JPH08326893A - Continuously variable transmission - Google Patents

Abstract

PURPOSE: To change the axial force of both pulleys by forward-rearward changeover in a continuously variable transmission in which the torque transmission direction of its belt type continuously variable transmission is changed over through circulation of torque. CONSTITUTION: In the case of range D low mode, high pressure oil from a regulator valve 56 is supplied to a secondary hydraulic servo 33 via ports (f), (l) of a manual valve 59 while low pressure oil from a ratio control valve 57 to a primary hydraulic servo 32 through ports (h), (g). The supply of pressure oil is inverted through the shift of the valve 59 to range R.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a continuously variable transmission using a belt type continuously variable transmission in which a belt is wound around a primary pulley and a secondary pulley each having two sheaves. The present invention relates to an axial force control of a belt type continuously variable transmission when switching between forward and backward movements in a continuously variable transmission including a combination of the belt type continuously variable transmission and a planetary gear.

[0002]

2. Description of the Related Art Recently, an automatic continuously variable transmission incorporating a belt type continuously variable transmission as a transmission of an automobile has been attracting attention due to demands such as improvement of fuel consumption rate and improvement of driving performance.

Conventionally, as disclosed in, for example, Japanese Unexamined Patent Publication No. 56-52653, a planetary gear train capable of shifting gears between two stages of direct connection and forward rotation deceleration, a belt type continuously variable transmission, and the planetary gear train described above. And a planetary gear device that combines and outputs the output rotation of the belt type continuously variable transmission and the output rotation of the belt type continuously variable transmission. , A continuously variable transmission that continuously changes gears in the forward and reverse directions has been devised.

In the continuously variable transmission, the planetary gear train is decelerated and positively rotated by operating the select lever to the D position.
This is combined with the variable speed rotation of the belt type continuously variable device to extract the positive direction continuously variable speed rotation from the planetary gear device, and by operating the select lever to the R position, the planetary gear train is directly connected and rotated. In combination with the rotation of the belt type continuously variable transmission, the reverse continuously variable rotation is extracted from the planetary gear unit. At this time, the belt type continuously variable transmission is
Due to the torque circulation, in the forward drive state (D position) and the reverse drive state (R position), the power transmission direction, that is, the drive / driven relationship between the primary pulley and the secondary pulley is switched.

[0005]

Generally, in a belt type continuously variable transmission, a pulley ratio is changed by changing an axial force (belt clamping pressure) acting on each of a driving side pulley and a driven side pulley. are doing. Then, in order to set both pulleys to a predetermined pulley ratio and transmit power, it is necessary to set the ratio of the axial force acting on both pulleys to a predetermined value.
In this case, the axial force of the driving pulley needs to be larger than the axial force of the driven pulley.

However, in the continuously variable transmission described above, due to torque circulation, the power transmission direction of the belt type continuously variable transmission is changed between forward rotation and reverse rotation, but the axial force of the pulley is Is not taken into consideration, the shift control at the time of switching between forward and reverse is troublesome, it takes time to stabilize,
There is a possibility that a response may be delayed.

Therefore, the present invention is
An object of the present invention is to provide a continuously variable transmission that solves the above problems by changing the axial force acting on both pulleys of a belt type continuously variable transmission.

[0008]

The present invention according to claim 1 has been made in view of the above circumstances, and includes an input shaft (3) interlocking with an engine output shaft (2) and an output interlocking with wheels. A shaft (5), a first pulley (7) interlocking with the input shaft, a second pulley (9), a belt (10) wound around these pulleys, and the first and second pulleys. A belt type continuously variable transmission (11) having axial force actuating means (32) (33) for exerting an axial force on both pulleys to change the pulley ratio, and at least first, second and third rotating elements. (19s) (19c) (19r), and the first
A planetary gear (3) in which the rotating element of the second rotating element is linked to the input shaft (3), the second rotating element is linked to the second pulley (9), and the third rotating element is linked to the output shaft (5). 19) and at least a selecting means for selecting a forward drive state and a reverse drive state of the vehicle, based on the switching between the forward drive state and the reverse drive state of the selecting means, the first and the first of the belt type continuously variable transmission (11). In the continuously variable transmission (1), the torque transmission direction between the two pulleys (7) and (9) is changed, and the output torque direction of the output shaft is changed. Control means (54, 54 ', 54 ₁ , 54 ₂ ) for controlling the axial force actuating means so that the axial force acting on the pulley produces a difference corresponding to the pulley ratio,
Switching means (59, 59 ') for changing so that the magnitude relation of the axial forces acting on the first and second pulleys by the control means is reversed based on the switching between the forward drive state and the reverse drive state by the selecting means, It is characterized by comprising.

In the invention according to claim 2, the axial force actuating means includes a first hydraulic servo (32) acting on the first pulley (7) and the second pulley (9). And a second hydraulic servo (33) that acts on the first and second hydraulic servos to supply a hydraulic pressure based on an oil pump (17) to the first and second pulleys. The switching means 8 is a switching valve (59, 59 ') for switching between the hydraulic pressure acting on the first hydraulic servo and the hydraulic pressure acting on the second hydraulic servo.

In the invention according to claim 3, the switching valve is a manual valve (59, 59 ') which is integrally linked with the selecting means.

In the invention according to claim 4, for example, as shown in FIGS. 4 to 8 and 14, the first and second hydraulic servos (32) and (33) are respectively provided with a plurality of hydraulic pressures. Comprising chambers (45) (47) (46) (49), the switching valve (59) selectively switches between the hydraulic chambers for supplying hydraulic pressure to operate the first and second hydraulic servos. The effective pressure receiving area is reversed.

In the invention according to claim 5, the first
And the second hydraulic servos (32) (33) each have at least a first hydraulic chamber (45) (46) and a second hydraulic chamber (47) (49), respectively, The first hydraulic chambers (45) (46) have the same effective pressure receiving area, and the control means (54, 54 ₁ ) has a regulator valve (56) and a ratio control valve (57). Is constantly supplied to the first hydraulic chambers (45) (46) of the both hydraulic servos, and the hydraulic pressure of the ratio control valve (57) is supplied to the second hydraulic chamber of either of the hydraulic servos. The switching valve (59) is supplied to the hydraulic chamber (47 or 49), and the switching valve (59) is the ratio control valve (57).
And the communication between the second hydraulic chamber (49 or 47) is reversed to the other.

In the invention according to claim 6, the regulator valve (56) and the ratio control valve (57) are arranged between the oil pump (17) and the switching valve (59). Become.

[0014] In the invention according to claim 7, for example, as shown in FIGS. 9 to 13 and 15, wherein said control means (54 ', 54 _2), the hydraulic pressure based on the oil pump (17) It has first and second regulator valves (56) (57) for adjusting the respective hydraulic pressures, and these first
One of the first and second hydraulic servos, and the other of the regulator valves communicates with the other of the hydraulic servos, respectively, and the switching valve (59 ') Is an inversion of communication between the first and second regulator valves and the first and second hydraulic servos.

In the invention according to claim 8, the first
The second regulator valve (56) (57) is arranged between the oil pump (17) and the switching valve (59).

According to a ninth aspect of the present invention, the hydraulic pressure of the hydraulic servo having a high hydraulic pressure is changed from the hydraulic pressure of the hydraulic servo having a low hydraulic pressure based on the determination of the coasting state of the vehicle. A relief valve (58) for lowering the pressure is provided.

According to the tenth aspect of the present invention, as shown in FIGS. 1 and 24 to 26, for example, the input shaft (3) receives the torque transmitted from the first pulley (7). The first pulley and the first rotation are such that the torque direction in the input shaft and the torque direction in the input shaft of the torque transmitted from the first rotating element are opposite to each other. It is connected to each element.

[0018]

Based on the above construction, the engine output from the input shaft (3) is appropriately shifted by the belt type continuously variable transmission (11) and transmitted to the second rotating element of the planetary gear (19), and at the same time, it is constant. The fast rotation is transmitted to the first rotating element, and both of these rotations are combined by the planetary gear (19) and taken out from the third rotating element to the output shaft (5).

The control means (54, 54 ', 54)
By controlling the axial force actuating means (32) (33) with ₁ , 54 ₂ ), the first or second pulley (7) (9)
Axial force of the belt is changed, belt type continuously variable transmission (11)
Is changed to a predetermined pulley ratio and the shift control is appropriately performed, and when the selection means switches to a forward drive state or a reverse drive state,
Due to the torque circulation, the output torque direction of the output shaft (5) is changed to forward rotation and reverse rotation, and the torque transmission direction of the belt type continuously variable transmission (11) is changed.

