are connected to the Hi-side oil chamber 516 communicating with the Hi-side oil passage 40 and Lo.
The arrangement of the Lo side oil chamber 518 communicating with the side oil passage 41 is reversed, and the trunnion 8 is moved in the direction in which the gear ratio becomes Hi.
The Hi-side oil chamber 516 that drives 3, 85
The hydraulic servo cylinder 87 of No. 3 is disposed on the right side of the piston in the figure, whereas the hydraulic servo cylinder 89 of the opposing trunnion 85 is disposed on the left side of the piston in the figure.

At the time of gear shifting, the hydraulic pressure in the Hi-side oil passage 40 and the Lo-side oil passage 41 is relatively changed, so that the pressure difference between the pistons of the hydraulic servo cylinders 87 and 89 (hereinafter referred to as differential pressure), Hi-side oil chamber 516 and Lo-side oil chamber 518
The trunnions 83, 85 are displaced synchronously in opposite axial directions by changing the differential pressure of the trunnions 83, 8
The power rollers 30, 31 tilt (turn around the trunnion axis) in response to the axial displacement of No. 5 so that the gear ratio can be continuously changed. Then, after reaching the predetermined gear ratio, the force applied to the piston in accordance with the pressure difference between the front and rear of the Hi-side oil chamber 516 and the Lo-side oil chamber 518 is applied to the trunnion 8.
3, 85 to support the torque reaction force (traction force) of the power rollers 30, 31 applied to them.

Therefore, the torque transmission capacity of the power rollers 30, 31 is determined based on the differential pressure between the hydraulic servo cylinders 87, 89 supporting the trunnions 83, 83. Note that the second toroidal transmission unit 24 is similarly configured.

In such a toroidal type continuously variable transmission 10, the step ratio is controlled while hydraulic servo is applied by the step motor 152, the shift control valve 150, and the precess cam 136. A sleeve 156 connected to the step motor 152 via pinions 152a and 154c;
Is composed of a spool 158 that can be relatively displaced on the inner circumference of
The spool 158 is connected to the first or second toroidal transmission unit 2.
A precess cam 136 provided on any one of the trunnions 83 and 85 of the second and 24 trunnions;
6 is driven by the feedback link 142 that follows.

For example, in the forward state, the forward / reverse switching valve 524 communicates the Hi-side oil passage 166 of the transmission control valve 152 with the Hi-side oil passage 40 of the transmission mechanism, while the transmission control valve 1
52 communicates with the Lo-side oil passage 168 of the transmission mechanism 52 and the Lo-side oil passage 41 of the transmission mechanism. At this time, when the target gear ratio is on the Hi side,
The step motor 152 drives the sleeve 156 downward in FIG. 3 to transfer the pressure oil of the line pressure circuit 534 to the Hi-side oil passage 16.
6 and 40, while connecting the Lo side oil passage 168 to the tank to drive the trunnion 83 to the left in the figure.
The opposing trunnion 85 is displaced rightward in the figure.

When the target gear ratio coincides with the actual gear ratio, the feedback link 142 oscillated in accordance with the rotation of the precess cam 136 drives the spool 158 downward, and the Hi-side oil passage 166 and the Lo-side oil passage 168 To maintain the target gear ratio.

At this time, the hydraulic servo cylinders 87, 89
The differential pressure between the Hi-side oil chamber 516 and the Lo-side oil chamber 518 determines the capacity of the torque that can be transmitted by the power rollers 30, 31, and 36, 37.

Next, in the normal shift control performed by the shift control controller 1, a target gear ratio is calculated based on a map preset according to the throttle opening TVO (or the amount of depression of an accelerator pedal) and the vehicle speed VSP. And
In addition to the hydraulic servo described above,
Feedback control is performed by the same PI control as in JP-A-772.

On the other hand, as described in the conventional example, in order to suppress the occurrence of abrupt shift when a normal hydraulic pressure cannot be obtained due to damage to the hydraulic piping system or the like and the target gear ratio cannot be maintained, The shift control controller 1 performs control as shown in FIG.

The flowchart shown in FIG. 4 is executed every predetermined time, for example, every 10 msec.

First, at step S1, the input shaft speed N
t and the output shaft speed No. are read, and the actual speed ratio RRTOold during the previous control is read. In step S2, the current actual speed ratio RRTO is calculated from the ratio between the input shaft speed Nt and the output shaft speed No. I do.

