JP3505767B2 - 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 a hydraulic control device for a toroidal type continuously variable transmission.

2. Description of the Related Art A toroidal type continuously variable transmission is provided with an input disk, an output disk and a power roller interposed between the two disks, and the tilt angle of the power roller is changed. Thus, the rotation ratio of the output disc with respect to the input disc, that is, the gear ratio is changed. In order to change the tilt angle of the power roller, hydraulic gear ratio adjusting means is provided. Then, it has been proposed that the hydraulic pressure supply / discharge control for the speed ratio adjusting means be performed by a speed ratio control valve.

By the way, when the amount of oil relieved from the gear ratio control valve increases, that is, when the amount of oil required for gear shifting of the toroidal type continuously variable transmission increases, the driving loss of the pump driven by the engine increases. .

Japanese Laid-Open Patent Publication No. 62-171557 discloses controlling the line pressure so that the line pressure corresponds to the input torque of the toroidal type continuously variable transmission. However, in this case, there is no problem when considering only the shift control of the toroidal type continuously variable transmission, but the pump normally supplies the hydraulic pressure to the hydraulic clutch interposed in the power transmission path. In consideration of the relationship with this clutch, the technique described in the above publication cannot be practically adopted. That is, when the line pressure required in the toroidal type continuously variable transmission is small or unnecessary, the line pressure required in the clutch becomes large,
Reducing the line pressure in accordance with the toroidal type continuously variable transmission becomes undesirable in terms of clutch control.

Further, in Japanese Patent Laid-Open No. 1-135958, oil leaked from the relief port of the gear ratio control valve is
What is utilized for lubrication of a toroidal type continuously variable transmission is disclosed. However, when the relief amount is small,
While it is a time when a large amount of oil is required for gear ratio control, this is also a time when a large amount of oil is required for lubrication, and both the requirements for both gear ratio control and lubrication must be satisfied. Will be difficult.

The present invention has been made in consideration of the above circumstances, and in a toroidal type continuously variable transmission, the amount of oil required for the gear ratio control itself is reduced as much as possible to reduce the size of the pump. An object of the present invention is to provide a hydraulic control device for a toroidal type continuously variable transmission that can reduce driving power.

In order to achieve the above object, the present invention provides a first invention corresponding to claim 1, a second invention corresponding to claim 2 and a third invention corresponding to claim 3 in the claims.
And a fourth invention corresponding to claim 4. The first to fourth inventions have the following basic configuration as a common configuration. That is, in a toroidal type continuously variable transmission that is equipped with a hydraulic gear ratio adjusting means, and hydraulic pressure is supplied to and discharged from the gear ratio adjusting means by a gear ratio control valve, a relief port of the gear ratio control valve is provided. A relief valve for adjusting the relief pressure so as to reduce the amount of oil that is connected and is relieved from the gear ratio control valve is provided, and the gear ratio adjusting means is set to perform a gear shift by a hydraulic pressure difference. The gear ratio control valve is configured so as to supply a high hydraulic pressure and a low hydraulic pressure for giving the hydraulic pressure difference to the gear ratio adjusting means.

Each of the first to fourth inventions is further provided with the following configuration on the assumption that each has the above-mentioned basic configuration. That is, in the first invention corresponding to claim 1 in the claims, the relief valve is set to increase the relief pressure as the low hydraulic pressure increases.
It is further equipped with such a configuration.

In the second invention described in claim 2 of the claims, the relief valve increases the relief pressure as the differential pressure between the line pressure and the high hydraulic pressure increases. It is further equipped with the set configuration.

In the third invention described in claim 3 in the claims, the relief valve increases the relief pressure as the differential pressure between the high hydraulic pressure and the low hydraulic pressure decreases. It is further equipped with the set configuration.

According to a fourth aspect of the present invention set forth in claim 4, when the relief valve sets a differential pressure between the high hydraulic pressure and the low hydraulic pressure to a predetermined differential pressure, the predetermined pressure is set. The relief pressure is set to increase as the pressure difference between the pressure difference and the line pressure increases.

[0011]

According to the present invention, the amount of oil that is relieved from the gear ratio control valve is controlled to be small, so that the amount of oil required for gear ratio control is reduced and the pump is downsized. be able to. In particular, according to the first aspect of the present invention, the relief pressure can be set to a value close to the low hydraulic pressure, and the amount of oil required for shifting can be reduced.

According to the second aspect of the present invention, it is possible to reduce the amount of oil required for gear shifting by setting the high hydraulic pressure for obtaining the hydraulic pressure difference to a value close to the line pressure. According to the third aspect of the present invention, the fact that the hydraulic pressure difference required for shifting is small means that the hydraulic pressure difference can be sufficiently secured even if the relief pressure is large.
The amount of oil required for shifting can be reduced. According to the fourth aspect of the present invention, the relief pressure is increased by an amount that can secure the hydraulic pressure difference required for the shift with respect to the line pressure as the most basic hydraulic pressure, and the amount of oil required for the shift is increased. Can be reduced.

[0013]

EXAMPLES Examples of the present invention will be specifically described below. Overall Overview As shown in FIG.
The output torque of the engine 1 including the fourth cylinders # 1 to # 4 is transmitted via a torque converter 2 (torque converter 2), a gear reduction mechanism 3 including a planetary gear system, and a forward / reverse switching mechanism 4 to a transmission output shaft. A first transmission unit for outputting to 5 is provided. Further, the output torque of the engine 1 is output to the transmission output shaft 5 via the switching clutch 6 and the toroidal type continuously variable transmission C provided with the first toroidal type transmission mechanism 7 and the second toroidal type transmission mechanism 8. A second speed changing unit is provided. When the switching clutch 6 is in the off state, torque is transmitted via the first speed change portion, and when it is in the on state, the torque is transmitted via the second speed change portion.

Since the torque converter 2 has a strong torque increasing function, the first speed change portion is mainly used when a relatively large speed change ratio (torque ratio) is required at the time of starting, accelerating, etc. The speed changer is mainly used when a vehicle is driven at a relatively small speed change ratio, such as during high-speed traveling.

The specific structure of the transmission TM will be described below. First, the first transmission unit will be described. The torque converter 2 is substantially composed of a pump impeller 12, a turbine liner 13, and a stator 14. The pump impeller 12 is connected to the engine output shaft 1 via the pump cover 15.
1 and is configured to rotate integrally with the engine output shaft 11. Further, the pump impeller 12 has a first
The hollow shaft 16 is coaxially connected, and the oil pump 1 is attached to the rear end portion (right end portion in FIG. 1) of the first hollow shaft 16.
7 are connected. The turbine liner 13 is coaxially connected to the torque converter output shaft 18 (turbine shaft). The stator 14 is connected to the second hollow shaft 21 via a one-way clutch 19. The second hollow shaft 21 is fixed to the transmission case 22.

The gear reduction mechanism 3 substantially includes the sun gear 2.
5, a first pinion 26, a reverse ring gear 27, a second pinion 28, a forward ring gear 29, and a carrier 30. The sun gear 25 has a torque converter output shaft 18
The torque of is input. Also, the first
The second pinions 26 and 28 are rotatably supported by a carrier 30. The carrier 30 is fixed to the second hollow shaft 21 fixed to the transmission case 22.