At this time, the switching means (54, 54 ', 54)
₁ and 54 ₂ ), the magnitude relationship between the axial forces acting on the first and second pulleys (7) and (9) is adjusted so as to match the changed torque transmission direction of the belt type continuously variable transmission. It is reversed.

The reference numerals in parentheses are for comparison with the drawings, but do not limit the structure of the present invention.

[0022]

According to the first aspect of the present invention, by switching between the forward drive state and the reverse drive state of the vehicle, the switching means inverts the magnitude relation of the axial forces acting on the two pulleys, and the torque transmission direction based on the torque circulation. It is possible to match the axial force relationship between both pulleys of the belt type continuously variable transmission which has been reversed, and it is possible to perform forward / reverse switching reliably and quickly.

According to the second aspect of the present invention, since the hydraulic pressures acting on the first and second hydraulic servos are reversed by the switching valve, it is not necessary to control the hydraulic pressures of both hydraulic servos individually, and the switching is performed. By switching the oil passage by means of a valve, forward / backward switching can be performed easily and reliably.

According to the third aspect of the present invention, since the switching valve is a manual valve, it is possible to reliably switch between forward and reverse movements based on the driver's intention.

According to the fourth aspect of the present invention, since the first and second hydraulic servos each have a plurality of hydraulic chambers, the hydraulic pressure does not have to be controlled even if the pressure control acting on the hydraulic servos is controlled. By switching the hydraulic pressure supply to the chamber, the effective pressure receiving areas of both hydraulic servos can be reversed, and forward / backward switching can be performed easily and reliably.

According to the fifth aspect of the present invention, the hydraulic pressure is constantly supplied to the first hydraulic chambers of both hydraulic servos, and the hydraulic pressure supply to one of the second hydraulic chambers of both hydraulic servos is switched. As a result, it is possible to easily and reliably switch between forward and reverse.

According to the sixth aspect of the present invention, when arranging the regulator valve and the ratio control valve on the downstream side of the switching valve, it is necessary to increase the number of valves or the function of one valve. However, by arranging the regulator valve and the ratio control valve between the oil pump and the switching valve, the above need is eliminated and the configuration can be simplified.

According to the seventh aspect of the present invention, the two regulator valves regulate different hydraulic pressures, and the supply of the two regulated pressures is switched by the both hydraulic servos in the forward traveling state and the reverse traveling state. Reverse the axial force of
It is possible to reliably and easily switch between forward and reverse.

According to the present invention of claim 8, claim 6
Similarly, it does not require an increase in the number and function of valves,
It can have a simple configuration.

According to the ninth aspect of the present invention, the relief valve reduces the hydraulic pressure of the hydraulic servo acting on the high hydraulic pressure. Therefore, even in the coast state in which the transmission torque direction is reversed, the torque transmission direction is matched. The axial force acting on the pulley is changed so that the coast state can be reliably maintained.

According to the tenth aspect of the present invention, torque circulation is reliably performed by the planetary gears.

[0032]

Embodiments of the present invention will be described below with reference to the drawings.

First, the power transmission mechanism of the continuously variable transmission according to the present invention will be described with reference to FIG. 1 (a). The in-vehicle automatic continuously variable transmission 1 includes a first shaft 3, a second shaft 5 aligned with an engine crankshaft 2, and a third shaft 6 aligned with a front axle.
(A, b), a primary (first) pulley 7 is supported on the first shaft 3, and a secondary (second) pulley 9 is supported on the second shaft 5, A belt 10 is wound around these pulleys 7 and 9 to form a belt type continuously variable transmission 11.

Further, the first shaft 3 is connected to the engine crankshaft 2 via a damper 12 to form an input shaft, and the input shaft 3 has an input side member 1 of the low clutch C _L.
3, the output side member 15 is rotatably supported, and the output side member 15 is integrally connected with a primary side sprocket 16 which constitutes a power transmission means. A fixed sheave 7a of the primary pulley 7 is fixed to the input shaft 3, and an oil pump 17 is connected to the tip of the fixed sheave 7a. A movable sheave 7b moves axially in the fixed sheave 7a. Supported as possible.

A secondary pulley 9 is rotatably supported on the second shaft 5, and the secondary pulley 9 includes a fixed sheave 9a, a movable sheave 9b and a fixed sheave which are axially movably supported by the fixed sheave. It has the secondary shaft 9c connected integrally. Further, a high clutch C _H is provided between the second shaft 5 and the secondary shaft 9c, a planetary gear 19 is arranged on the second shaft 5, and the secondary side sprocket 20 is rotatable. Supported by. An output gear 21 is fixed to one end of the second shaft 5 to form an output shaft.

The planetary gear 19 is a sun gear 19
s, a ring gear 19r, and a single pinion planetary gear having a carrier 19c that rotatably supports one pinion 19p that meshes with both gears. The sun gear 19s is connected to the secondary shaft 9c to form a second rotating element, the ring gear 19r is connected to the output shaft 5 to form a third rotating element, and the carrier 19c is connected to the secondary sprocket 20. To form a first rotating element.
Also, the primary side and secondary side sprocket 1
A winding body 22 such as a silent chain, a roller chain, or a timing belt is wound around 6 and 20.

The gear 21 fixed to the output shaft 5 meshes with the large gear 23a of the reduction gear unit 23, and the small gear 23b of the unit meshes with the ring gear 24 of the differential device 25. The differential device 25 outputs differential rotations to the left and right axle shafts 6a and 6b forming the third shaft.

Next, the operation of the continuously variable transmission 1 based on the power transmission mechanism will be described with reference to FIGS. 1 (a), 1 (b), 2 and 3. FIG. The rotation of the engine crankshaft 2 is transmitted to the input shaft 3 via the damper 12. In the low mode in which the low clutch C _L is connected and the high clutch C _H is disconnected, the rotation of the input shaft 3 is transmitted to the primary pulley 7, the primary side sprocket 16 and the wrapping body 22. And to the carrier 19c of the planetary gear 19 via the power transmission device 26 including the secondary side sprocket 20. On the other hand, the rotation of the primary pulley 7 is continuously changed and transmitted to the secondary pulley 9 by appropriately adjusting the pulley ratio of the primary and secondary pulleys by an axial force actuating means such as a hydraulic servo which will be described later. The variable speed rotation of the pulley 9 is transmitted to the sun gear 19s of the planetary gear 19.

In the planetary gear 19, as shown in the velocity diagram of FIG. 1 (b), the carrier 19c to which constant speed rotation is transmitted via the power transmission device 26 becomes a reaction force element.
The continuously variable rotation from the belt type continuously variable transmission (CVT) 11 is transmitted to the sun gear 19s, and the rotations of the carrier and the sun gear are combined and transmitted to the output shaft 5 via the ring gear 19r. At this time, since the ring gear 19r, which is a rotating element other than the reaction force supporting element, is connected to the output shaft 5, the planetary gear 19 causes torque circulation, and the sun gear 19s and the carrier 19c rotate in the same direction. , Output shaft 5 rotates forward with zero rotation in between (Lo)
And in the reverse (Rev) direction. That is, on the basis of the torque circulation, when the output shaft 5 is rotated in the forward (forward) direction, the belt-type continuously variable transmission 11 transmits torque from the secondary pulley 9 to the primary pulley 7 to reverse the output shaft (reverse). In the directional rotation state, torque is transmitted from the primary pulley 7 to the secondary pulley 9.

Then, in the high mode in which the low clutch C _L is disconnected and the high clutch C _H is connected, the transmission to the planetary gear 19 via the power transmission device 26 is cut off, and the planetary gear 19 is disconnected. , And the high clutch C _H is engaged to be in an integrated rotation state. Therefore, the rotation of the input shaft 3 is limited to the belt type continuously variable transmission (CVT) 1
1 and the high clutch C _H to the output shaft 5. That is, the CVT 11 transmits power from the primary pulley 7 to the secondary pulley 9. Furthermore, the output shaft 5
Is transmitted to the differential device 25 via the output gear 21 and the reduction gear unit 23, and is transmitted to the left and right front wheels via the left and right axle shafts 6a and 6b.

