JP3465552B2 - Hydraulic control device for toroidal type continuously variable transmission - Google Patents

Description

Detailed Description of the Invention

[0001]

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

[0002]

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

In the toroidal type continuously variable transmission, gear change control and torque transmission are performed by the hydraulic pressure applied to the hydraulic servo cylinders that support the trunnion, so that normal hydraulic pressure can be obtained when the hydraulic piping system is damaged. If it is not possible, the shift control will be disabled.

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

[0005]

However, in the above-mentioned 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 the control program malfunctions, and the gear ratio is forced to the maximum. Since it is set to Lo, there is a possibility that a sudden gear shift may occur during normal traveling and a large gear shift shock may occur.

Therefore, the present invention has been made in view of the above problems, and suppresses a shift shock when a sudden shift occurs due to damage of the hydraulic system or malfunction of the control means,
Allows the vehicle to run.

[0007]

[0008]

A first aspect of the present invention is directed to a trunnion which is supported by an input / output disk and which tiltably supports a power roller, a hydraulic cylinder which axially drives the trunnion, and a vehicle operation. In a hydraulic control device for a toroidal type continuously variable transmission, which comprises a shift control means for controlling a hydraulic pressure to the hydraulic cylinder so as to obtain a target speed ratio determined according to a state, an actual shift for detecting an actual shift speed. When the detected actual shift speed exceeds a predetermined value, the speed detection means and a fail-separator different from the oil passage used in the normal shift control.
Front-rear differential pressure reducing means for reducing or eliminating the front-rear differential pressure of the piston that constitutes the hydraulic cylinder by connecting to the oil passage for the harf.

In a second aspect based on the first aspect , the hydraulic cylinder includes an accelerating side oil chamber and a decelerating side oil chamber defined by pistons, and the front-rear differential pressure reducing means is increased. By connecting the speed side oil chamber to the deceleration side oil chamber or the deceleration side oil chamber to the speed increasing side oil chamber, the front-rear differential pressure is reduced.

Further, the third invention, in the first invention, the hydraulic cylinder includes a piston with fraction made the speed increasing side oil chamber and the deceleration side oil chamber, the longitudinal differential pressure lowering means increase The line pressure is supplied to each of the speed side oil chamber and the deceleration side oil chamber to reduce the front-rear differential pressure.

In a fourth aspect based on the first aspect , the hydraulic cylinder includes an accelerating side oil chamber and a decelerating side oil chamber defined by pistons, and the front-rear differential pressure reducing means is increased. By connecting the high speed side oil chamber and the deceleration side oil chamber to the tank respectively, the front-rear differential pressure is reduced.

[0012]

[0013]

[Effect of the Invention] A first invention, during the rapid shifting of the actual shift speed exceeds a predetermined value, or the differential pressure of the hydraulic cylinder is reduced, or the pressure difference will be disappears, counteract the traction force of the power roller Since the hydraulic pressure to be applied decreases or becomes zero, the torque transmission capacity of the continuously variable transmission sharply decreases, and the differential pressure across the hydraulic cylinder becomes extremely small, or the differential pressure disappears, so that rapid gear shifting is not possible. Therefore, it is possible to suppress the occurrence of gear shift shock during a sudden gear shift when an abnormality occurs in the hydraulic system or the control means.

The second aspect of the invention is to reduce the differential pressure across the piston of the hydraulic cylinder by connecting one of the speed increasing side oil chamber and the speed reducing side oil chamber defined by the piston to the other. Since the torque transmission capacity of the continuously variable transmission is sharply reduced, it is possible to suppress the sudden gear shift and also to prevent the gear shift shock from occurring during the sudden gear shift when an abnormality occurs in the hydraulic system or the control means.

The third aspect of the invention is to reduce the differential pressure across the piston of the hydraulic cylinder by supplying the line pressure to the speed increasing side oil chamber and the speed reducing side oil chamber defined by the piston, respectively. Since the torque transmission capacity of the stepped transmission is rapidly reduced, it is possible to suppress the sudden shift and to prevent the occurrence of a shift shock during the sudden shift when an abnormality occurs in the hydraulic system or the control means.