Then, the sun gear 25 and the first pinion 26
The front part of the gear meshes with the first pinion 26 and the reverse ring gear 27 meshes with each other to form a planetary gear system having a reverse reduction function. In this planetary gear system, a torque in the reverse rotation direction larger than the torque input to the sun gear 25 is output from the reverse ring gear 27.

The first gear meshing with the sun gear 25
The rear part of the pinion 26 and the second pinion 28 mesh with each other,
Further, the second pinion 28 and the forward ring gear 29 mesh with each other, and these form a planetary gear system having a forward rotation deceleration function. In this planetary gear system, the forward rotation direction torque that is larger than the torque input to the sun gear 25 is output from the forward ring gear 29.

The forward / reverse switching mechanism 4 is substantially composed of a reverse clutch 31, a clutch case 32, a forward clutch 33 and a one-way clutch 34.
The clutch case 32 is connected to the transmission output shaft 5 via an output disc 44f of the first toroidal transmission mechanism 7 described later. Here, when the reverse clutch 31 is engaged (the forward clutch 3
3 is disconnected), reverse ring gear 27 and clutch case 3
2 is connected, and the torque of the reverse ring gear 27 is transmitted to the transmission output shaft 5. On the other hand, the forward clutch 3
3 is engaged (the reverse clutch 31 is disengaged), the torque of the forward ring gear 29 is applied to the transmission output shaft 5.
Be transmitted to. The one-way clutch 34 runs idle when the rotation speed of the transmission output shaft 5 is higher than the rotation speed of the forward drive ring gear 29 to prevent the forward drive ring gear 29 from being reversely driven by the transmission output shaft 5. It is provided in.

Next, the second transmission section will be described. The second speed change unit is substantially a toroidal type transmission including a changeover switch 6 for connecting and disconnecting a torque input to the second speed change unit, and first and second toroidal type speed change mechanisms 7 and 8. It is composed of a stepped transmission C and a gear mechanism 35 that transmits torque from the switching clutch 6 to the toroidal continuously variable transmission C. Here, the gear mechanism 35 sequentially engages the torque of the first hollow shaft 16, that is, the torque of the engine output shaft 11, when the switching clutch 6 is engaged, with the drive gear 37, the idle gear 38, and the driven gear 39. The torque of the bypass shaft 40 is further transmitted to the toroidal type continuously variable transmission C via the driving gear 41 and the driven gear 42 which are meshed with each other.

The toroidal type continuously variable transmission C includes a first toroidal type transmission mechanism 7 arranged on the front side (left side in FIG. 1) so as to surround the transmission output shaft 5 and a second toroidal type transmission mechanism arranged on the rear side. It is composed of a toroidal type speed change mechanism 8. Here, the first and second toroidal type speed change mechanisms 7 and 8 are arranged symmetrically in the front and rear, but since the configurations and functions of both are basically the same, corresponding members are not included. In principle, each member of the first toroidal type speed change mechanism 7 is attached with a subscript f, and the second toroidal type speed change mechanism 8 is attached.
Each member is marked with a subscript r. However, such as rollers
The members arranged in twos in each toroidal transmission mechanism 7 and 8 cannot be distinguished only by the subscripts f and r. Therefore, the subscripts a and a are respectively added to the respective members disposed to the left and right of the first toroidal transmission mechanism 7. b, and the subscripts c and d are attached to the respective members arranged on the left and right of the second toroidal transmission mechanism 8. Therefore, in the following, in principle, the description made for one member also applies to other members having the same number but different subscripts.

The first toroidal transmission mechanism 7 has a first input disk 43f loosely fitted around the transmission output shaft 5.
And the first output disc 44 fixed to the transmission output shaft 5.
f and the first and second rollers 45a, 45b for transmitting the torque of the first input disk 43f to the first output disk 44f.
And are provided. Then, the first input disk 43f
Is the input cam 4 to which the driven gear 42 is attached.
8 through the first cam roller 49f, and the larger the input torque to the first input disk 43f, the stronger the input cam 48 is pressed against the first input disk 43f.

The first and second rollers 45a and 45b are rotatable about the axis Y, and their peripheral surfaces are curved surfaces of the first input disk 43f and the first output disk 44f. It is in contact with the surface. Therefore, the first
The torque (rotation) of the input disk 43f is transmitted to the first output disk 44f via the first and second rollers 45a and 45b. Here, the speed change ratio (torque ratio) in torque transmission from the first input disk 43f to the first output disk 44f is the first and second rollers 45.
It is determined by the ratio R2 / R1 of the radius R2 of the first output disc 44f and the radius R1 of the first input disc 43f at the position in contact with a and 45b.

Then, the first and second rollers 45a, 45b
As will be described later, the abutting positions of the two discs 43f and 44f are determined by the tilt angles of the first and second rollers 45a and 45b, which will be described later. By changing the tilt angle, the gear ratio can be arbitrarily set within a predetermined range. Of course, the second toroidal transmission mechanism 8 is basically the same as the first toroidal transmission mechanism 7.

Details of Toroidal Type Continuously Variable Transmission The specific structure of the toroidal type continuously variable transmission C will be described below. As shown in FIGS. 2 to 5, in the first toroidal transmission mechanism 7, the first output disk 44f is spline-fitted to the transmission output shaft 5. Further, the first output disk 44f is rotatably supported by the transmission case 22 via the first bearing 47f while being positioned by the ring-shaped positioning member 46 fitted to the transmission output shaft 5. Has been done. The second output disk 44r of the second toroidal transmission mechanism 8 is provided between the enlarged diameter portion 5g formed integrally with the transmission output shaft 5 and the transmission case 22, and the transmission output is provided. It is positioned by a second bearing 47r that rotatably supports the shaft 5.

Between the first output disk 44f and the second output disk 44r, the first and second input disks 43f,
43r are arranged close to each other with their rear surfaces facing each other, and between both input disks 43f and 43r, an input cam 48 which is rotatable relative to them is arranged. Then, the input cam 48 and the first and second
Between the input discs 43f and 43r, the first and second discs are respectively formed.
Cam rollers 49f and 49r are interposed. here,
The first and second cam rollers 49f and 49r are provided with the first and second input disks 43f and 43r when the input cam 48 and the first and second disks 43f and 43r rotate relative to each other.
Has a function of generating a pressing force for pressing the output discs 44f and 44r side, and the larger the input torque to the first and second input discs 43f and 43r, the first and second cam rollers 49f. , 49r increases the pressing force on the first and second input disks 43f, 43r.