As shown in the velocity diagram of FIG. 1 (b), the output torque diagram of FIG. 2, and the output rotational speed diagram of FIG. 3, in the low mode, the belt type continuously variable transmission (hereinafter referred to as CVT) is used. )
When 11 is at the limit (O / D end) in the speed-up direction (position a in FIG. 1), the ring gear 19r is reversely rotated with respect to the rotation of the carrier 19c having a constant rotation speed based on the maximum rotation of the sun gear 19s. Then, the reverse rotation (REV) is transmitted to the output shaft 5. When the CVT 11 shifts in the deceleration (U / D) direction, the reverse rotation speed is reduced, and the rotation speed of the output shaft 5 is reduced at a predetermined pulley ratio determined by the gear ratio of the planetary gear 19 and the power transmission device 26. The neutral position (NEU) becomes zero. Further, when the CVT 11 shifts in the deceleration direction, the ring gear 19r is switched to the forward rotation direction, and the forward rotation, that is, the forward rotation is transmitted to the output shaft 5. At this time, as is apparent from FIG. 2, the torque of the output shaft 5 diverges infinitely in the vicinity of the neutral position NEU.

Next, the CVT 11 is in the deceleration direction (U / D).
At the end, the high clutch C _H is connected and switched to the high mode. In the high mode, the output rotation of the CVT 11 is transmitted to the output shaft 5 as it is.
In the velocity diagram of (b), it becomes parallel lines as shown in b. Then, this time, as the CVT 11 is shifted in the speed increasing (O / D) direction, the rotation of the output shaft 5 is also changed to the speed increasing direction, and the transmission torque is reduced accordingly. In addition, λ in FIG. 1 (b) is the number of teeth Zs of the sun gear and the number of teeth Z of the ring gear.
It is the ratio to r (Zs / Zr).

Next, an embodiment of the hydraulic control mechanism of the continuously variable transmission according to the present invention will be described.

Primary and secondary pulleys 7, 9
As shown in FIG. 4, the movable sheaves 7b and 9b are axially movably supported by the boss portions 7c and 9c of the fixed sheaves 7a and 9a via ball splines 30 and 31, respectively. , 9b are respectively provided with hydraulic servos 32, 33 constituting axial force actuating means for exerting an axial force on the pulleys. Both hydraulic servos 3
2, 33 are partition members 35, 36 and cylinder members 37, 39 fixed to the fixed sheave boss portions 7c, 9c.
And the drum members 40 and 41 and the second piston members 42 and 43 fixed to the rear surfaces of the movable sheaves 7b and 9b, and the partition members 35 and 36 are the drum members 40 and 4.
1 is oil-tightly fitted to the first piston member 42,
43 is a cylinder member 37, 39 and a partition member 35, 3
6 oil-tightly fitted to the first hydraulic chambers 45, 4 respectively.
6 and the second hydraulic chambers 47, 49 have a double piston structure.

The first hydraulic chambers 45 and 46 in the hydraulic servos 32 and 33 are respectively provided with movable sheaves 7
The back surfaces of b and 9b form a piston surface, and the effective pressure receiving areas of the piston surface are equal on the primary side and the secondary side. Further, oil passages 50, 51, 52, 53 communicating with the first hydraulic chambers 45, 46 and the second hydraulic chambers 47, 49 are formed in the primary and secondary fixed sheave boss portions 7c, 9c, respectively. A spring 55 for preloading is contracted in the first hydraulic chamber 45 of the primary hydraulic servo 32.

Further, the hydraulic control mechanism (means) 5 of this embodiment
As shown in FIG. 5, 4 has a primary regulator valve 56, a ratio control valve 57, a downshift relief valve 58, a manual valve (switching valve, selecting means) 59, and a low-high control valve (low-high switching means) 60. The hydraulic pressure from the oil pump 17 is appropriately adjusted and switched, and the first and second hydraulic chambers 45 and 4 of the hydraulic servos 32 and 33 are switched.
6, 47, 49 and low clutch (for hydraulic servo) C
It is supplied to _L and the high clutch (hydraulic servo) C _H.

Next, the operation of the hydraulic control mechanism 54 will be described with reference to FIGS.

As shown in FIG. 8, in the low (L) mode in the D range, a predetermined hydraulic pressure is supplied to the first hydraulic chamber 45 of the primary side hydraulic servo 32, and the first hydraulic chamber 33 of the secondary side hydraulic servo 33 is supplied. A predetermined hydraulic pressure is applied to both the hydraulic chamber 46 and the second hydraulic chamber 49, and the low clutch C _L is actuated by hydraulic pressure to connect. That is, in the low mode, the manual valve 59 is operated to the D range position shown in FIG. 5 to communicate the ports d and e, f and g, h and i, k and l, and the port j to the drain. It communicates with the port Ex. Also, low-high control valve 6
0 is in the low position (L) and connects ports o and p, q and r, s and t, u and v, w and x, and port y.
Are switched and held so as to communicate with the drain port Ex.

Therefore, the primary regulator valve 5
The hydraulic pressure from the output port c of No. 6 is transferred to the low clutch hydraulic servo C _L via the ports d and e of the manual valve 59 (see FIG. 5) located in the D range position and the ports o and p of the low / high control valve 60. is supplied to engage the clutch C _L, also port f and g of the manual valve is supplied to the first hydraulic pressure chamber 45 of the primary hydraulic servo 32 further through the ports q and r low high control valve 60. Further, when the downshift operation is not performed, the downshift relief valve 58 is in the normal position where the ports m and n communicate with each other, and the hydraulic pressure from the output port c is the port m,
It is supplied to the first hydraulic chamber 46 of the secondary hydraulic servo 33 via n, the ports h and i of the manual valve 59, and the ports s and t of the low-high control valve 60.
Then, the hydraulic pressure from the output port c is appropriately adjusted by the ratio control valve 57 so that the hydraulic pressure corresponds to the target pulley ratio, and the hydraulic pressure from the output port z is adjusted to the ports k and l of the manual valve 59. , Through the ports w and x of the low-high control valve 60 to the second hydraulic chamber 49 of the secondary hydraulic servo 33.
In this state, the high clutch hydraulic servo C _H
The low-high control valve 60 is in a released state by communicating with the drain port Ex from the port y, and the second hydraulic chamber 47 of the primary hydraulic servo 32 has the ports u and v of the low-high control valve 60 and the port j of the manual valve 59. Through the drain port Ex.

As a result, the low clutch C _L is connected and the CVT 11 is connected to the first and second hydraulic chambers 46, 4
Secondary side hydraulic servo 33 in which hydraulic pressure acts on both 9
Is greater than the axial force by the primary side hydraulic servo 32 in which the hydraulic pressure acts only on the first hydraulic chamber 45,
By appropriately adjusting the opening / closing of the ratio control valve 57 in the axial force state of both pulleys 9 and 7 corresponding to the torque transmission from the secondary pulley 9 to the primary pulley 7, the second hydraulic chamber 49 of the secondary hydraulic servo 33 is controlled. The hydraulic pressure is adjusted, the axial force by the hydraulic servo 33 is appropriately adjusted, and the pulley ratio (torque ratio) is appropriately changed. In this state, the carrier 19c of the planetary gear 19 is transmitted from the input shaft 3 via the low clutch C _L and the power transmission device 26.
The engine torque transmitted to the output shaft 5 is taken out from the output shaft 5 via the ring gear 19r while being regulated by the CVT 11 having the predetermined pulley ratio via the sun gear 19s.

In the high (H) mode of the D range, a predetermined hydraulic pressure is supplied to the first and second hydraulic pressure chambers 45 and 47 of the primary side hydraulic servo 32 as shown in FIG. The high-clutch hydraulic servo C _H is supplied to the first hydraulic chamber 46 of the secondary-side hydraulic servo 33 and is also for the high clutch.
Is supplied to. That is, in the high mode (H),
As shown in FIG. 6, although the manual valve 59 is in the same D range position as in the previous low mode, the low / high control valve 60 is switched to the high position (H),
The ports o and y, q and t, s and r, x and v, w and u are in communication with each other, and the port p is in communication with the drain port Ex.

Therefore, the primary regulator valve 5
The output hydraulic pressure from 6 is the port d of the manual valve 59.
And e, and is supplied to the high clutch hydraulic servo C _H via the ports o and y of the low / high control valve 60 to engage the clutch C _H , and the manual valve 5
9 ports f and g, low-high control valve 60
Is supplied to the first hydraulic chamber 46 of the secondary hydraulic servo 33 via the ports q and t. Further, the ports m and n of the downshift relief valve 58, the ports h and i of the manual valve 59, the low-high control valve 6
Primary hydraulic servo 32 via ports 0 and s and r
Of the primary hydraulic servo 32 through the output port z of the ratio control valve 57, the ports k and l of the manual valve 59, and the ports w and u of the low / high control valve 60, which are supplied to the first hydraulic chamber 45 of the low pressure control valve 60 and are appropriately adjusted. Is supplied to the second hydraulic chamber 47. In this state, the low clutch hydraulic servo C _L is operated from the port p of the low / high control valve 60 to the drain port Ex.
The second hydraulic chamber 49 of the secondary hydraulic servo 33 communicates with the drain port Ex via the ports x and v of the low / high control valve 60 and the port j of the manual valve 59.