The fourth aspect of the invention is to continuously reduce the differential pressure across the piston of the hydraulic cylinder by connecting the speed-increasing side oil chamber and the decelerating-side oil chamber defined by the piston to the tank, respectively. With the torque transmission capacity of the transmission decreasing rapidly,
It is possible to suppress a sudden shift and also to prevent a shift shock from occurring during a sudden shift when an abnormality has occurred in the hydraulic system or the control means.

[0017]

BEST MODE FOR CARRYING OUT THE INVENTION An embodiment of the present invention will be described below with reference to the accompanying drawings.

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

In the continuously variable transmission 10, the input shaft 20 side is connected to the engine 11 via a torque converter 12 having a lockup mechanism L / U, while the output shaft 21 side is connected to drive 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, etc., and the step motor 1 is operated in response to a command from the shift control controller 1.
A gear shift is performed by driving the gear shift control valve 150 by 52, and the line pressure PL supplied to the gear shift mechanism is a line pressure solenoid valve 52 that responds to a command from the gear shift control controller 1.
Controlled by 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.
Of the output shaft 21 from the output shaft rotation sensor 3 for 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 to drive the vehicle. The target speed ratio corresponding to the state is calculated, and the control amount STP (the number of steps) such that the actual speed ratio of the continuously variable transmission 10 matches the target speed ratio is instructed to the step motor 152, and the vehicle speed VSP is reached. 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 line pressure solenoid valve 5
28 is driven by duty control or the like to control the line pressure PL according to the operating state. In the present embodiment, the product of the output shaft rotation number No and a predetermined constant is used as the vehicle speed VSP.

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

Hydraulic pump 15 driven by the engine 11
Is regulated to a predetermined line pressure PL by the pressure regulator valve 2 (line pressure control means) arranged in the line pressure circuit 534, and the toroidal type via the shift control valve 150 and the forward / reverse switching valve 524. Continuously variable transmission 10
Drive the transmission mechanism. 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 value.

The pressure regulator valve 2 regulates the line pressure PL of the line pressure circuit 534 according to the signal pressure PLsol from the line pressure solenoid valve 528 whose duty ratio is controlled by the shift control controller 1. Valve 2 has a signal pressure PLso
Depending on l, the line pressure circuit 534 is connected to the torque converter pressure circuit 535.
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 accumulation control valve 514 via the oil passage 601 and the accumulation control valve 514 receives the oil passage 54 according to the signal pressure PLsol.
By adjusting the hydraulic 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 lockup control valve 508 via a relief valve 512 which controls the supplied oil pressure so as not to exceed the upper limit of the withstand pressure of the torque converter 12.
Then, a lubricating oil passage 38 of the transmission mechanism is formed downstream of the lockup control valve 508 via a cooler 530. Lock-up control valve 5
08 is controlled based on the signal pressure from the lock-up solenoid valve 526, which is duty-controlled by the shift control controller 1, to engage or release the lock-up clutch of the torque converter 12.

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

Here, as shown in FIGS. 1 and 3, the transmission mechanism of the toroidal type continuously variable transmission 10 is composed of a half toroidal type toroidal transmission section 22 and a second toroidal transmission section 24. An example including a pair of input / output disks 28, 32 and 29, 33 is shown. The pair of power rollers 30, 31 sandwiched between the input disk 28 and the output disk 29 of the first toroidal transmission unit 22 are Trunnions 83, 85 that are driven in opposite directions by hydraulic servo cylinders provided at the ends and are rotatable about their axes
Supported by.

The hydraulic servo cylinders 87 and 89 for driving the trunnions 83 and 85 are Hi side oil chamber 516 (acceleration side oil chamber) and Lo side oil chamber 518 defined on the left and right sides of the piston in the figure.
(Deceleration-side oil chamber), the trunnions 83, 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, 31 are changed in accordance with the axial displacement. By changing it, the gear ratio is continuously changed.

Therefore, the hydraulic servo cylinders 87, 89
Is the Hi-side oil chamber 516 that communicates with the Hi-side oil passage 40 and Lo.
The arrangement of the Lo-side oil chamber 518 that communicates with the side oil passage 41 is reversed, and the trunnion 8 moves in the direction in which the gear ratio is on the Hi side.
The Hi-side oil chamber 516 that drives 3, 85 is
The hydraulic servo cylinder 87 of No. 3 is arranged on the right side of the piston in the drawing, whereas the hydraulic servo cylinder 89 of the trunnion 85 facing the other is arranged on the left side of the piston in the drawing.