Between the first and second input disks 43f and 43r, the output shaft 5 of the transmission is loosely fitted, and both ends are in contact with the back surfaces of the first and second input disks 43f and 43r, respectively. An engaging member 50 that is spline-fitted to the first and second input disks 43f and 43r is arranged. A disc spring 51 is provided between the engagement member 50 and the second input disc 43r, and the disc spring 51 preloads the first input disc 43f and the second input disc 43r in a direction in which they are separated from each other. It has become so. The disc spring 51 abuts against the back surface of the second input disk 43r and urges the second input disk 43r toward the second output disk 44r, and the urging reaction force of the disk spring 51 causes the first input disk via the engaging member 50. 43f is urged toward the first output disc 44f,
Between the first input disk 43f and the first output disk 44f, and between the second input disk 43r and the second output disk 4
A predetermined preload is applied between 4r and 4r.

Next, a hydraulic mechanism for tilting the first to fourth rollers 45a to 45d will be described. First
The toroidal type speed change mechanism 7 includes first and second rollers 45.
First and second trunnions 59a and 59b that rotatably support a and 45b, respectively, are provided. And the first,
By the second trunnions 59a and 59b, respectively,
The first and second rollers 45a and 45b are rotatably supported via the second eccentric shafts 60a and 60b. Further, the first and second trunnions 59a and 59b are respectively integrated with first and second shaft members 61a and 61b which extend downward (in a direction orthogonal to the transmission output shaft 5). Installed in place.

The upper connecting member 6 is attached to the transmission case 22 slightly above the first and second rollers 45a and 45b.
2 is attached. On the other hand, the first and second rollers 45
A lower connecting member 63 is attached to a partition wall portion 53 fixed to the transmission case 22 below a and 45b. Then, the first and the first formed on the upper connecting member 62
The upper ends of the first and second trunnions 59a and 59b are rotatably supported by the second shaft holes 65a and 65b via the first and second upper spherical bushes 64a and 64b, respectively. On the other hand, due to the first and second shaft holes 67a and 67b formed in the lower connecting member 63, the lower end portions of the first and second trunnions 59a and 59b respectively correspond to the first and second lower spherical bushes 66a and 66a. It is rotatably supported via 66b.

The lower portions of the first and second shaft members 61a and 61b are attached to the lower surface of the upper housing 55 by penetrating the opening 55g of the upper housing 55 attached to the lower surface of the partition wall 53. Lower housing 56
The concave portion 56g of the first and second support bearings 54
It is rotatably supported via a and 54b.

In the partition wall portion 53, there are provided first and second trunnions 59a and 59b for operating respectively.
Hydraulic cylinders 76a and 76b are provided, and these first and second hydraulic cylinders 76a and 76b are vertically partitioned by a partition wall portion 53g which is a part of the partition wall portion 53. Then, the first and second hydraulic cylinders 76a, 7
The upper half of 6b has first and second upper pistons 77a,
77b is fitted, and the first and second lower pistons 78a, 78b are fitted in the lower half part. Therefore, the first and second upper pistons 77a and 77b and the partition 53g define the first and second upper hydraulic chambers 79a and 79b, respectively, while the first and second lower pistons 78a and 78b and the partition are separated from each other. The portion 53g and the first and second lower hydraulic chambers 80a,
80b is defined.

Here, the first and second upper hydraulic chambers 79a, 7a
When hydraulic pressure is applied to 9b, the first and second upper pistons 77a and 77b cause the first and second trunnion 5 to move.
9a and 59b are displaced upward, while the first and second
When hydraulic pressure is applied to the lower hydraulic chambers 80a and 80b, the first and second lower pistons 78a. By 78b,
The first and second trunnions 59a and 59b are adapted to be displaced downward. And, like this, the first and second
When the trunnions 59a and 59b are displaced in the vertical direction, the trunnions 59a and 59b are accordingly displaced according to the displacement amount.
5b is tilted, and the gear ratio of the first toroidal transmission mechanism 7 is changed. Along with this, the first and second trunnions 59a and 59b are adapted to rotate about their axes. The first to fourth upper hydraulic chambers 79
a to 79d and the first to fourth lower hydraulic chambers 80a to 80d, the gear ratio control device for controlling the gear ratio by controlling the supply of hydraulic pressure, and the specific gear shift operation by the gear ratio control device are: More on this later.

In order to back up the synchronization of the rotations of the first and second trunnions 59a and 59b when the hydraulic mechanism fails, both trunnions 59a and 59b (59c, 59c,
The interlocking wires 57 and 58 are wound around 59d).

The upper connecting member 62 has a first upper positioning hole 68f formed in the middle of the first shaft hole 65a and the second shaft hole 65b, and a middle of the third shaft hole 65c and the fourth shaft hole 65d. A second upper positioning hole 68r is formed. And
A first support portion 69f integrally formed with the transmission case 22 is inserted through the first upper positioning hole 68f. In addition,
The first roller lubrication member 70f is attached to the first support portion 69f by using the first attachment member 71f. Also,
The upper spherical bearing 75r attached to the second support portion 69r is inserted into the second upper positioning hole 69r. The second roller lubrication member 70r is attached to the second support portion 69r.
Are attached using the second attachment member 71r.
In this way, the upper connecting member 62 is fixed or positioned with respect to the transmission case 22 by the first supporting portion 69f and the upper spherical bearing 75r.

In the lower connecting member 63, a first lower positioning hole 72f is formed at an intermediate portion between the first shaft hole 67a and the second shaft hole 67b, and an intermediate portion between the third shaft hole 67c and the fourth shaft hole 67d. A second lower positioning hole 72r is formed in the portion. And
In the first and second lower positioning holes 72f and 72r, respectively,
The first and second mounting bolts 74f, 7 are provided on the upper surface of the partition wall portion 53.
First and second lower spherical bearings 73 fixed using 4r
f and 73r are inserted. In this way, the lower connecting member 63 is fixed or positioned with respect to the partition wall portion 53 (transmission case 22) by the first and second lower spherical bearings 73f and 73r.

Gear Ratio Control Device Hereinafter, the first to fourth upper hydraulic chambers 79a to 79d and the first to fourth upper hydraulic chambers will be described.
A gear ratio control device that controls the gear ratio by controlling the supply of hydraulic pressure to the fourth lower hydraulic pressures 80a to 80d will be described. In the gear ratio control device, the first to fourth upper hydraulic chambers 79a to 79d and the first to fourth lower hydraulic chambers 80a are included.
To 80d from the gear ratio control valve V according to the operating state,
Hydraulic pressure is supplied via a hydraulic circuit which will be described later. The speed ratio control valve V is a so-called three-layer valve in which a sleeve 83 is fitted in the valve body 82 and a spool 84 is fitted in the sleeve 83, as will be described later in detail. It has a spring 85, a rotating member 86, a pin member 87, etc. as an operating mechanism, and is driven or controlled by a stepping motor 88 that operates according to a signal from a control unit (not shown), depending on the operating state. The hydraulic pressure (line pressure) received in the source pressure receiving port P1 is passed through the up-shift control port P2 or the down-shift control port P3 to the predetermined hydraulic chambers 79a-79d, 80a-. Hydraulic pressure is supplied to 80d. The speed ratio control valve V is provided with feedback means 90.