As a result, the high clutch C _H is connected and the CVT 11 is connected to the first and second hydraulic chambers 45, 4
Primary side hydraulic servo 32 to which hydraulic pressure is supplied to 7
Is increased by the axial force of the secondary side hydraulic servo 33 that is supplied only to the first hydraulic chamber 46, and in the axial force state corresponding to the torque transmission from the primary pulley 7 to the secondary pulley 9, the ratio is increased. By appropriately adjusting the control valve 57, the hydraulic pressure of the second hydraulic chamber 47 of the primary hydraulic servo 32 is adjusted,
The axial force by the hydraulic servo 32 is adjusted to obtain an appropriate pulley ratio (torque ratio). In this state, the torque transmitted from the engine to the input shaft 3 is appropriately changed by the CVT 11 transmitted from the primary pulley 7 to the secondary 9, and is further extracted from the output shaft 5 via the high clutch C _H.

In the reverse range (R), as shown in FIG. 8, the predetermined hydraulic pressure is the primary hydraulic servo 32.
Are supplied to the first and second hydraulic chambers 45 and 47 of the secondary side hydraulic servo 33, are supplied to the first hydraulic chamber 46 of the secondary hydraulic servo 33, and are supplied to the low clutch hydraulic servo C _L. That is, in the reverse range, the manual valve 59 is in the reverse (R) range position as shown in FIG.
0 is in the low position (L). In this state, the manual valve 59 has ports d and e, f and i, h and g, k and j.
Communicate with each other, and port l is drain port Ex
Communicate with Also, low-high control valve 60
Is similar to the low (L) mode described above, the ports o and p, q
, R, s, t, v and u, x and w, and the port y communicates with the drain port Ex.

Therefore, the primary regulator valve 5
The hydraulic pressure from the output port c of 6 is the manual valve 59.
Port d and e, hydraulic row through the ports o and p of low high control valve 60 clutch servo C _L for
And the ports f and i of the manual valve 59 and the ports s and t of the low-high control valve 60.
Is supplied to the first hydraulic chamber 46 of the secondary hydraulic servo 33 via. Further, ports m and n of the downshift relief valve 58, ports h and g of the manual valve 59, and ports q and r of the low / high control valve 60.
Is supplied to the first hydraulic chamber 45 of the primary hydraulic servo 32 via the pressure control valve 57, and the pressure is appropriately adjusted by the ratio control valve 57.
9 ports k and j, low-high control valve 60
Is supplied to the second hydraulic chamber 47 of the primary hydraulic servo 32 via the ports v and u.

As a result, the low clutch C _L is connected and the CVT 11 is connected to the first and second hydraulic chambers 45, 4
The axial force by the primary side hydraulic servo 32 in which the hydraulic pressure acts on 7 becomes higher than that by the secondary side hydraulic chamber 33 by only the first hydraulic chamber 46, and corresponds to torque transmission from the primary pulley 7 to the secondary pulley 9. The hydraulic pressure of the second hydraulic chamber 47 of the primary hydraulic servo 32 is adjusted by adjusting the ratio control valve 57, and the axial force of the hydraulic servo 32 is adjusted to obtain an appropriate pulley ratio. . In this state, CV
When the pulley ratio of T11 is in the predetermined acceleration (O / D) state,
The engine torque from the input shaft 3 is transmitted to the carrier 19c of the planetary gear 19 via the low clutch C _L and the power transmission device 26, and at the same time, the sun gear 19s is transmitted via the CVT 11 in which the torque is transmitted from the primary pulley 7 to the secondary pulley 9. Is transmitted to the output shaft 5 through the ring gear 19r as reverse rotation.

Further, as shown in FIG. 8, at the parking position P and the neutral position N of the manual valve 59, both the low clutch C _L and the high clutch C _H are released, and the primary side and the secondary side are released. The first hydraulic chamber 45 of the hydraulic servos 32, 33,
A predetermined hydraulic pressure is supplied to 46. That is, in the manual valve 59, the ports f and g, and h and i communicate with each other, and the ports e, j, and l respectively communicate with the drain port Ex.
The low / high control valve 60 is held at the low position L described above.

Therefore, the primary regulator valve 5
The output hydraulic pressure from 6 is the port f of the manual valve 59.
And g, are supplied to the first hydraulic chamber 45 of the primary hydraulic servo 32 through the ports q and r of the low-high control valve 60, and the ports m and n of the downshift relief valve 58, the port h and n of the manual valve 59, respectively. i, low-high control valve 60 port s
And t to the first hydraulic chamber 46 of the secondary hydraulic servo 33. The hydraulic pressure of the high clutch hydraulic servo C _H is released via the port y and the drain port Ex of the low / high control valve 60, and the hydraulic pressure of the low clutch hydraulic servo C _L is released from the ports p and o of the low / high control valve 60. , Manual valve 59
Is released via the port e and the drain port Ex of the second hydraulic chamber 47 of the primary hydraulic servo 32.
The hydraulic pressure of the port u of the low-high control valve 60
And v, the port j of the manual valve 59 and the drain port Ex are released, and the secondary hydraulic servo 3
The hydraulic pressure in the second hydraulic chamber 49 of No. 3 is released through the ports x and w of the low / high control valve 60, the port l of the manual valve 59, and the drain port Ex.

As a result, both the low clutch C _L and the high clutch C _H are released, and the primary hydraulic servo 32 and the secondary hydraulic servo 33 have the same hydraulic pressure acting only on the first hydraulic chambers 45 and 46. Then, substantially the same axial force acts on both the primary and secondary pulleys 7 and 9.

In the positions D; N, R, low (L) mode, and high (H) mode described above, the primary hydraulic chambers 45 and 46 of both the primary and secondary hydraulic servos 32 and 33 are respectively placed in the primary hydraulic chambers 45 and 46. A predetermined hydraulic pressure is supplied from the regulator valve 56, whereby a predetermined axial force corresponding to the transmission torque is secured so that the belt does not slip, and the second hydraulic chamber 47 or 49 of either one of the hydraulic servos 32 and 33 is secured. Then, the pressure control from the ratio control valve 57 acts to adjust the ratio of the axial forces of the two pulleys 7 and 9 to perform the shift operation so that the predetermined pulley ratio is achieved.

When downshifting in the coast state, the downshift relief valve 58 is switched so that the port n communicates with the drain port Ex. As a result, the hydraulic pressure of the first hydraulic chamber 45 or 46 in the supply state in both of the hydraulic servos 32 and 33 described above causes the predetermined port, the port n of the relief valve 58 and the drain port E.
The CVT 11 is discharged through x, and the hydraulic pressure of the hydraulic servo acting on the high hydraulic pressure becomes lower than the hydraulic pressure of the hydraulic servo acting on the low hydraulic pressure.

Next, a control mechanism according to another embodiment will be described with reference to FIGS. 9 to 13.

The primary side and secondary side hydraulic servos 32 and 33 of this embodiment shown in FIG. 9 are similar to those shown in FIG. 4, and the first and second hydraulic chambers 73 ₁ and 75, respectively.
₁ , 73 ₂ , 75 ₂ , but holes 70, 70 are formed in the partition parts 35, 36 to simply increase the pressure receiving area, and the hydraulic pressure is supplied from the oil passages 71, 72, respectively. The hydraulic actuators 73 and 75 of both hydraulic servos 32 and 33 have the same pressure receiving area. In addition, in the first hydraulic chamber 75 ₁ of the secondary side hydraulic servo 33, the spring 5 for preload is used.
5 is reduced. As shown in FIGS. 10 to 12, the control mechanism 54 'constitutes a primary (first) regulator valve 56 and a secondary (second) regulator valve that are continuously connected in series from the oil pump 17. 1 has a ratio control valve 57 and a downshift relief valve 58, and also has a manual valve 59 'and a low / high control valve 60' having an oil passage structure different from that of the above embodiment.
It operates as shown in the operation table of No. 3.