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

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. The second toroidal transmission unit 24 has the same structure.

In the toroidal type continuously variable transmission 10 as described above, the step motor 152, the shift control valve 150 and the recess cam 136 are used to control the gear ratio while hydraulic servo is applied. A sleeve 156 connected to the step motor 152 via the pinions 152a and 154c, and a sleeve 156.
It is composed of a spool 158 capable of relative displacement on the inner periphery of
The spool 158 is used for the first or second toroidal transmission unit 2.
The precess cam 136 provided on any one of the 2, 24 trunnions 83, 85, and the precess cam 13
It is driven by a feedback link 142 that follows 6.

For example, in the forward drive state, the forward / reverse switching valve 524 connects the Hi-side oil passage 166 of the shift control valve 152 and the Hi-side oil passage 40 of the transmission mechanism while the shift control valve 1
The Lo-side oil passage 168 of 52 and the Lo-side oil passage 41 of the transmission mechanism communicate with each other. 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 so that the pressure oil of the line pressure circuit 534 is transferred 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 side in the drawing,
The opposing trunnion 85 is displaced to the right side in the figure.

When the target gear ratio matches the actual gear ratio, the feedback link 142 that swings in accordance with the rotation of the precess cam 136 drives the spool 158 downward to drive the Hi side oil passage 166 and the Lo side oil passage 168. The target gear ratio is maintained by sealing the.

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 torque capacity that can be transmitted by the power rollers 30, 31, 36 and 37.

Next, in the normal shift control performed by the shift control controller 1, a target shift ratio is calculated based on a map preset according to the throttle opening TVO (or the accelerator pedal depression amount) and the vehicle speed VSP. And
In addition to the hydraulic servo as described above, JP-A-8-270
Feedback control is performed by PI control similar to that of Japanese Patent No. 772.

On the other hand, as described in the above-mentioned conventional example, in order to suppress the occurrence of a sudden gear shift when normal hydraulic pressure cannot be obtained due to damage of the hydraulic piping system 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, in step S1, the input shaft speed N
t and the output shaft speed No., the actual speed ratio RRTOold at the time of the previous control is read, and 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. To do.

Then, in step S3, a value obtained by dividing the difference between the current 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 change φ.

In step S4, it is determined whether or not the absolute value of the actual shift speed φ exceeds a predetermined value. If it exceeds the predetermined value, it is determined that a sudden shift has occurred, and step S5 is performed.
On the other hand, if not, 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 the sudden shift, and in step S6, the line pressure PL of the line pressure circuit 534 is set to the minimum value PLmin. Value PLmi
After the line pressure solenoid valve 528 is driven so as to be n, in step S7, the current actual speed ratio RRTO is substituted into the previous value RRTOold, and the process ends.

By performing the above control, normal hydraulic pressure cannot be obtained due to damage of the hydraulic piping system, etc., and the target gear ratio cannot be maintained, and the actual gear ratio rapidly changes to Lo side or Hi side. Is detected, the line pressure circuit 53
As shown in FIG. 5, the line pressure PL of No. 4 is the line pressure set value PL0 during normal traveling after the time t1 when the control is started.
To a predetermined minimum value PLmin.

The shifting of the toroidal type continuously variable transmission is started based on the axial displacement of the trunnions 83 and 85 as described above, so that the rapid shifting is started due to an abnormality in the hydraulic system or a malfunction of the control program. Also in the case of the normal shift, 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 via the transmission control valve 150.
4 and the tank communicate with each other to change gears.

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. The line pressure PL is the minimum value PLmi because it is determined by the differential pressure of the piston.
By reducing the value to n, the torque transmission capacity of the continuously variable transmission 10 also decreases.