As shown in FIGS. 3, 4, and 8, in the gear ratio control valve V, a part of the lower housing 56 is a valve body 82, and in the valve body 82,
A sleeve 83 adapted to reciprocate in the axial direction of the valve body is fitted therein, and a spool 84 adapted to reciprocate in the axial direction of the valve body is fitted in the sleeve 83. . The sleeve 83 has a main port 8 which is always in communication with the source pressure receiving port P1.
3i, the first port 83j communicating with the control port P2 for always shifting up, and the control port P for constantly shifting down.
A second port 83k communicating with the third port 3 is formed. In addition, the spool 84 is formed with an annular group 84i which is always in communication with the main port 83i, and first and second land portions 84j and 84k which are adjacent to the left and right of the group 84i. There is.

FIG. 10 shows the details near the ports P1 to P3. In FIG. 10, the spool 84 is shown in the neutral position. At this time, the input port P1 is slightly communicated with the output ports P2 and P3 by the underlaps δ1 and δ2. In the neutral position, the output ports P2 and P3 are
With respect to the drain passage 99, underlap δ3, δ4
Slightly communicated by. At such a neutral position, a pressure difference between the output ports P2 and P3 does not occur. Then, the spool 84 slightly moves within the range of the underlap δ1 or δ2 to cause a predetermined pressure difference between the output ports P2 and P3, and this pressure difference causes the input torque to the input disks 43f and 43r. It becomes the size according to.

When shifting, the spool 84
By making a large movement to set the underlap δ1 or δ2 to zero, a large pressure difference is generated between the output ports P2 and P3. It should be noted that δ1 to δ4 are equal in size, and δ4 (δ3) is set to zero when δ1 (δ2) becomes zero. Further, as will be described later, at the time of feeding back by the feeding back means 90, the spool 84 is returned to a substantially neutral position. The returning position at this time is the input torque to the input disks 43f and 43r described above. The position is set according to the size of.

In the following description, for simplification, the port P1 is set to P when the spool 84 is in a substantially neutral position, that is, when both δ1 and δ2 are not zero.
2, P3 is blocked or ports P2, P3
Is cut off from the drain passage 99. Also, δ1
Is zero, the port P1 is in communication with P3 and the port P2 is drained. Furthermore, when δ2 is zero, the port P1 is in communication with P2 and the port P3 is drained.

The stepping motor 8 is attached to the side wall of the oil pan 36 attached to the lower end of the transmission case 22.
8, a rotary member 86 is fixed and connected to a rotary shaft 88i of the stepping motor 88. A male screw portion 86i is formed near the tip of the rotary member 86, and a female threaded collar 92 is screwed onto the male screw portion 86i. A pin member 87 is fixed to the female threaded collar 92, and both ends of the pin member 87 are
Locked by a pair of upper and lower groove portions 82i formed in the valve body 82, the female threaded collar 92 does not rotate about the valve body axis. Therefore, when the rotary member 86 rotates about the valve body axis, the female threaded collar 92 moves in the valve body axis direction. Further, the pin member 87 allows the internal threaded collar 92 and the sleeve 83 to interlock in the axial direction of the valve body.

Inside the sleeve 83, a collar with an internal thread
A spring 85 is provided between 92 and the spool 84 to apply a biasing force to each other so as to separate them from each other in the axial direction of the valve body. That is, the spring 85 constantly urges the female threaded collar 92 in the stepping motor direction (leftward in FIG. 7). In the following, this direction (left direction in FIG. 8) is simply referred to as “left” for convenience, unless otherwise specified.
The opposite direction (rightward in FIG. 8) is simply referred to as "right". Therefore, the sleeve 83 interlocking with the female threaded collar 92 is always biased to the left. Of course, the spring 85 always urges the spool 84 rightward.

When the rotating member 86 rotates with the rotation of the stepping motor 88, the female threaded collar 92, the rotation of which is restricted by the pin member 87, is moved in the axial direction of the valve body. The sleeve 83 moves in the axial direction of the valve body, and the main port 83i
Through the group 84i to communicate with the first port 83j or the second port 83k, and the operating oil (hydraulic pressure) in the source pressure receiving port P1 is shifted up to the control port P2. Alternatively, the data is output to the downshift control port P3. Note that the hydraulic oil here is hydraulic oil accompanied by a predetermined hydraulic pressure, and the same applies below.

Specifically, for example, at the time of upshifting, the stepping motor 88 rotates forward at a rotation angle corresponding to the pulse, and along with this, the sleeve 83 moves to the right, and at this time, the main port 83i. Communicates with the first port 83j via the group 84i, and the hydraulic oil of the source pressure receiving port P1 is output from the upshift control port P2. In this case, the hydraulic oil in the downshift control port P3 is released to the drain passage 99.

On the other hand, at the time of downshifting, the stepping motor 88 reversely rotates at a rotation angle corresponding to the pulse, and accordingly, the sleeve 83 moves leftward, and the main port 8 moves.
3i communicates with the second port 83k via the group 84i, and the working oil (hydraulic pressure) of the source pressure receiving port 1 is output from the downshift control port P3. In this case, the hydraulic oil in the shift-up control port P2 is released into the drain passage 99.

The right end portion of the spool 84 and the first shaft member 61a
A feed back means 90 is provided between the lower end portion and the lower end portion. The feedback means 90 is provided with a recess cam 100 which is fixed to the first shaft member 61a and rotates integrally with the first shaft member 61a.
An inclined surface 100i is formed at 0. Further, a first arm 102i and a second arm 102 are attached to a rotary shaft 101 rotatably provided at a predetermined position of the up housing 55.
j and are fixed. Here, the tip of the first arm 102i engages with the inclined surface 100i of the recess cam 100, and the tip of the second arm 102j engages with the slit 84m formed at the right end of the spool 84. I am fit.

Since the basic function of the feedback means 90 is similar to that of a generally used feedback means, detailed description thereof will be omitted, but each process is generally performed in the following process. The tilt angles (gear ratios) of the rollers 45a to 45d are maintained at the target tilt angle (target gear ratio). That is, when the stepping motor 88 rotates by an angle corresponding to the target tilt angle (target gear ratio) during gear shifting, the sleeve 83 moves in the valve body axial direction by an amount corresponding to this rotation angle, The hydraulic pressure is supplied to the predetermined hydraulic chambers 79a to 79d, 80a to 80d,
Each of the rollers 45a to 45d tilts to the target tilt angle. On the other hand, when the rollers 45a to 45d are tilted in this manner, the trunnions 59a to 59d and the shaft members 6 are correspondingly moved.
1a to 61d rotate, and along with this, the precess cam 1
00 rotates. At this time, the rotating shaft 101 is rotated by the precess cam 100 via the first arm 102i, and the rotating shaft 101 is further rotated by the second arm 1.
The spool 84 is moved in the same direction as the above-mentioned moving direction of the sleeve 83 via 02j, and the rollers 45a to 45a.
When the tilt angle of d reaches the target tilt angle, the spool 84
Of the sleeve 83 becomes equal to the amount of movement of the sleeve 83, where the communication between the main port 83i and the first port 83j or the second port 83k is cut off, and the hydraulic pressure 79
The supply of hydraulic pressure to a to 79d and 80a to 80d is stopped, the change of the tilt angle is stopped, and the tilt angle is maintained at the target value tilt angle. However, when the feedback operation is performed, the tip portion of the second arm 102j is scrolled by the spool 8.
When the spool 84 is displaced in the direction away from No. 4 (that is, rightward), the spring 85 causes the spool 84 to follow the displacement of the second arm 102j.