In the low mode (L) in the D range, as shown in FIG. 13, a relatively low line pressure (PL- _L ) from the ratio control valve 57 is supplied to the hydraulic chamber 73 of the primary hydraulic servo 32. and secondary relatively high line pressure from the primary regulator valve 56 to the hydraulic chamber 75 of the hydraulic servo 33 (P _L -H) is supplied, and the hydraulic servo C _L to the line pressure is supplied for the low clutch. That is, the manual valve 59 'is D
In the range position, as shown in FIG. 10, ports d and e, f and i, h and g communicate with each other, and the low-high control valve 60 'is in the low (L) position, and ports o and p are connected. , Q and t, s and r communicate with each other, and the port y communicates with the drain port Ex.

Therefore, the primary regulator valve 5
Relatively high line pressure regulated by 6 (P _L -H) is
It is supplied to the low clutch hydraulic servo C _L via the ports d and e of the manual valve 59 ′ and the ports o and p of the low / high control valve 60 ′, and also the ports m and n of the downshift relief valve 58 and the manual valve 59 ′. Are supplied to the secondary hydraulic servo 33 through the ports f and i of the low-high control valve 60 'and the ports s and r of the low-high control valve 60'. Further, the output pressure from the primary regulator valve 56 is further adjusted appropriately by a ratio control valve (secondary regulator valve) 57, and the adjusted relatively low line pressure (P _L -L).
Is supplied from its output port z to the primary hydraulic servo 32 via the ports h and g of the manual valve 59 'and the ports q and t of the low-high control valve 60'. The hydraulic pressure of the high clutch hydraulic servo C _H is released through the port y and the drain port Ex.

[0066] Thus, the low clutch C _L is connected, CVT 11 is axial force of the secondary pulley 9 by the secondary hydraulic servo 33 of relatively high line pressure P _L -H acts is relatively low line pressure P _L -Primary pulley 7 by primary hydraulic servo 32 on which L acts
It is higher than the axial force of. In this state, the secondary pulley 9
It corresponds to the low (L) mode in the D range in which torque is transmitted from the primary pulley 7 to the primary pulley 7, and the axial force of the primary pulley 7 is adjusted by appropriately adjusting the ratio control valve 57 to adjust the pulley ratio ( The torque ratio) is changed appropriately.

Further, in the high (H) mode of the D range, as shown in FIG. 13, a relatively high line pressure P _L -H is supplied to the primary hydraulic servo 32 and the secondary hydraulic servo 33 relatively. The low line pressure P _L -L is supplied, and the line pressure is supplied to the high clutch hydraulic servo C _H. That is, the manual valve 59 'is held in the D range position as described above.
As shown in, the low-high control valve 60 'is
Switched to high (H) position, ports o and y, q and r,
s and t communicate with each other, and the port p communicates with the drain port Ex.

Therefore, the primary regulator valve 5
A relatively high pressure (line pressure) P _L -H from 6 is supplied to the high clutch hydraulic servo C _H via ports d and e of the manual valve 59 ′ and ports o and y of the low / high control valve 60 ′. The pressure is supplied to the primary hydraulic servo 32 through the ports m and n of the downshift relief valve 58, the ports f and i of the manual valve 59 ', and the ports s and t of the low / high control valve 60'. On the other hand, a relatively low pressure (line pressure) P _L -L from the ratio control valve 57 that further regulates the output pressure of the primary regulator valve 56.
Is supplied from the output port z to the secondary hydraulic servo 33 via the ports h and g of the manual valve 59 'and the ports q and r of the low-high control valve 60'. Incidentally, the hydraulic servo C _L the low clutch hydraulic is
It is released through port p and drain port Ex.

As a result, the high clutch C _H is engaged, and the CVT 11 has a relatively low line pressure P _L when the axial force of the primary pulley 7 by the primary hydraulic servo 32 on which the relatively high line pressure P _L -H acts. -Secondary pulley 9 by secondary hydraulic servo 33 on which L acts
It is higher than the axial force of. In this state, the primary pulley 7
High (H) that transmits torque from the secondary pulley 9 to the secondary pulley 9.
The secondary pulley 9 is compatible with the mode and the ratio control valve 57 is appropriately pressure-controlled.
The axial force of is adjusted and the pulley ratio (torque ratio) is changed appropriately.

In the case of the reverse range (R), as shown in FIG.
As shown in FIG. 3, a relatively high line pressure P _L -H acts on the primary hydraulic servo 32, a relatively low line pressure P _L -L acts on the secondary hydraulic servo 33, and the low clutch hydraulic servo C is operated. Line pressure acts on _L. That is, as shown in FIG. 12, the manual valve 5
9'is switched to the reverse (R) range position, the ports d and e, f and g, and h and i are communicated, and the low / high control valve 60 'is set to the above-mentioned low (L).
Held in position.

Therefore, the primary regulator valve 5
The relatively high pressure (line pressure) P _L -H from 6 is supplied to the low clutch hydraulic servo C _L via the ports d and e of the manual valve 59 ′ and the ports o and p of the low / high control valve 60 ′. At the same time, it is supplied to the primary hydraulic servo 32 via the ports m and n of the downshift relief valve 58, the ports f and g of the manual valve 59 ', and the ports q and t of the low-high control valve 60'. On the other hand, the relatively low pressure regulation (line pressure) P _L -L from the ratio control valve 57 is
It is supplied to the secondary hydraulic servo 33 via the ports h and i of the manual valve 59 'and the ports s and r of the low-high control valve 60'. The hydraulic pressure of the high clutch hydraulic servo C _H is discharged from the port y to the drain port Ex.

As a result, the low clutch C _L is connected, and in the CVT 11, a relatively high hydraulic pressure P _L -H based on the primary regulator valve 56 acts on the primary hydraulic servo 32, and the ratio control valve 5 is used.
A relatively low hydraulic pressure P _L -L based on 7 acts on the secondary hydraulic servo 33. In the reverse (R) range, torque is transmitted from the primary pulley 7 to the secondary pulley 9 in the CVT 11, and the axial force of the primary pulley 7 corresponds to the torque transmission.
The axial force of the secondary pulley 9 is adjusted by adjusting the pressure of the ratio control valve 57, and the pulley ratio is appropriately changed.

In the neutral (N) range and the parking (P) range, as shown in FIG. 13, a relatively low hydraulic pressure P _L -L acts on the primary hydraulic servo 32,
In addition, the secondary hydraulic servo 33 has a relatively low hydraulic pressure P _L −.
L acts. That is, the manual valve 59 'is switched to the neutral or parking position, the port e communicates with the drain port Ex, and the port h
Communicates with the ports g and i, and the low / high control valve 60 'is held at the low (L) position described above.

Therefore, the hydraulic pressure from the output port z of the ratio control valve 57 is supplied from the port h of the manual valve 59 'to the primary hydraulic servo 32 via the port g and the ports q and t of the low / high control valve 60'. At the same time, it is supplied to the secondary hydraulic servo 33 via the port i and the ports s and r of the low-high control valve 60 '. Further, in this state, the hydraulic pressure of the high clutch hydraulic servo C _H is discharged from the port y of the low high control valve 60 ′ to the drain port Ex, and the hydraulic pressure of the low clutch hydraulic servo C _L is
Ports p and o of the low / high control valve 60 ',
It is discharged through the port e and the drain port Ex of the manual valve 59 '.

As a result, both the low and high clutches C
_{When L} and C _H are both cut off, the same hydraulic pressure from the ratio control valve 57 acts on the primary hydraulic servo 32 and the secondary hydraulic servo 33 having the same pressure receiving area, so that both the primary and secondary pulleys 7, 9
Have almost the same axial force.

When the downshift operation is performed, the downshift relief valve 58 is switched so that the port n communicates with the drain port Ex, so that the hydraulic pressure of the hydraulic servo acting on the high hydraulic pressure acts on the low hydraulic pressure. It is lower than the hydraulic pressure of the hydraulic servo.

Next, a partially modified control mechanism will be described with reference to FIG. The control mechanism 54 ₁ includes the primary and secondary hydraulic servos 32 and 33 shown in FIG.
That is, the first and second hydraulic chambers 45, 46, 47, 4 respectively.
It is applied to a so-called double-chamber type having 9. In addition, the primary regulator valve 56,
Ratio control valve 57, the down-shift relief valve 58 and manual valve 59, although those similar to the embodiment shown in FIGS. 5 to 7 is used, low high control valve 60 ₁ is simple four-port two-position switching solenoid valve Consists of. A pilot oil passage 77a is branched into an oil passage 77 from the port y of the low / high control valve 60 ₁ to the high clutch servo C _H , and a switching valve 79 for switching by the pilot pressure is provided.