For example, a sudden shift (downshift) to the Lo side
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 to the Lo side -φ 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 so that the minimum torque transmission capacity that can be traveled is set, it is possible to reduce the shift shock that occurs, and when an abnormality that prevents shifting is generated in the hydraulic system or the shift control controller 1. Even in this case, the vehicle can be traveled by transmitting the minimum necessary torque, and the fail safe of the vehicle provided with the toroidal type continuously variable transmission 10 can be ensured.

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

In FIG. 6, the shift control controller 1
In the same manner as in the flow chart of FIG. 4, when the sudden shift is detected, the oil passage switching solenoid valve 5 is turned on and the oil passage switching valve 6 is driven.

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 side of the speed change mechanism. The port connected to the output port 6c is connected to the input port 6a according to the hydraulic pressure applied to the signal pressure port 6p formed at the other end of the spool 61 urged by the spring 62.
Alternatively, it is switched to 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 (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 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, as shown in FIGS. 7 and 8, 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, The throttle 7 is provided on the upstream side (line pressure solenoid 528 side).

During normal traveling, 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, but the spring 62 resists this pilot pressure and causes the spool 61 to move. Energize to connect the input port 6a and the output port 6c while blocking the input port 6b, as shown in FIG. Therefore, the pressure oil in the Hi-side oil passage 40 is added to the hydraulic servo cylinders 87, 89 via the oil passage switching valve 6 as in the first embodiment,
Normal shift control is performed.

On the other hand, when the shift control controller 1 detects a sudden shift 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, so that the spool 61 springs. As shown in FIG. 10, the spool 61 is biased against 62 and is displaced to the right side in the drawing to connect the input port 6b and the output port 6c, while blocking 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 a sudden shift, the Hi side oil chamber 516 and the Lo side oil chamber 518 of the hydraulic servo cylinders 87 and 89.
Of the power rollers 30 and 31 decreases or becomes zero, and the torque transmission capacity of the continuously variable transmission 10 decreases sharply. , Hydraulic servo cylinder 8
Since the differential pressure between 7 and 89 becomes extremely small or the differential pressure disappears, rapid gear shifting cannot be performed, and a gear shift shock occurs during a sudden gear shift when an abnormality occurs in the hydraulic system or the gear shift control controller 1. Can be suppressed.

Here, 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, and the power rollers 30 and 31 move in the direction of sudden gear change. Although the tilting of 31 continues, the power roller 30,
Since a stopper (not shown) for preventing excessive tilting of 31 is formed, the tilting angle of the power rollers 30, 31, that is, the gear ratio, is the maximum L that is locked by this stopper.
o The gear ratio is set to the highest gear ratio or the highest gear ratio, and the vehicle can be allowed to travel by the minimum torque transmission by the reaction force locked by the stopper, and the hydraulic system and the gear shift controller 1 have an abnormality. It is possible to suppress the occurrence of a shift shock at the time of a sudden gear shift that occurs, to enable a minimum amount of travel, and to ensure fail-safe of a vehicle including the toroidal type continuously variable transmission 10.

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 other configurations are the same as those in the second embodiment.

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

When a 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 toward the right side in the figure and the input port 6b is moved.
And the output port 6c are communicated with each other, while the input port 6a is blocked. Therefore, the oil pressure downstream of the Lo-side oil passage 41 is
It becomes equal to the hydraulic pressure Phi of the Hi-side oil passage 40.

That is, at the time of a sudden shift, the Hi side oil chamber 516 and the Lo side oil chamber 518 of the hydraulic servo cylinders 87 and 89.
Of the power rollers 30 and 31 decreases or becomes zero, and the torque transmission capacity of the continuously variable transmission 10 decreases sharply. , Hydraulic servo cylinder 8
Since the differential pressure between 7 and 89 becomes extremely small or the differential pressure disappears, rapid gear shifting cannot be performed, and a gear shift shock occurs during a sudden gear shift when an abnormality occurs in the hydraulic system or the gear shift control controller 1. Can be suppressed.

13 and 14 show a fourth embodiment,
The oil passage switching valve 6 of the second embodiment is replaced with the Hi side oil passage 40.
And a line pressure circuit 53 selectively on the downstream side of the Lo side oil passage 41.
It is replaced with an oil passage switching valve 6'connected to No. 4 and other configurations are the same as those in the second embodiment.