A hydraulic circuit for supplying hydraulic oil (hydraulic pressure) from the hydraulic pressure supply source to the gear ratio control valve V and for supplying hydraulic oil from the gear ratio control valve V to the hydraulic chambers 79a to 79d and 80a to 80d. explain. As shown in FIG. 9, in such a hydraulic circuit, after the hydraulic oil in the oil pan 36 is discharged from the oil pump 17, the line pressure control unit 12
After being adjusted to a predetermined pressure (source pressure) in step 2, the source pressure receiving port P1 of the gear ratio control valve V is supplied via the source pressure supply passage 123.
To be supplied to.

The hydraulic oil output from the shift-up control port P2 flows into the common shift-up oil passage 130, and thereafter, the first to fourth branch shift-up oil passages 130a to 13th.
0d to the first lower hydraulic chamber 80a, the second upper hydraulic chamber 79b, the third lower hydraulic chamber 80c, and the fourth upper hydraulic chamber 79d, respectively. On the other hand,
The hydraulic oil output from the shift-down control port P3 flows into the common shift-down oil passage 131, and thereafter,
The first upper hydraulic chamber 79a and the second lower hydraulic chamber 8 are respectively passed through the fourth branch downshift oil passages 131a to 131d.
0b, the third upper hydraulic chamber 79c, and the fourth lower hydraulic chamber 80.
It is designed to be supplied to d and.

In such a shift control device, the sleeve 83 moves to the right in FIG. 9 during downshifting,
The hydraulic oil of the source pressure receiving port P1 is the control port for downshifting.
Output from the engine P3, this hydraulic oil is supplied to the first upper hydraulic chamber 7
9a, a second lower hydraulic chamber 80b, and a third upper hydraulic chamber 79.
c and the fourth lower hydraulic chamber 80d. It should be noted that the gear ratio control valve V in FIG. 9 is shown to have a left-right positional relationship opposite to that of the gear ratio control valve V in FIGS. 4 and 8.

At this time, as shown in FIG.
Trunnions 59a, 59c, that is, rollers 45a, 45c
Is displaced upward, and the second and fourth trunnions 59b, 59
d, that is, the rollers 45b and 45d are displaced downward, and the first to fourth rollers 45a to 45d are decelerated by the displacement of the trunnions 59a to 59d. At this time, the trunnions 59a to 59d and therefore the shaft members 61a
Although the first cam member 61a rotates, the precess cam 100 is in contact with the inclined surface 100i along with the rotation of the first shaft member 61a, the first arm 102i, the rotary shaft 101, and the second arm.
The spool 84 is connected to the source pressure receiving port through the frame 102j.
The port P1 and the downshift control port P3 are moved to the right until the communication is cut off. In this way, the tilt angle is maintained at the target tilt angle.

On the other hand, when shifting up, the sleeve 83
1 moves to the left in FIG. 1, the hydraulic oil of the source pressure receiving port P1 is output from the shift-up control port P2, and this hydraulic oil is supplied to the first lower hydraulic chamber 80a and the second upper hydraulic chamber 80a. Hydraulic chamber 79b, third lower hydraulic chamber 80c, and fourth upper hydraulic chamber 7
9d.

At this time, as shown in FIG.
Trunnions 59a, 59c, that is, rollers 45a, 45c
Is displaced downward, and the second and fourth trunnions 59b, 59
d, that is, the rollers 45b and 45d are displaced upward, and the displacement of the trunnions 59a to 59d causes the first to fourth rollers 45a to 45d to change to the speed increasing side (overdrive side). At this time, each trunnion 59a to 59d
(Each shaft member 61a to 61d) rotates, and the first shaft member 6
With the rotation of 1a, the precess cam 100 moves the inclined surface 1
The first arm 102i in contact with 00i and the rotary shaft 1
01 and the second arm 102j are actuated. In this case, unlike the case of the above-mentioned downshift, the tip end of the second arm 102j is separated from the spool 84 (see FIG. In 1 it moves to the left). For this reason the spring 8
By the urging force of 5, the spool 84 is moved leftward in FIG. 1 and is moved leftward until the communication between the source pressure receiving port P1 and the upshift control port P2 is cut off. In this way, the tilt angle is maintained at the target tilt angle.

Entire Hydraulic Circuit (FIG. 11) The entire hydraulic circuit will be described with reference to FIG. First, the pressure discharged from the pump 17 is controlled by the line pressure control valve 122 so that the pressure in the line 123 becomes the line pressure. Further, the gear ratio control valve V described above is
A front gear ratio control valve Vf for forward rotation, that is, for forward traveling,
Two rear gear ratio control valves Vr for reverse traveling are provided. The front gear ratio control valve Vf is used during forward traveling when the input disk 43f (43r) rotates counterclockwise in FIGS. 6 and 7,
It is used when performing normal shift control. On the other hand, the rear speed ratio control valve Vr has the input disc 43f (43
r) is used during the backward traveling, which is rotating clockwise in FIGS. 6 and 7, but the toroidal type continuously variable transmissions 7 and 8 are not involved in the power transmission during the backward traveling. It is for fixing to a predetermined gear ratio. That is, the power rollers 45a to 45f are connected to the respective disks 43a.
The gear ratio is set so that it does not come into contact with the stopper provided at the stroke end of the tilt direction in the range that does not deviate from 43f to 43f with excessive force.

The changeover valve VFR selects which of the front and rear gear ratio control valves is used. This switching valve VF
When the front speed ratio control valve Vf is selected by R,
The line pressure is supplied to the input port P1 of the front speed ratio control valve Vf, and the output ports P2 and P3 are connected to the continuously variable transmissions 7 and 8 in the mode shown in FIG. At this time,
The ports P1 to P3 of the reverse gear ratio control valve Vr are closed.

When the reverse gear ratio control valve Vr is selected by the switching valve VFR, the rear gear ratio control valve Vr is selected.
The line pressure is supplied to the input port P1 of the output port P2, and the output ports P2 and P3 are connected to the continuously variable transmissions 7 and 8 in the mode shown in FIG. At this time, the ports P1 to P3 of the forward gear ratio control valve Vf are closed. The switching valve VFR is switched by the ON / OFF solenoid SL1.

A part of the oil discharged from the pump 17 is discharged to the line 150, and the pressure in the line 150 is reduced to a constant pressure by the pressure reducing valve 151. The pressure of the line 150, which has been set to this constant pressure, is the same as the electromagnetic control valve 1
In addition to being used as control pressure for line pressure control by 22a, it is also used as pressure for clutch control, which will be described later. The surplus oil generated by the line pressure control valve 122 passes through the line 152, and the toroidal type continuously variable transmission 7,
8 is used for lubrication, and the pressure in the line 152 is controlled by the pressure reducing valve 153 so as to be a predetermined pressure.