In the low (L) mode in the D range, the low / high control valve 60 ₁ is set to the low (L) mode.
At the position, the ports o and p communicate with each other, and the port y and the drain port Ex communicate with each other. Also, the oil passage 77 is connected to the drain port E.
In this state, which is in communication with x, the switching valve 7
9 is at the position shown in the figure, and the ports q and r, s and t, u
And v, w and x communicate. In the D range position, the manual valve 59 has the ratio pressure from the ratio control valve 57 introduced to the oil passage and the oil passage communicates with the drain port Ex. Clutch pressure P _C
Is supplied to the low clutch hydraulic servo C _L via the ports o and p of the low / high control valve 60 ₁ .
Also, the first and second line pressures P from the manual valve
_L 1 and P _L 2 are supplied to the first hydraulic chamber 45 of the primary hydraulic servo 32 and the first hydraulic chamber 46 of the secondary hydraulic servo 33 via the ports q and r and the ports s and t of the switching valve 79, respectively. It Further, the ratio pressure from the manual valve is supplied to the second hydraulic chamber 49 of the secondary hydraulic servo 33 via the oil passage and the ports w and x of the switching valve 79, and then to the second hydraulic chamber 47 of the primary hydraulic servo 32. The hydraulic pressure communicates with the ports u and v and the drain port Ex of the manual valve via the oil passage.

In this state, similarly to the above, the low clutch C _L is connected, and the CVT 11 is secondary in proportion to the ratio pressure acting on the second hydraulic chamber 49 of the secondary hydraulic servo 33 in accordance with the torque transmission direction. The axial force of the pulley 9 is higher than that of the primary pulley 7.

In the high (H) mode in the D range, the low / high control valve 60 ₁ is switched to the high (H) position so that the ports o and y communicate with each other and the port p communicates with the drain port Ex. In this state in which the clutch pressure P _C is supplied to the oil passage 77 via the ports o and y, the switching valve 79 is switched by the pilot pressure from the oil passage 77a, and the ports q and t, s and r, x and v communicate with w and u.

Therefore, the clutch pressure P _C is supplied to the high clutch hydraulic servo C _H via the ports o and y of the low / high control valve 60 ₁ and the line pressure P _L.
1 and P _L 2 are connected to the first hydraulic chamber 4 of the secondary hydraulic servo 33 via the ports q and t of the switching valve 79.
6 as well as to the first hydraulic chamber 45 of the primary hydraulic servo 32 via the ports s and r. Further, the ratio pressure from the manual valve 59 is supplied to the second hydraulic chamber 47 of the primary hydraulic servo 32 via the oil passage and the ports w and u, and the hydraulic pressure of the second hydraulic chamber 49 of the secondary hydraulic servo 33 is equal to the hydraulic pressure. , Through the ports x and v and the oil passage, and is discharged from the drain port of the manual valve 59.

In this state, the high clutch C _H is connected, and the CVT 11 causes the axial force of the primary pulley 7 by the primary hydraulic servo 32, which applies the ratio pressure to the second hydraulic chamber 47, in accordance with the torque transmission direction. It is higher than the pulley 9.

In the reverse (R) range,
As shown in FIG. 7, the manual valve 59 is switched,
And the low-high control valve 60 ₁ is low (L)
Held in position. Therefore, in this state, the switching valve 79 is held at the lower position in the figure, as described above. Also, by switching the manual valve 59,
The ratio pressure is supplied to the oil passage, and the oil passage communicates with the drain port Ex.

Therefore, the clutch pressure P _C is supplied to the low clutch hydraulic servo C _L via the ports o and p,
The first and second line pressures are supplied to the first hydraulic chamber 45 of the primary hydraulic servo via ports q and r, and to the first hydraulic chamber 46 of the secondary hydraulic servo via ports s and t. Supplied. Further, the ratio pressure through the manual valve 59 is the oil passage, the ports v and u.
The hydraulic pressure in the second hydraulic chamber 47 of the primary hydraulic servo is discharged from the drain port Ex of the manual valve 59 via the ports x and w and the oil passage. It

In this state, the low clutch C _L is connected, and the CVT 11 causes the axial force of the primary pulley 7 by the primary hydraulic servo 32, which applies the ratio pressure to the second hydraulic chamber 47, in accordance with the torque transmission direction. Higher than pulleys.

Next, a partially modified control mechanism will be described with reference to FIG. This control mechanism 54 ₂ has the primary and secondary hydraulic servos 32, 33 shown in FIG. 9, that is, one hydraulic chamber 73, 75 having the same pressure receiving area, and supplies different hydraulic pressures to the so-called double regulators. Applies to types. The primary regulator valve 56, the ratio control valve (secondary regulator valve) 57, the downshift relief valve 58 and the manual valve 59 'are shown in FIG.
Or similar to that shown in FIG. 12 is used,
The low-high control valve 60 ₁ is composed of a simple 4-port 2-position switching solenoid valve, and is provided with a switching valve 80 capable of switching the hydraulic pressure from the port y of the valve to the high clutch hydraulic servo C _H using the pilot pressure. .

In the low (L) mode in the D range, the low / high control valve 60 ₁ is set to low (L).
At the position, the ports o and p communicate with each other, and the port y and the drain port Ex communicate with each other. Also, the oil passage 77 is connected to the drain port E.
In this state, which is in communication with x, the switching valve 8
0 is at the position shown in the figure, and ports q and t and s and r communicate with each other. At this time, a relatively high line pressure P _L -H from the primary regulator valve 56 is guided to the oil passage based on the D range position of the manual valve 59 ′, and a relatively low line from the ratio control valve 57. The pressure PL- _L is introduced into the oil passage.

Therefore, the clutch pressure P _C from the manual valve 59 is supplied to the low clutch hydraulic servo C _L via the ports o and p of the low / high control valve 60 ₁ , and the relatively low line pressure from the oil passage. P _L −
L is supplied to the primary hydraulic servo 32 via the ports q and t of the switching valve 80, and the relatively high line pressure P _L -H from the oil passage is supplied to the secondary hydraulic servo 33 via the ports s and r. Supplied.

In this state, similarly to the above, the low clutch C _L is connected, and the CVT 11 adjusts the axial force of the secondary pulley 9 by the secondary hydraulic servo 33 on which a higher line pressure acts in accordance with the torque transmission direction. Is higher than that of the primary pulley 7.

In the high (H) mode in the D range, the low-high control valve 60 ₁ is switched to the high (H) position so that the ports o and y communicate with each other and the port p communicates with the drain port Ex. In this state in which the clutch pressure P _C is being supplied to the oil passage 77 via the ports o and y, the switching valve 80 is switched by the pilot pressure from the oil passage 77a, and the ports q and r, s and t are switched. Communicate.

Therefore, the clutch pressure P _C is supplied to the high clutch hydraulic servo C _H , and the relatively low line pressure P _L -L from the oil passage is secondary through the ports q and r of the switching valve 80. It is supplied to the hydraulic servo 33, and a relatively high line pressure P _L -H from the oil passage
Is the primary hydraulic servo 32 via ports s and t.
Is supplied to.

In this state, the high clutch C _H is engaged, and the CVT 11 causes the axial force of the primary pulley 7 by the primary hydraulic servo 32, which acts on the higher line pressure in accordance with the torque transmission direction, to be greater than that of the secondary pulley 9. Get higher

In the reverse (R) range,
As shown in FIG. 12, the manual valve 59 'is switched, and the low / high control valve 60 ₁ is held in the low (L) position. Therefore, in this state, the switching valve 80 is held in the lower position in the figure, as described above. Further, by switching the manual valve 59 ', the oil passage, which leads to relatively high line pressure P _L -H from the primary regulator valve 56, also the oil passage, a relatively low line pressure from the ratio control valve 57 P _L −L is derived.

Therefore, the clutch pressure P _C is supplied to the low clutch hydraulic servo C _L via the ports o and p,
The high line-pressure P _L -H from the oil passage, the port q
And is supplied to the primary hydraulic servo 32 via t,
And the low line pressure P _L -L from the oil passage is
And r to the secondary hydraulic servo 33.

[0095] In this state, the low clutch C _L is connected, CVT 11 is in accordance with the torque transmission direction, the axial force of the primary pulley 7 by the primary hydraulic servo 32 high eye line pressure P _L -H acts It becomes higher than the secondary pulley 9.