Then, in step S3, a value obtained by dividing the difference between the present actual speed ratio RRTO and the previous actual speed ratio RRTOold by the control cycle dt (dt = 10 msec in this case) is calculated as the actual speed speed φ.

In step S4, it is determined whether or not the absolute value of the actual shift speed φ exceeds a predetermined value. If the absolute value exceeds the predetermined value, it is determined that a rapid shift has occurred and step S5 is performed.
Otherwise, the process proceeds to step S7.

In step S5, the line pressure PL of the line pressure circuit 534 is set to a preset minimum value PLmin in order to suppress a sudden shift, and in step S6, the line pressure of the line pressure circuit 534 is set to the minimum value PLmin. Value PLmi
After driving the line pressure solenoid valve 528 so as to reach n, in step S7, the current actual gear ratio RRTO is substituted for the previous value RRTOold, and the process ends.

By performing the above control, normal hydraulic pressure could not be obtained due to damage to the hydraulic piping system and the target gear ratio could not be maintained, and the actual gear ratio started to change rapidly to the Lo or Hi side. Is detected, the line pressure circuit 53
As shown in FIG. 5, the line pressure PL of the normal running is set after the time t1 when the control is started.
To a predetermined minimum value PLmin.

As described above, since the shift of the toroidal type continuously variable transmission is started based on the axial displacement of the trunnions 83 and 85, a sudden shift is started due to an abnormality in the hydraulic system or a malfunction of the control program. In this case as well, the Hi-side oil passage 40 and the Lo-side oil passage 41 of the transmission mechanism are connected to the line pressure circuit 53
4 and the tank communicate with each other to change the speed.

At this time, by reducing the line pressure PL of the line pressure circuit 534 to a predetermined minimum value PLmin,
The maximum value of the hydraulic pressure that can be applied to the pistons of the hydraulic servo cylinders 87 and 89 is limited, and as described above, the torque transmission capacity of the toroidal type continuously variable transmission 10 opposes the traction force of the power rollers 30 and 31. Since it is determined by the differential pressure of the piston, the line pressure PL becomes the minimum value PLmi
By reducing it to n, the torque transmission capacity of the continuously variable transmission 10 also decreases.

For example, a sudden shift to the Lo side (downshift)
In this case, when the torque transmission capacity of the continuously variable transmission 10 is sufficiently large, the magnitude of the shift shock (deceleration shock) depends on the shift speed −φ to the Lo side and the magnitude of the inertia moment of the input system. Decided.

Therefore, the minimum line pressure PLmin at the time of sudden shift
Is set to be the minimum torque transmission capacity that can be traveled, it is possible to reduce the shift shock that occurs, and further, when an abnormality occurs that makes it impossible to shift in the hydraulic system or the shift control controller 1. Even in this case, traveling can be performed by transmitting the minimum necessary torque, and fail-safe of the vehicle including the toroidal-type continuously variable transmission 10 can be ensured.

FIGS. 6 to 10 show a second embodiment, in which the oil passage switching valve 6 is interposed in the Hi-side oil passage 40 and the Lo-side oil passage 41 for driving the transmission mechanism of the first embodiment. An oil passage switching solenoid valve 5 for driving the oil passage switching valve 6 is provided to perform the same control as that shown in FIG. 4. In this case, instead of changing the set value of the line pressure PL ( In steps S5 and S6), when a sudden shift is detected, the oil path switching solenoid valve 5 is driven.

In FIG. 6, the shift control controller 1
As in the flowchart of FIG. 4, when the rapid shift is detected, the oil path switching solenoid valve 5 is turned on to drive the oil path switching valve 6.

As shown in FIGS. 8 and 9, the oil passage switching valve 6 is interposed in the Hi-side oil passage 40 between the forward / reverse switching valve 524 and the hydraulic servo cylinders 87 and 89 on the transmission mechanism side. According to the oil pressure applied to the signal pressure port 6p formed at the other end of the spool 61 biased by the spring 62, the port connected to the output port 6c is connected to the input port 6a.
Or 6b.

The input port 6a of the oil passage switching valve 6 is connected to the upstream side of the Hi-side oil passage 40 (the forward-reverse switching valve 524 side), while the output port 6c is connected to the downstream side of the Hi-side oil passage 40 (the downstream side of the Hi-side oil passage 40). The hydraulic servo cylinders 87 and 89 are connected, and the hydraulic pressure of the Lo-side oil passage 41 is supplied to the input port 6b.