The drain passage 99 of each gear ratio control valve Vf, Vr is connected to the common relief line 171 from the relief line 171f or 171r through the check ball type switching valve 172. The relief valve 161 is connected to the common relief line 171. A line 152 branched from the line 150 having a constant pressure is connected to the relief valve 161. The line 152 is for controlling the relief pressure from the relief valve 161, and therefore the pressure in the line 152 is controlled by the electromagnetic control valve 162 (linear solenoid).

A branch line 123i branched from the line pressure passage 123 is connected to forward / reverse switching clutches 31 and 33. In FIG. 11, 201 is a switching valve for supplying line pressure to either the forward clutch 33 or the reverse clutch 31 to switch between a forward drive state and a reverse drive state. The switching valve 201 also switches the supply of lubricating oil to the forward and backward clutches 31 and 32. Such switching of the switching valve 201 is performed by
This is performed by the ON / OFF solenoid SL2. Further, in FIG. 11, 203 is a proportional solenoid 205.
Is a clutch pressure control valve for gradually changing the line pressure to form a half-clutch state.
Switching during the formation of this half-clutch state is O
This is performed by the FF solenoid SL3.

Outline of Relief Pressure Control (FIGS. 12 to 15) Next, referring to FIGS. 12 to 15, the electromagnetic control valve 162 will be described.
An outline of the relief pressure control by will be described. First,
In FIG. 13, U is a control unit configured by using a microcomputer, which receives inputs from the sensors S1 to S4 and outputs them to the electromagnetic control valve 162. The sensor S1 detects the pressure of the line 123, that is, the line pressure LINE, the sensor S2 detects the pressure of the common relief passage 171, that is, the relief pressure PR, and the sensor S3 detects the pressure of the output port P2, that is, the line 130. Is to detect the pressure PP2 of
The sensor S4 detects the pressure at the output port P3, that is, the line 131.
The pressure of is detected. The control unit U
Also performs switching control of the clutches 31 and 33, etc., but since this point is not directly related to the present invention, its explanation is omitted.

FIG. 12 shows a toroidal type continuously variable transmission 7, 8.
Shows the pressure difference ΔP between the pressures P1 and P2 supplied from the transmission ratio control valve V with respect to the input torque. As is clear from FIG. 12, when the gear ratio is the same, the pressure difference ΔP is increased as the input torque increases. Further, if the input torque is the same, the pressure difference ΔP is increased as the gear ratio is on the deceleration side, that is, on the lower speed side.

In FIG. 14, the pressure at the port P1, that is, the line pressure is shown as PLINE, and the larger pressure of the pressures PP2 and PP3 at the ports P2 and P3 is PH,
The smaller pressure is shown as PL and the relief pressure is shown as PR. In FIG. 14, the pressure difference ΔP is
It is a pressure difference required for gear shifting, and the line pressure LINE is set as a constant pressure so as to obtain this pressure difference. Then, during the shift control, what is shown as ΔH and ΔL in FIG. 14 corresponds to the amount of oil consumed during the shift. This ΔH is the line pressure LINE and the high hydraulic pressure PH.
Is the differential pressure between the low pressure PL and the relief pressure PR.
It is the differential pressure with. Therefore, by decreasing this ΔH or ΔL, the amount of oil consumed at the time of shifting is reduced, but this ΔH or ΔL is decreased by increasing the relief pressure PR. . With such a method, as shown in FIG. 15, while maintaining a predetermined pressure difference ΔP, ΔH and ΔL are reduced to reduce the oil consumption.

In order to reduce the amount of oil consumed during a shift, the present invention adopts the following four methods. The first method is to increase the relief pressure PR as the low hydraulic pressure PL increases. The second method is line pressure LINE
The higher the differential pressure between the high hydraulic pressure PH and the high hydraulic pressure PH, the higher the relief pressure PR.

In the third method, when the differential pressure between the high hydraulic pressure PH and the low hydraulic pressure PL is set to a predetermined differential pressure ΔP, this predetermined differential pressure ΔP.
Is smaller, the relief pressure PR is to be increased. In the fourth method, when the differential pressure between the high hydraulic pressure PH and the low hydraulic pressure PL is set to a predetermined differential pressure ΔP, the relief pressure PR increases as the difference between the predetermined differential pressure ΔP and the line pressure PLINE increases. It is to be.

Concrete example of relief pressure control (electronic control, FIG.
6 to 19) FIGS. 16 to 19 show control examples for performing the above-mentioned four methods by electronic control. In the following description, Q indicates a step. First, FIG. 16 is for performing the above-mentioned first method, and in Q1, each output port P2, P
The pressures PP2 and PP3 of 3 are read. Then Q2
At, it is determined whether PP2 is equal to or greater than PP3. When the determination in Q2 is YES, the pressure PP3 is set as the low hydraulic pressure PL in Q3. When the determination in Q2 is NO, the pressure PP2 is set to the low hydraulic pressure PL in Q4.

After Q3 or Q4, in Q5,
The relief pressure PR is read. Thereafter, in Q6, it is determined whether or not the absolute value of the difference between the relief pressure PR and the low oil pressure PL is equal to or less than a predetermined value Δk. When the determination in Q6 is YES, the relief pressure PR is a sufficiently high value in the vicinity of the low oil pressure PL, and therefore the engine is returned as it is at this time.

When the determination in Q6 is NO, it is determined in Q7 whether the low hydraulic pressure PL is equal to or higher than the relief pressure PR. If YES in the determination in Q7, in Q8, a predetermined correction value Δ with respect to the current relief pressure PR.
A value obtained by adding C is set as the relief pressure PR of this time. When going through the process of Q8, the relief pressure PR
Is gradually approaching the low oil pressure PL. Q7
When the determination is NO, the value obtained by subtracting the predetermined correction value ΔC from the current relief pressure PR is set as the current relief pressure PR in Q9. Of course, the proportional solenoid 162 of FIG. 11 is controlled so that the relief pressure PR set by Q8 or Q9 is obtained. As is already clear from the above description, the relief pressure PR is controlled to be a value within the range of the predetermined allowable width Δk with respect to the low oil pressure PL.

FIG. 17 shows a control example for performing the above-mentioned second method. First, in Q11, after the pressures PP2 and PP3 are read, the higher pressure of the pressures PP2 and PP3 is set as the high hydraulic pressure PH by the processing of Q12 to Q14. Q16-Q19 are
This corresponds to the processing of Q5 to Q9 in FIG. That is, in Q15, the relief pressure PR is read. After that, in Q16, it is determined whether or not the absolute value of the difference between the line pressure LINE and the high hydraulic pressure PH is equal to or less than a predetermined value Δk. If the determination in Q16 is YES, it means that the high hydraulic pressure PH is a sufficiently high value near the line pressure LINE, so at this time it is returned as it is.