Next, the control of the continuously variable transmission of this embodiment will be described.

FIG. 16 is a block diagram of the electronic control unit (EUC) 90. Reference numeral 91 is a sensor installed in the continuously variable transmission 1 for detecting the rotational speed of the input shaft 2 of the transmission, and 92 is
A sensor that detects the rotation of the secondary pulley 9 of the CVT 11, 93 is a vehicle speed sensor that detects the rotation of the output shaft 5 of the continuously variable transmission, and 94 is a vehicle speed sensor that is also used as a speedometer signal. It is also for backup when the vehicle speed sensor 93 fails. 95 is a shift lever of a continuously variable transmission, that is, a manual valve is P,
A sensor for detecting the position of each of the R, N, and D shift positions, 96 is a potentiometer installed in the engine, and a sensor for detecting the opening degree of the throttle. 97 is a throttle opening sensor 96. It is a sensor that is installed in and detects that the accelerator is fully closed. Then, the signals from the respective sensors are taken into the CPU, ROM or RAM via the input processing circuit and the input interface circuit, respectively.

Reference numeral 101 denotes a solenoid for the low / high control valve 60 for switching between the low (L) mode and the high (H) mode, which is turned on and off. Reference numeral 102 denotes a solenoid for the downshift relief valve 58 for draining the high-voltage side circuit, which is a duty or linear solenoid. Reference numeral 103 is a solenoid for the ratio control valve 57 for adjusting the shift control hydraulic pressure, and is composed of a duty or linear solenoid. Reference numeral 105 denotes a solenoid for the primary regulator valve 56 for controlling the line pressure, which is a linear solenoid. Then, each of the solenoids is driven through a solenoid drive circuit 106 that generates a predetermined voltage or output based on a signal from the output interface circuit, and the operation of each solenoid is checked by a monitor circuit 107 and a failure occurs. Is determined and self-determination is performed.

Reference numeral 109 denotes an electronic control unit for controlling the engine, and reference numeral 110 temporarily retards the torque generated by the engine to retard the ignition timing, cut the fuel, and so on to alleviate the shock at the time of shifting. And 111 is a processing circuit for inputting the engine speed. Reference numeral 112 is a checker member including an indicator lamp and the like, which outputs a self-diagnosis result when the electronic control unit 90 fails, and 113 is a circuit for outputting a self-diagnosis result when the failure occurs.
Reference numeral 115 denotes a display device for displaying the state of the continuously variable transmission such as a low (L) mode and a high (H) mode display lamp,
And 116 is a drive circuit therefor.

FIG. 17 is a diagram showing a general flow of the continuously variable transmission, S1 is a step of performing all initial settings at the start of calculation, and S2 is a step from the rotation sensors 91, 92, 93. This is a step of calculating the rotation speed based on the signal. In S3, the signal processing of the neutral start switch is performed, the shift position is read, and in S4, the throttle opening is calculated. And S
In step 5, it is determined whether or not the manual valve is in the reverse (R) range based on the above step S3.
In the case of S, R range control described later is executed (S
6). In the case of NO, it is determined whether or not the D range is selected (S7). In the case of the D range, it is further determined whether or not the current vehicle speed V is equal to or lower than a predetermined vehicle speed V ₁ (for example, 5 km / h) (S8).
When the vehicle speed is equal to or lower than the predetermined vehicle speed, the start control is executed (S
9) Further, when the vehicle speed is equal to or higher than the predetermined vehicle speed, the line pressure control is executed, and the line pressure is set based on the low (L) mode or the high (H) mode, or the throttle opening and the pulley ratio (S10). Further, in S11, the target engine speed is read from the map and determined from the preset best fuel consumption curve based on the throttle opening. Then, compare the current engine speed with the target engine speed,
It is determined whether it is a downshift, an upshift, or the current state is maintained (S12), and a downshift control (S13) or an upshift control (S14) described later is executed.

FIG. 18 is a diagram showing a subroutine of the reverse (R) range control. First, at S15, it is judged from the pulley ratio and the vehicle speed whether or not the vehicle is moving forward. And
When moving forward, the current vehicle speed V is the predetermined vehicle speed V ₂ (for example, 5K
m / h) is determined (S16), and if it is faster than a predetermined vehicle speed, for example, low clutch C _L and high clutch C
Reverse control is executed by releasing both of _{H and} the like (S17), and if slow, neutral (N) control is executed (S18).

The N control is performed so that the axial forces of the primary pulley 7 and the secondary pulley 9 are substantially equalized, or at least the difference between the axial forces of the primary and secondary pulleys is determined when the output torque direction is positive. C at that time
A value smaller than the difference between the axial forces of the two pulleys, which is determined from the input torque of the VT and the pulley ratio, within the range in which the magnitude relationship is not reversed, or the input torque of the CVT at that time when the output torque direction is negative. Also, the difference between the axial forces of the primary and secondary pulleys determined from the pulley ratio is controlled to be a small value within the range in which the magnitude relationship is not reversed. Specifically, in the case of the double chamber type control mechanism 54, 54 ₁ shown in FIGS. 4 to 7 and 14, both the primary and secondary hydraulic servos 32, 3 are provided.
In a state in which the hydraulic pressure is supplied to both the first hydraulic chambers 45 and 46 in 3, the hydraulic pressures in the second hydraulic chambers 47 and 49 are released to equalize the axial forces of the pulleys 7 and 9. Also, FIGS.
And double regulator type control mechanism 54 'shown in FIG. 15, the case of 54 _2, by fully opening the ratio control valve (secondary regulator valve) 57, and supplies the same oil pressure to the primary and secondary both hydraulic servos 32, 33, The axial forces of both pulleys 7 and 9 are made equal.
As a result, regardless of the number of rotations of the output shaft, the CVT 11
The output from the output shaft 5 is self-converged so as to be zero, and is stably held at the neutral position where the rotation of the output shaft 5 is zero. Regarding the N control, Japanese Patent Application No. 7-
It is described in detail in the specification and drawings relating to No. 66234, and the contents thereof are applied to this embodiment as they are.

If it is determined in S15 that the vehicle is moving backward, the current vehicle speed V is compared with a predetermined vehicle speed V ₁ (for example, 5 km / h) (S19). If the current vehicle speed V is slower than the predetermined vehicle speed V _{1, the} vehicle starts. The control is executed (S20). If the control is fast, the line pressure control for setting the line pressure based on the throttle opening and the pulley ratio is executed (S21). Further, the target engine speed is determined by the map reading similar to the above (see S10) (S22), and like the above S12, it is determined whether the upshift, the downshift, or the current state is maintained (S23). Downshift control (S24) and upshift control (S25) described below are executed.

In the reverse (R) control, the CVT 11 is at the O / D end, and the accelerating rotation transmitted from the CVT to the sun gear 19 _S and the power transmission device 26.
The constant rotation transmitted from the carrier 19c to the carrier 19c is combined by the planetary gear 19, and the reverse rotation is taken out from the output shaft 5 integrated with the ring gear 19r. At this time, as shown in FIG. CVT11 across the neutral position (output from output shaft 5 is zero)
Output torque reverses, and CVT changes the primary pulley 7
The power is transmitted from the secondary pulley 9 to the secondary pulley 9. During the reverse control, the manual valves 59, 59 'are
As shown in FIGS. 8 and 12, FIG. 13, FIG. 14, and FIG. 15, in the reverse (R) position and at the time of transmitting positive torque (transmitting to the engine and the wheels), the primary hydraulic servo 32 is operated. Hydraulic pressure is secondary hydraulic servo 3
23. Therefore, as shown in FIG. 23, the axial force F _pr of the primary pulley 7 is higher than the axial force F _sr of the secondary pulley 9, which matches the torque transmission direction described above. There is an axial force relationship. Then, in this state, the ratio control valve 57 is appropriately controlled to change the axial forces F _pr and F _{sr of the} pulleys 7 and 9 so that the gears are appropriately shifted.

FIG. 19 is a diagram showing a flow of upshift control in the D range shown in step S14.
At 26, it is determined whether to execute a high (H) mode shift from a low (L) mode described later. The step S26 is not the R range control. Further, the deviation between the current engine speed and the target engine speed and the acceleration of the deviation are calculated (S27), and the shift speed is calculated based on this calculation (S28). As a result, the apply control of the ratio control valve 57 is executed (S29).