On the other hand, the oil passage switching solenoid valve 5 for driving the oil passage switching valve 6 is interposed in an oil passage 601 'branched from the oil passage 601 of the line pressure solenoid 528, as shown in FIGS. The throttle 7 is provided on the upstream side (on the line pressure solenoid 528 side).

During normal running, the oil path switching solenoid valve 5 is turned off, and a predetermined pilot pressure is applied to the signal pressure port 6p of the oil path switching valve 6, but the spring 62 moves the spool 61 against this pilot pressure. As shown in FIG. 9, the input port 6a and the output port 6c are connected to each other, and the input port 6b is shut off. Therefore, the pressure oil in the Hi-side oil passage 40 is added to the hydraulic servo cylinders 87 and 89 via the oil passage switching valve 6 in the same manner as in the first embodiment,
Normal shift control is performed.

On the other hand, when the shift control controller 1 detects a rapid shift and turns on the oil passage switching solenoid valve 5, the pilot pressure applied to the signal pressure port 6p of the oil passage switching valve 6 increases, so that the spool 61 is resilient. As shown in FIG. 10, the spool 61 is displaced to the right in the figure to communicate the input port 6b and the output port 6c, while shutting off the input port 6a. Therefore, the oil pressure downstream of the Hi-side oil passage 40 becomes equal to the oil pressure Plow of the Lo-side oil passage 41.

That is, at the time of rapid shifting, the Hi-side oil chamber 516 and the Lo-side oil chamber 518 of the hydraulic servo cylinders 87 and 89 are provided.
Or the differential pressure disappears, and the force for supporting the traction force of the power rollers 30 and 31 decreases or becomes zero, so that the torque transmission capacity of the continuously variable transmission 10 sharply decreases. , Hydraulic servo cylinder 8
If the differential pressure at 7, 89 becomes very small or the differential pressure disappears, it is not possible to perform a rapid shift, and a shift shock occurs at the time of a sudden shift when an abnormality occurs in the hydraulic system or the shift control controller 1. Can be suppressed.

Here, when the differential pressure between the hydraulic servo cylinders 87 and 89 becomes small or becomes zero, the power rollers 30 and 31 cannot transmit torque, so that the power rollers 30 and 31 move in the direction of rapid shifting. 31 continues to tilt, but the power rollers 30,
Since a stopper (not shown) for preventing excessive tilting of the power roller 31 is formed, the tilt angle of the power rollers 30, 31, that is, the gear ratio, is the maximum L locked by the stopper.
o The transmission ratio is set to the maximum transmission ratio or the maximum transmission ratio, and the vehicle can run by transmitting the minimum torque by the reaction force locked by the stopper. Thus, it is possible to minimize the occurrence of a shift shock at the time of a sudden shift in which the occurrence of a shift occurs, and at the same time, to enable a minimum traveling, thereby ensuring the fail-safe of the vehicle including the toroidal-type continuously variable transmission 10.

FIGS. 11 and 12 show a third embodiment.
In the second embodiment, the relationship between the Hi-side oil passage 40 and the Lo-side oil passage 41 connected to the oil passage switching valve 6 is reversed, and the other configuration is the same as that of the second embodiment.

That is, the input port 6a of the oil passage switching valve 6
At the upstream side of the Lo side oil passage 41 (forward / backward switching valve 524 side)
Is connected to the output port 6c.
Is connected to the hydraulic servo cylinders 87 and 89, and the hydraulic pressure of the Hi-side oil passage 40 is supplied to the input port 6b.

When the sudden shift is detected and the oil passage switching solenoid valve 5 is turned on, the pilot pressure applied to the signal pressure port 6p of the oil passage switching valve 6 increases, and the spool 61 is attached against the spring 62. As shown in FIG. 12, the spool 61 is displaced to the right in FIG.
And the output port 6c, while blocking the input port 6a. Therefore, the hydraulic pressure downstream of the Lo side oil passage 41 is
It becomes equal to the oil pressure Phi of the Hi-side oil passage 40.

That is, at the time of rapid shifting, the Hi-side oil chamber 516 and the Lo-side oil chamber 518 of the hydraulic servo cylinders 87 and 89 are provided.
Or the differential pressure disappears, and the force for supporting the traction force of the power rollers 30 and 31 decreases or becomes zero, so that the torque transmission capacity of the continuously variable transmission 10 sharply decreases. , Hydraulic servo cylinder 8
If the differential pressure at 7, 89 becomes very small or the differential pressure disappears, it is not possible to perform a rapid shift, and a shift shock occurs at the time of a sudden shift when an abnormality occurs in the hydraulic system or the shift control controller 1. Can be suppressed.