If NO in Q16, it is determined in Q17 whether or not the line pressure LINE is higher than the high hydraulic pressure PH. If YES in the determination in Q17,
In Q18, a value obtained by adding a predetermined correction value ΔC to the current relief pressure PR is set as the current relief pressure PR. When the process of Q18 is performed, the high hydraulic pressure PH is gradually brought closer to the line pressure LINE. If NO in Q17, in Q19, a value obtained by subtracting a predetermined correction value ΔC from the current relief pressure PR is set as the current relief pressure PR.
Of course, the proportional solenoid 162 of FIG. 11 is controlled so that the relief pressure PR set by Q18 or Q19 is obtained.

FIG. 18 shows a control example for performing the above-mentioned third method. First, at Q21, pressures PP2 and P
After P3 and P3 are read, in Q22, PP2 and P
The difference from P3 is set as the predetermined differential pressure ΔP. Q23
Then, it is determined whether or not the absolute value of the predetermined differential pressure is greater than or equal to the predetermined value Δk. If NO in Q23, Q24
In, the current relief pressure PR is set by adding a predetermined correction value ΔC to the current relief pressure PR. If the answer in Q23 is YES, Q2
5, a predetermined correction value Δ from the relief pressure PR of this time
By subtracting C, the relief pressure PR of this time is set. In the present control example, when the predetermined differential pressure ΔP is small, there is a large margin for increasing the relief pressure PR (large ΔH and ΔL in FIG. 14). .

FIG. 19 shows a control example for performing the above-mentioned fourth method. First, in Q31, the pressures PP2, P
After P3 and LINE are read, the absolute value of the difference between PP2 and PP3 is set as the predetermined differential pressure ΔP in Q32. At Q33, it is judged if the difference between the line pressure LINE and the predetermined differential pressure ΔP is a predetermined value Δk or more. If YES in the determination in Q33, in Q34, a predetermined correction value Δ for the relief pressure PR of this time is set.
The relief pressure PR of this time is set by adding C. When the determination in Q33 is NO, in Q35, the current relief pressure PR is set by subtracting the predetermined correction value Δc from the current relief pressure PR.

Concrete example of relief pressure control (hydraulic operation type, FIG.
20 to 23) FIGS. 20 to 23 show the above-described first to fourth methods which are hydraulically operated instead of electronically controlled. First, FIG. 20 corresponds to the first method, in which the control oil pressure of the relief valve 161 is changed to the oil pressure control valve 210.
It is controlled by. Hydraulic control valve 210
Receives the two pressures PP2 and PP3, and when the pressure PP2 is greater than PP3, the spool 21
0i is set to the right side in the figure, and the lower pressure P is supplied to the pilot pressure supply line 211 for the relief valve 161.
P3 is supplied. Further, when the pressure PP2 is lower than PP3, the spool 210i is set to the left position in the figure, and the lower pressure PP2 is supplied to the pilot pressure supply line 211 for the relief valve 161. Thus, the pressure supplied to the pilot pressure passage 211 is
Low oil pressure PL is used. In the relief valve 161, as the pressure from the pilot passage 211, that is, the low hydraulic pressure PL, increases, the degree of draining the relief passage 171 decreases and the relief pressure PR increases.

FIG. 21 corresponds to the second method. In this example, the relief valve 161B corresponding to the relief valve 161 of FIG. 20 is operated by receiving the line pressure LINE from the line pressure passage 123 and the pilot pressure from the pilot passage 212. In this operation, as the line pressure LINE is larger than the pressure in the pilot passage 212, the degree of draining the relief passage 171 is reduced and the relief pressure PRPR is increased. Then, the pressure in the pilot passage 212 is set by the control valve 220 to the larger one of the pressures PP2 and PP3, that is, the high hydraulic pressure PH. More specifically, the control valve 22
For the spool 220i of 0, the pressure PP3 is higher than the pressure PP2.
When it is larger than the above value, it is located on the left side in the drawing, and the larger pressure PP3 is supplied to the pilot passage 212. On the contrary, when the pressure PP2 is higher than the pressure PP3, the spool 220i is positioned at the right side in the figure, and the higher pressure PP2 is supplied to the pilot passage 212.

FIG. 22 corresponds to the fourth method. In this example, the relief valve 161C corresponding to the relief valve 161 of FIG. 20 receives the line pressure LINE and the pressure from the two pilot passages 213 and 214.
The spool 230i of the control valve 230 corresponding to the control valve 220 of FIG. 20 is set to the left position in the drawing when the pressure PP3 is higher than PP2, and the pressure in the pilot passage 214 is set to the higher pressure PP3, that is, PH. On the other hand, the pressure in the pilot passage 213 is set to the lower pressure PP2, that is, the low hydraulic pressure PL. Further, when PP2 is larger than pressure PP3, the spool 230i is set to the right position in the drawing to set the pressure in the pilot passage 214 to the larger pressure PP2, that is, the high hydraulic pressure PH, while the pilot passage 2
The pressure of 13 is set to the smaller pressure PP3, that is, the low hydraulic pressure PL. In this way, the pressure in the pilot passage 214 is always the high hydraulic pressure PH, and the pressure in the pilot passage 213 is always the low hydraulic pressure PL.

The relief valve 161C has a line pressure PLIN.
The larger E is, the more to the left in the figure, the relief passage 1
Although the degree of draining 71 is reduced to increase the relief pressure PR, the high hydraulic pressure PH from the pilot passage 214 is increased.
Has an action against the line pressure PLINE, and the low hydraulic pressure PL from the pilot passage 213 has the action of reducing the action of the high hydraulic pressure PH. That is, when the differential pressure between the high hydraulic pressure PH and the low hydraulic pressure PL is a predetermined differential pressure, the relief pressure PR is increased as the differential pressure between the predetermined differential pressure and the line pressure PLINE is increased.

FIG. 23 corresponds to the third method. In this example, the relief valve 161C in FIG.
The relief valve 161D corresponding to the
Is set not to be used as the pilot hydraulic pressure, and other points are the same as in the case of FIG. That is, the relief valve 161D increases the relief pressure PR as the pressure difference between the high hydraulic pressure PH and the low hydraulic pressure PL decreases.

Modified Example (FIG. 24) FIG. 24 shows a modified example of the present invention, in which the positional relationship of the hydraulic cylinders 76a and 76b as the speed change actuator with respect to the speed ratio control valve V is changed. It is a thing. That is, the gear ratio control valve V is arranged at a low position in the oil pan 36, but the hydraulic cylinders 76a and 76b.
Is disposed at a high position on the opposite side of the transmission ratio control valve V with the power rollers 45a and 45b interposed therebetween. Hydraulic cylinder 76
Although a and 76b act to deform the members around them because they generate a large driving force, the hydraulic cylinders 76a and 76b have the above-mentioned arrangement relationship.
This is preferable for preventing the gear ratio control valve V from being deformed due to the driving force of b, and ensuring the reliable operation of the gear ratio control valve V.