Hi is connected to the input port 6A of the oil passage switching valve 6 '.
The upstream side of the side oil passage 40 (forward / reverse switching valve 524 side) is connected to the same upstream side of the Lo side oil passage 41 (hydraulic servo cylinders 87, 89 side), respectively, while the output port 6B 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 '
The line pressure circuit 534 is connected to E and 6F, and the line pressure PL is supplied.

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

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

On the other hand, when the shift control controller 1 detects a sudden shift 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. 14, the spool 61a is displaced to the right in the figure to block the input ports 6A and 6C communicating with the transmission control valve 524, while the hydraulic servo cylinder 8 is being urged against the spring 62.
By connecting the output ports 6B and 6D that communicate with Nos. 7 and 89 to the input ports 6E and 6F that communicate with the line pressure circuit 534, the hydraulic pressures downstream of the Hi-side oil passage 40 and the Lo-side oil passage 41 are
It becomes equal to the line pressure PL.

That is, at the time of a sudden shift, the Hi side oil chamber 516 and the Lo side oil chamber 518 of the hydraulic servo cylinders 87 and 89.
Of the power rollers 30 and 31 decreases or becomes zero, and the torque transmission capacity of the continuously variable transmission 10 decreases sharply. , Hydraulic servo cylinder 8
Since the differential pressure between 7 and 89 becomes extremely small or the differential pressure disappears, rapid gear shifting cannot be performed, and a gear shift shock occurs during a sudden gear shift when an abnormality occurs in the hydraulic system or the gear shift control controller 1. Can be suppressed.

Then, similarly to the 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 the power rollers 30 and 31 move in the direction of rapid gear shift. Power rollers 30, 3
Although the tilting of No. 1 continues, the trunnions 83 and 85 have stoppers (not shown) for preventing excessive tilting of the power rollers 30 and 31.
The tilt angles of 0 and 31, that is, the gear ratio is set to the highest Lo gear ratio or the highest Hi gear ratio locked by this stopper, and minimum torque transmission is performed by the reaction force locked by the stopper. Can enable the running of the vehicle,
A failsafe for a vehicle equipped with the toroidal-type continuously variable transmission 10 is realized by suppressing 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 shift control controller 1 and enabling a minimum running. It can be secured.

FIG. 15 and FIG. 16 show the fifth embodiment,
Input port 6 of oil passage switching valve 6'of the fourth embodiment
E and 6F are connected to a tank instead of the line pressure circuit 534, and other configurations are the same as those in the fourth embodiment.

During normal traveling, 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 ', but the spring 62 of the hydraulic pressure switching valve 6'is Spool 61a against pilot pressure
Is urged to the left side of FIG. 15 to connect the input port 6A and the output port 6B and the input port 6C and the output port 6D, respectively. Therefore, the Hi-side oil passage 40 and the Lo-side oil passage 4
The upstream side of No. 1 is communicated with the forward / reverse switching valve 524, respectively, and generates a differential pressure in the oil chambers 516 and 518 of the hydraulic servo cylinders 87 and 89 in accordance with the hydraulic pressure from the shift control valve 150, thereby performing normal shifting. Take control.

On the other hand, when the shift control controller 1 detects a sudden shift 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. As shown in FIG. 16, the spool 61a is biased against the spring 62 and is displaced to the right side in the figure to block the input ports 6A and 6C communicating with the transmission control valve 524, while the hydraulic servo cylinder 8
The output ports 6B and 6D communicating with 7, 89 are communicated with 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, at the time of a sudden shift, the Hi side oil chamber 516 and the Lo side oil chamber 518 of the hydraulic servo cylinders 87 and 89.
Is discharged to the tank to reduce the differential pressure or the differential pressure disappears.
Since the force supporting the traction force of the hydraulic pressure decreases or becomes zero, the torque transmission capacity of the continuously variable transmission 10 decreases sharply, and the differential pressure between the hydraulic servo cylinders 87 and 89 becomes extremely small or the differential pressure disappears. Therefore, it is possible to suppress the occurrence of a gear shift shock during a sudden gear shift when an abnormality has occurred in the hydraulic system or the gear shift control controller 1.