FIGS. 13 and 14 show a fourth embodiment.
The oil passage switching valve 6 of the second embodiment is connected to the Hi-side oil passage 40.
And a line pressure circuit 53 for selectively selecting the downstream side of the Lo side oil passage 41.
4 is replaced with an oil passage switching valve 6 ′, and the other configuration is the same as that of the second embodiment.

The input port 6A of the oil passage switching valve 6 'is Hi
The upstream side of the side oil passage 40 (the forward / reverse switching valve 524 side) is connected to the input port 6C, and the upstream side of the Lo side oil passage 41 (the hydraulic servo cylinders 87, 89 side) is connected to the input port 6C. Is connected to the downstream side of the Hi-side oil passage 40 (the forward-reverse switching valve 524 side), and the output port 6D is connected to the same downstream side of the Lo-side oil passage 41.

The input port 6 of the oil passage switching valve 6 '
Line pressure circuit 534 is connected to E and 6F, and line pressure PL is supplied.

The oil passage switching solenoid valve 5 for driving the oil passage switching valve 6 'is interposed in the oil passage 601' branched from the oil passage 601 of the line pressure solenoid 528, as in the second embodiment. The throttle 7 is provided on the (line pressure solenoid 528 side).

Then, the oil passage switching solenoid valve 5 is turned off.
In this case, a predetermined pilot pressure is applied to the signal pressure port 6p of the oil passage switching valve 6 ', but the spring 62 of the hydraulic switching valve 6' urges the spool 61a to the left in FIG. , Input port 6A and output port 6B
And the input port 6C and the output port 6D communicate with each other. Therefore, the upstream sides of the Hi-side oil passage 40 and the Lo-side oil passage 41 are communicated with the forward / reverse switching valve 524, respectively.
Normal shift control is performed by generating a differential pressure in the oil chambers 516, 518 of the hydraulic servo cylinders 87, 89 according to the oil pressure from the shift control valve 150.

On the other hand, when the shift control controller 1 detects a sudden shift and turns on the oil passage switching solenoid valve 5, the pilot pressure applied to the signal pressure port 6p of the oil passage switching valve 6 'increases, so that the spool 61a As shown in FIG. 14, the spool 61a is displaced to the right in the drawing to shut off the input ports 6A and 6C communicating with the shift control valve 524, while the hydraulic servo cylinder 8 is biased against the spring 62.
The output ports 6B and 6D communicating with the lines 7 and 89 are communicated with the input ports 6E and 6F communicating with the line pressure circuit 534, so that the hydraulic pressure downstream of the Hi-side oil passage 40 and the Lo-side oil passage 41 becomes
It becomes equal to the line pressure PL.

That is, during a rapid shift, the Hi-side oil chamber 516 and the Lo-side oil chamber 518 of the hydraulic servo cylinders 87 and 89 are provided.
Or the differential pressure disappears, and the force for supporting the traction force of the power rollers 30 and 31 decreases or becomes zero, so that the torque transmission capacity of the continuously variable transmission 10 sharply decreases. , Hydraulic servo cylinder 8
If the differential pressure at 7, 89 becomes very small or the differential pressure disappears, it is not possible to perform a rapid shift, and a shift shock occurs at the time of a sudden shift when an abnormality occurs in the hydraulic system or the shift control controller 1. Can be suppressed.

As described above, when the pressure difference between the hydraulic servo cylinders 87 and 89 becomes small or becomes zero, the power rollers 30 and 31 cannot transmit torque, so that the power rollers 30 and 31 move in the direction of rapid shifting. Power rollers 30, 3
Although the tilting of the power roller 3 continues, the stoppers (not shown) for preventing the power rollers 30 and 31 from excessive tilting are formed on the trunnions 83 and 85.
The tilt angles of 0 and 31, ie, the gear ratios, are set to the maximum Lo gear ratio or the maximum Hi gear ratio locked to the stopper, and the minimum torque transmission is performed by the reaction force locked to the stopper. Can make the vehicle run,
In the case of a sudden shift in which an abnormality has occurred in the hydraulic system or the shift control controller 1, while suppressing occurrence of a shift shock, a minimum traveling is enabled, and a fail-safe of the vehicle equipped with the toroidal-type continuously variable transmission 10 is achieved. It can be secured.