Modified Example (FIG. 25) FIG. 25 shows a modified example of the gear ratio control valve. The same components as those of the member shown in FIG. 8 are designated by the same reference numerals and the description thereof will be omitted. In the gear ratio control valve VB in this example, the spool 84 is directly fitted to the valve body 82. Further, the drive member 110 is attached to the rotary shaft 88i of the motor 88.
On the other hand, a pin 111 is fixed to the spool 84, and the pin 111 is connected to the drive member 110 so as to rotate integrally therewith and slidably displaced in the axial direction. Further, the spool 84 has a male screw member 11
3 is screwed, and this male screw member 113 is engaged with the arm 102j as the feedback means. In such a configuration, the spool 84 is moved in the axial direction with the rotation of the motor 88, and the amount of this movement is changed according to the rotational position of the arm 102j as the feedback means.

[Brief description of drawings]

FIG. 1 is an overall mechanical view showing an example of a toroidal type continuously variable transmission.

FIG. 2 is a plan sectional view of a toroidal type continuously variable transmission portion.

FIG. 3 is a side sectional view of the toroidal type continuously variable transmission shown in FIG.

FIG. 4 is a sectional view taken along line AA of FIG.

5 is a sectional view taken along line BB of FIG.

FIG. 6 is an explanatory diagram showing the principle of changing the gear ratio, showing the time of upshifting.

FIG. 7 is an explanatory view showing a principle of changing a gear ratio, showing a downshift.

FIG. 8 is a sectional view of a gear ratio control valve portion.

FIG. 9 is a hydraulic circuit diagram showing a connection relationship between a gear ratio control valve and a hydraulic cylinder for changing a gear ratio.

FIG. 10 is a sectional view showing an input / output port portion of the gear ratio control valve.

FIG. 11 is an overall hydraulic circuit diagram of the toroidal type continuously variable transmission.

FIG. 12 is a diagram showing a relationship between an input torque and a differential pressure required for shifting.

FIG. 13 is a diagram showing a control system when electronically controlling the relief pressure.

FIG. 14 is a diagram for explaining the principle of oil consumption reduction during shifting, showing the state when relief pressure control is not being performed.

FIG. 15 is a view corresponding to FIG. 14, showing a state when the oil consumption is reduced by performing relief pressure control.

FIG. 16 is a diagram showing a first example when electronically controlling relief pressure control.

FIG. 17 is a diagram showing a second example in which relief pressure control is electronically controlled.

FIG. 18 is a diagram showing a third example in which relief pressure control is electronically controlled.

FIG. 19 is a diagram showing a fourth example in which relief pressure control is electronically controlled.

FIG. 20 is a diagram showing a first example in which relief pressure control is hydraulically operated.

FIG. 21 is a diagram showing a second example in which relief pressure control is hydraulically operated.

FIG. 22 is a diagram showing a third example in the case where relief pressure control is hydraulically operated.

FIG. 23 is a diagram showing a fourth example in which relief pressure control is hydraulically operated.

FIG. 24 is a diagram showing an example of changing the arrangement positions of a gear ratio control valve and a hydraulic cylinder.

FIG. 25 is a view showing a modified example of the gear ratio control valve.

[Explanation of sign]

7, 8: Toroidal type continuously variable transmission 17: Oil pump 43: Input disc 44: Output disc 45: Power Roller 59: Trunnion 76: Hydraulic cylinder (for changing gear ratio) 161 to 161D: Relief valve 162: Proportional solenoid (for relief pressure control) 210, 220, 230: control valve PH: High hydraulic pressure PL: Low hydraulic pressure ΔP: predetermined differential pressure LINE: Line pressure PR: Relief pressure V, VB: Gear ratio control valve P1: Input port P2, P3: Output port

─────────────────────────────────────────────────── ─── Continuation of the front page (56) References JP-A-62-171557 (JP, A) JP-A-62-127554 (JP, A) JP-A-5-39847 (JP, A) Actual development Sho-62- 199562 (JP, U) (58) Fields investigated (Int.Cl. ⁷ , DB name) F16H 15/38 F16H 59/00-61/12 F16H 61/16-61/24 F16H 63/40-63/48

Claims (4)

(57) [Claims] 1. A toroidal type continuously variable transmission comprising a hydraulic gear ratio adjusting means, wherein hydraulic pressure is supplied to and discharged from the gear ratio adjusting means by a gear ratio control valve. is connected to the relief port is provided with a relief valve for adjusting the relief pressure as the amount of oil is relieved from the speed-change ratio control valve is reduced, the speed ratio adjusting means, to perform the speed change by the hydraulic pressure difference
The gear ratio control valve is set to the front of the gear ratio adjusting means.
Supply high and low hydraulic pressure to give hydraulic pressure difference.
Is set so that the relief valve becomes
A hydraulic control device for a toroidal-type continuously variable transmission , which is set so as to increase leaf pressure . 2. A gear ratio adjusting means of hydraulic type is provided, and the gear ratio is provided.
Supply and discharge of hydraulic pressure to the adjusting means by the gear ratio control valve
In the toroidal type continuously variable transmission, the gear ratio control valve is connected to a relief port,
To reduce the amount of oil that is relieved from the ratio control valve,
A relief valve for adjusting the leaf pressure is provided, and the gear ratio adjusting means performs gear shifting by a hydraulic pressure difference.
The gear ratio control valve is set to the front of the gear ratio adjusting means.
Supply high and low hydraulic pressure to give hydraulic pressure difference.
The relief valve has a large differential pressure between the line pressure and the high hydraulic pressure.
The higher the pressure, the higher the relief pressure is set.
Of the toroidal type continuously variable transmission characterized by
Hydraulic control device . 3. A gear ratio adjusting means of hydraulic type is provided, and the gear ratio
Supply and discharge of hydraulic pressure to the adjusting means by the gear ratio control valve
In the toroidal type continuously variable transmission, the gear ratio control valve is connected to a relief port,
To reduce the amount of oil that is relieved from the ratio control valve ,
A relief valve for adjusting the leaf pressure is provided, and the gear ratio adjusting means performs gear shifting by a hydraulic pressure difference.
The gear ratio control valve is set to the front of the gear ratio adjusting means.
Supply high and low hydraulic pressure to give hydraulic pressure difference.
Is set so that the pressure difference between the high oil pressure and the low oil pressure is small.
Is set to increase the relief pressure
Toroidal type continuously variable transmission oil characterized by
Pressure control device. 4. A hydraulic gear ratio adjusting means is provided, and the gear ratio is provided.
Supply and discharge of hydraulic pressure to the adjusting means by the gear ratio control valve
In the toroidal type continuously variable transmission, the gear ratio control valve is connected to a relief port,
To reduce the amount of oil that is relieved from the ratio control valve,
A relief valve for adjusting the leaf pressure is provided, and the gear ratio adjusting means performs gear shifting by a hydraulic pressure difference.
The gear ratio control valve is set to the front of the gear ratio adjusting means.
Supply high and low hydraulic pressure to give hydraulic pressure difference.
The relief valve sets the differential pressure between the high hydraulic pressure and the low hydraulic pressure to a predetermined value.
When the differential pressure is set, the difference between the predetermined differential pressure and the line pressure is large.
The higher the pressure, the higher the relief pressure is set.
Of the toroidal type continuously variable transmission characterized by
Hydraulic control device .