Then, similarly to the 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 the power rollers 30 and 31 move in the direction of rapid gear shifting. Power rollers 30, 3
Although the tilting of No. 1 continues, the trunnions 83 and 85 have stoppers (not shown) for preventing excessive tilting of the power rollers 30 and 31.
The tilt angles of 0 and 31, that is, the gear ratio is set to the highest Lo gear ratio or the highest Hi gear ratio locked by this stopper, and minimum torque transmission is performed by the reaction force locked by the stopper. Can enable the running of the vehicle,
A failsafe for a vehicle equipped with the toroidal-type continuously variable transmission 10 is realized by suppressing 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 shift control controller 1 and enabling a minimum running. It can be secured.

In the above embodiment, the actual shift speed φ is calculated from the input shaft rotation speed Nt and the output shaft rotation 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 speed change speed may be obtained from the above.

Further, in the above embodiment, an example in which the continuously variable transmission 10 is constituted by a double cavity is shown. However, although not shown, it may be constituted by a single cavity toroidal type.

[Brief description of drawings]

FIG. 1 is a schematic configuration diagram of a toroidal type continuously variable transmission according to an embodiment of the present invention.

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

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

FIG. 4 is a flow chart showing an example of control performed by the shift control controller.

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

FIG. 6 is a conceptual configuration diagram of a toroidal type continuously variable transmission according to the second embodiment.

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

FIG. 8 is likewise the latter half of the circuit diagram showing the 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 traveling.

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

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

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

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

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

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

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

[Explanation 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 aperture 10 Toroidal type continuously variable transmission 11 engine 12 Torque converter 15 Hydraulic pump 20 input axis 21 Output shaft 22 1st toroidal transmission 24 2nd toroidal transmission 28, 32 input disc 29, 33 output disc 30, 31, 36, 37 Power roller 38 Lubricating oil passage 40, 166 Hi side oil passage 41, 168 Lo side oil passage 61 spool 62 spring 83,85 trunnion 87, 89 hydraulic servo cylinder 136 Precessum 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 Reverse shift control valve 524 Forward-reverse switching valve 528 line pressure solenoid valve 534 line pressure circuit

Continuation of front page (58) Fields surveyed (Int.Cl. ⁷ , DB name) F16H 59/00-61/12 F16H 61/16-61/24 F16H 63/40-63/48

Claims (4)

(57) [Claims] 1. A trunnion sandwiched between an input / output disk and tiltably supporting a power roller, a hydraulic cylinder driving the trunnion in an axial direction, and a target gear ratio determined according to a driving state of a vehicle. Shift control means for controlling the hydraulic pressure to the hydraulic cylinder so that
Toroidal type continuously variable transmission hydraulic control device equipped with
The actual shift speed detecting means for detecting the actual shift speed and the normal shift speed when the detected actual shift speed exceeds a predetermined value.
Oil passage for fail-safe that is different from the oil passage used for speed control
Piston that constitutes the hydraulic cylinder by connecting to
A hydraulic control device for a toroidal-type continuously variable transmission, comprising: 2. The hydraulic cylinder is defined by a piston.
Equipped with a speed-increasing side oil chamber and a decelerating-side oil chamber.
For the step, increase the oil chamber on the acceleration side to the oil chamber on the deceleration side or increase the oil chamber on the deceleration side.
Lowering the differential pressure across the cylinder by communicating with the high-speed oil chamber
The hydraulic control device for a toroidal type continuously variable transmission according to claim 1 . 3. The hydraulic cylinder includes a speed-increasing side oil chamber and a decelerating-side oil chamber defined by pistons, and the front-rear differential pressure reducing means applies a line pressure to the speed-increasing side oil chamber and the decelerating side oil chamber. For each
A contract characterized by lowering the differential pressure across the supply
A hydraulic control device for a toroidal type continuously variable transmission according to claim 1 . 4. The hydraulic cylinder includes a speed-increasing side oil chamber and a decelerating-side oil chamber defined by pistons, and the front-rear differential pressure reducing means includes a speed-increasing side oil chamber and a decelerating-side oil chamber in tanks, respectively. Connection
Claim that is characterized by reducing the differential pressure across the
Item 2. A hydraulic control device for a toroidal-type continuously variable transmission according to Item 1 .