FIGS. 15 and 16 show a fifth embodiment.
The input port 6 of the oil passage switching valve 6 'according to the fourth embodiment.
E and 6F are connected to a tank in place of the line pressure circuit 534, and the other configuration is the same as that of the fourth embodiment.

During normal running, the oil passage switching solenoid valve 5 is turned off, and a predetermined pilot pressure is applied to the signal pressure port 6p of the oil passage switching valve 6 '. Spool 61a against pilot pressure
15 to the left side of FIG. 15 to make the input port 6A communicate with the output port 6B and the input port 6C communicate with the output port 6D. Therefore, the Hi-side oil passage 40 and the Lo-side oil passage 4
1 is connected to the forward / reverse switching valve 524 to generate a differential pressure in the oil chambers 516 and 518 of the hydraulic servo cylinders 87 and 89 in accordance with the oil pressure from the transmission control valve 150, so that the normal shift is performed. Perform control.

On the other hand, when the shift control controller 1 detects a sudden shift and turns on the oil passage switching solenoid valve 5, the pilot pressure applied to the signal pressure port 6p of the oil passage switching valve 6 'increases, so that the spool 61a As shown in FIG. 16, the spool 61a is displaced to the right in the figure to shut off the input ports 6A and 6C communicating with the shift control valve 524, while the hydraulic servo cylinder 8 is pressed.
The output ports 6B and 6D communicating with the ports 7 and 89 are connected to the input ports 6E and 6F communicating with the tank, and the hydraulic pressures downstream of the Hi-side oil passage 40 and the Lo-side oil passage 41 are discharged to the tank and are substantially equal. Become.

That is, during a rapid shift, the Hi-side oil chamber 516 and the Lo-side oil chamber 518 of the hydraulic servo cylinders 87 and 89 are provided.
Is discharged to the tank, and the differential pressure is reduced or the differential pressure is eliminated.
Traction force is reduced or becomes zero, so that the torque transmission capacity of the continuously variable transmission 10 is sharply reduced, and the pressure difference between the hydraulic servo cylinders 87 and 89 becomes extremely small or the pressure difference disappears. In this case, it is not possible to perform a rapid shift, and it is possible to suppress the occurrence of a shift shock at the time of a rapid shift in which an abnormality occurs in the hydraulic system or the shift control controller 1.

As described above, when the differential pressure between the hydraulic servo cylinders 87 and 89 becomes small or becomes zero, the power rollers 30 and 31 cannot transmit torque and move in the direction of rapid shifting. Power rollers 30, 3
Although the tilting of the power roller 3 continues, the stoppers (not shown) for preventing the power rollers 30 and 31 from excessive tilting are formed on the trunnions 83 and 85.
The tilt angles of 0 and 31, ie, the gear ratios, are set to the maximum Lo gear ratio or the maximum Hi gear ratio locked to the stopper, and the minimum torque transmission is performed by the reaction force locked to the stopper. Can make the vehicle run,
In the case of a sudden shift in which an abnormality has occurred in the hydraulic system or the shift control controller 1, while suppressing occurrence of a shift shock, a minimum traveling is enabled, and a fail-safe of the vehicle equipped with the toroidal-type continuously variable transmission 10 is achieved. It can be secured.

In the above embodiment, the actual speed change φ is calculated from the input shaft speed Nt and the output shaft speed No. However, the tilting speed of the power rollers 30 and 31 and the displacement speed of the trunnions 83 and 85 are calculated. The actual shift speed may be obtained from the above.

In the above-described embodiment, an example in which the continuously variable transmission 10 has a double cavity is shown. However, although not shown, the same applies to a single-cavity toroidal type.

[Brief description of the drawings]

FIG. 1 shows an embodiment of the present invention and is a schematic configuration diagram of a toroidal-type continuously variable transmission.

FIG. 2 is a first half of a circuit diagram showing a hydraulic control unit of the toroidal-type continuously variable transmission.

FIG. 3 is a second half of a circuit diagram showing a hydraulic control unit of the toroidal-type continuously variable transmission.

FIG. 4 is a flowchart showing an example of control performed by a shift control controller.

FIG. 5 is a graph showing a relationship between a change in line pressure and time.

FIG. 6 shows the second embodiment and is a conceptual configuration diagram of a toroidal-type continuously variable transmission.

FIG. 7 is a first half of a circuit diagram showing a hydraulic control unit of the toroidal-type continuously variable transmission.

FIG. 8 is a second half of a circuit diagram showing a hydraulic control unit of the toroidal-type continuously variable transmission.

FIG. 9 is also a schematic configuration diagram of the oil passage switching valve, showing a state during normal running.

FIG. 10 is also a schematic configuration diagram of an oil passage switching valve, showing a state at the time of rapid shifting.

FIG. 11 is a schematic configuration diagram of an oil passage switching valve according to a third embodiment, showing a state during normal running.

FIG. 12 is also a schematic configuration diagram of an oil passage switching valve, showing a state at the time of rapid shifting.

FIG. 13 is a schematic configuration diagram of an oil passage switching valve according to a fourth embodiment, showing a state during normal running.

FIG. 14 is also a schematic configuration diagram of an oil passage switching valve, showing a state at the time of rapid shifting.

FIG. 15 is a schematic configuration diagram of an oil passage switching valve according to a fifth embodiment, showing a state during normal running.

FIG. 16 is also a schematic configuration diagram of an oil passage switching valve, showing a state at the time of rapid shifting.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 1 Shift control controller 2 Input shaft rotation sensor 3 Output shaft rotation sensor 4 Throttle opening sensor 5 Oil passage switching solenoid valve 6 Oil passage switching valve 7 Throttle 10 Toroidal type continuously variable transmission 11 Engine 12 Torque converter 15 Hydraulic pump 20 Input shaft DESCRIPTION OF SYMBOLS 21 Output shaft 22 1st toroidal transmission part 24 2nd toroidal transmission part 28, 32 Input disk 29, 33 Output disk 30, 31, 36, 37 Power roller 38 Lubricating oil path 40, 166 Hi side oil path 41, 168 Lo side Oil passage 61 Spool 62 Spring 83, 85 Trunnion 87, 89 Hydraulic servo cylinder 136 Precess cam 142 Feedback link 152 Step motor 150 Shift control valve 502 Pressure regulator valve 516 Hi-side oil chamber 518 Lo-side oil chamber 522 Rear Use shift control valve 524 forward and reverse switching valve 528 the line pressure solenoid valve 534 line pressure circuit

Claims (5)

[Claims] 1. A trunnion that is held by an input / output disk to support a tiltable power roller, a hydraulic cylinder that drives the trunnion in an axial direction, and a target gear ratio determined according to a driving state of a vehicle. A transmission control means for controlling a hydraulic pressure to the hydraulic cylinder based on a line pressure, and a line pressure control means for supplying a predetermined line pressure to the transmission control means. The control device includes: an actual shift speed detecting unit that detects an actual shift speed; and a line pressure decreasing unit that decreases a set value of the line pressure control unit when the detected actual shift speed exceeds a predetermined value. A hydraulic control device for a toroidal-type continuously variable transmission. 2. A trunnion, which is held by an input / output disk and supports a tiltable power roller, a hydraulic cylinder that drives the trunnion in an axial direction, and a target gear ratio determined according to a driving state of the vehicle. An actual shift speed detecting means for detecting an actual shift speed; and a detected actual shift speed. A hydraulic pressure control device for a toroidal type continuously variable transmission, comprising: means for reducing or eliminating the differential pressure across the piston constituting the hydraulic cylinder when the pressure exceeds a predetermined value. 3. The hydraulic cylinder includes a speed-increasing oil chamber and a deceleration-side oil chamber defined by a piston, and the front-rear differential pressure reducing means shifts the speed-increasing oil chamber to the deceleration-side oil chamber or decelerates. The hydraulic control device for a toroidal-type continuously variable transmission according to claim 2, wherein the pressure difference between the front and rear is reduced by connecting the side oil chamber to the speed-increasing oil chamber. 4. The hydraulic cylinder includes a speed-increasing oil chamber and a deceleration-side oil chamber defined by a piston, and the front-rear differential pressure reducing means applies a line pressure to the speed-increasing oil chamber and the deceleration-side oil chamber. The hydraulic pressure control device for a toroidal-type continuously variable transmission according to claim 2, wherein the pressure difference is reduced by supplying the pressure. 5. The hydraulic cylinder includes a speed-increasing oil chamber and a deceleration-side oil chamber defined by a piston, and the front-rear differential pressure reducing means includes a speed-increasing oil chamber and a deceleration-side oil chamber in tanks, respectively. The hydraulic control device for a toroidal-type continuously variable transmission according to claim 2, wherein the pressure difference between the front and rear is reduced by connecting.