FIG. 20 shows Lo → Hi in S26.
This is a subroutine showing the determination, and the rotation sensor 9 of the primary pulley 7 and the secondary pulley 9 is SS30.
The pulley ratio (ratio) is calculated based on the signals of 1,92. Then, it is determined whether or not the pulley ratio is within the allowable range for shifting from the low mode to the high mode (S3).
1) If it is within the range, the low / high control valve 60 is switched to switch from the low clutch C _L to the high clutch C
_The replacement to _H is executed (S32).

FIG. 21 is a diagram showing the flow of the downshift control shown in step S13. In step S33, it is determined whether or not to execute a low (L) mode shift from a high (H) mode, which will be described later. (Not in R range control). Further, the deviation between the current engine speed and the target engine speed and the acceleration of the deviation are calculated (S34), and the shift speed is calculated based on the calculation result (S35).
Further, it is determined whether or not the downshift is a power-off down based on signals from the throttle opening sensor 96 and the idle switch 97 (S36). And power off,
That is, when the accelerator pedal is released, for example, in the coast state, the ratio control valve 57 is controlled and the downshift relief valve 58 is controlled (S37), and the power is turned on, that is, the accelerator pedal is rapidly depressed. When downshifting with acceleration,
The ratio control valve 57 is controlled (S3
8).

That is, as shown in step S37, during coasting, there is a negative torque state in which torque is transmitted from the wheels toward the engine. In the negative torque state, as shown by the chain line in FIG. Axial force F _p ′ (F _pr
′) And the axial force F _s ′ (F _sr ′) of the secondary pulley 9 reverses the axial force F _p , F _s (F _pr , F _sr ) in the positive torque (engine → wheel) state shown by the solid line. To do. On this occasion,
In the double-chamber type control mechanism 54 shown in FIG. 5, by draining the downshift relief valve 58, the first hydraulic chamber 46 of the secondary hydraulic servo 33 is released, whereby the axial force of the primary pulley 7 is increased. F _p 'is the axial force F _s of the secondary pulley 9' becomes higher than, CVT 11 by the ratio control valve 57 in this state is adjusted appropriately is shifting.

In the double regulator type control mechanism 54 'shown in FIG. 10, the downshift relief valve 58 is drained to release the secondary hydraulic servo 33, whereby the axial force F of the primary pulley is increased. _p 'is the axial force F _s of the secondary pulley' higher than.

At the time of the reverse (R) control shown in FIGS. 7 and 11, the oil passages to the primary hydraulic servo 32 and the secondary hydraulic servo 33 are reversed based on the manual valve. The higher hydraulic servo is depressurized and switched so that the axial force F _Sr ′ of the secondary pulley becomes higher than the axial force F _pr ′ of the primary pulley.

In the power-down downshift control in step S38, the ratio control valve 57 in the double-chamber type control mechanism 54 shown in FIG. Is relatively smaller than the hydraulic pressure of the primary hydraulic servo, and in the double regulator type control mechanism 54 'shown in FIG. 10, the ratio control valve 57 is controlled in the pressure increasing direction,
The hydraulic pressure of the secondary hydraulic servo 33 becomes relatively smaller than the hydraulic pressure of the primary hydraulic servo 32. Thus, the axial force F _s of the secondary pulley and the axial force F _s of the primary pulley are
The ratio of _P decreases, and the CVT 11 rapidly shifts in the speed increasing (O / D) direction, and thus in the deceleration (U / D) direction of the entire transmission 1.

FIG. 22 shows Hi in step S33.
→ It is a Lo determination subroutine. At S40, the current pulley ratio (ratio) is based on the signals from the rotation sensors of the primary pulley 7 (= input shaft 3) and the secondary pulley 9.
Is read (S40), and it is determined whether or not the pulley ratio is within the shift allowable range from the high mode to the low mode (S4).
1). If it is within the range, the low-high control valve 60 is switched, and the grip change from the high clutch C _H to the low clutch C _L is executed (S42).

By changing the grip of the low clutch C _L and the high clutch C _H in steps S32 and S42, the speed is changed between the low (L) mode and the high (H) mode. In the low mode, the above-described power transmission device 26 and CV
By combining with the planetary gear 19 of both rotations of T11,
Forward rotation is taken out from the output shaft 5, but at this time, CVT11
Is transmitted from the secondary pulley 9 to the primary pulley 7 by torque circulation. On the other hand, in the high mode, the output rotation of the CVT 11 is transmitted to the output shaft 5 as it is, so that the CVT 11 is transmitted from the primary pulley 7 to the secondary pulley 9.

At this time, FIG. 6, FIG. 11, FIG. 14 and FIG.
As shown in, low-high control valves 60, 6
By switching 0 ', 60 ₁ , the primary hydraulic servo 3
2 and the hydraulic pressure of the secondary hydraulic servo 33 are switched.
Thus, in the low mode, the axial force F _{s of the} secondary pulley 9 is set so as to match the power transmission direction.
Is higher than the axial force F _p of the primary pulley 7, and in the high mode, the axial force of the primary pulley is higher than the axial force of the secondary pulley. The above description is for positive torque transmission, and during negative torque transmission, as described above, the reverse rotation is performed by the downshift relief valve.

Next, a modification of the power transmission mechanism of the continuously variable transmission will be described with reference to FIGS. 24 to 26.

FIG. 24 shows an embodiment in which the dual pinion planetary gear 19 is used, and the ring gear 19r of the planetary gear is the gears 130 and 131 and the low clutch C.
_It is connected to the input shaft 3 via _L (first rotating element), its sun gear 19s is connected to the secondary pulley 9 via gears 132 and 133 (second rotating element), and its carrier 19c is connected to the output shaft 5. (The third rotating element), the ring gear 19r and the sun gear 19s can be connected via the high clutch C _H , and the ring gear 19r and the sun gear 19s rotate in the same direction. The third rotating element may be a sun gear instead of the carrier, and the first and second rotating elements are respectively composed of a carrier, a sun gear and a ring gear which do not serve as the third rotating element. Good.

In this embodiment, as shown in the velocity diagram of FIG. 24 (b), in the low mode (Lo), the input shaft 3
Engine torque from low clutch _CL , gear 13
1,130 through the ring gear 1 of the planetary gear 19
9R and the CVT 11 and the gear 133,
It is transmitted to the sun gear 19s via 132. On this occasion,
As in the previous embodiment, torque circulation is generated by the planetary gear 19, and the CVT 11 has the secondary pulley 9 on the drive side and the primary pulley 7 on the driven side.
Depending on the increasing / decreasing direction of VT, the carrier 19
Output from c to the output shaft 5. Further, when the CVT 11 enters a predetermined acceleration (O / D) state, the carrier 19c is reversely rotated, that is, reverse (Rev) rotated, and the CVT1 is
Reverse the torque transmission direction of 1. In addition, high mode (H
In the case of i), the engine torque from the input shaft 5 is CVT.
11, is transmitted to the shaft 135 via the gears 133, 132, and in the high mode, the planetary gear 19 is in an integrally rotating state due to the connection of the high clutch C _H ,
The CVT shift torque of the shaft 135 remains unchanged from the output shaft 5
Is transmitted to

In FIG. 25, the sun gear 19s of the single pinion planetary gear 19 is connected to the input shaft 3 via the chain 22 and the low clutch C _L (first rotating element), and the ring gear 19r is connected via the gears 132 and 133 to the secondary. It is connected to the pulley 9 (second rotating element), and the carrier 19c is connected to the differential device 25 via the output shaft 5 and gears 21 and 24 (third rotating element). Then, the sun gear 19s and the ring gear 19r rotate in opposite directions. The first rotating element can be a ring gear and the second rotating element can be a sun gear.

In this embodiment, as shown in the velocity diagram of FIG. 25 (b), the engine rotation from the input shaft 3 is transmitted through the low clutch C _L and the chain 22 to the planetary gear 1
9 sun gear 19s and CVT11,
The rotation is transmitted to the ring gear 19r via the gears 133 and 132, and the two rotations are combined to form the carrier 19c.
To the output shaft 5. At this time, similarly, torque circulation is generated, and the torque is changed in the forward and reverse directions and output from the output shaft 5.

FIG. 26 shows a modification of the embodiment shown in FIG. 25 in which a dual pinion planetary gear is used.
That is, the dual sun gear 19s of the pinion planetary gear 19 is connected to the input shaft 3 via the chain 22 and the low clutch C _L (first rotating element), a carrier 19c
Is coupled to the secondary pulley 9 via gears 132 and 133 (second rotating element), and the ring gear 19r is coupled to the differential device 2 via the output shaft 5 and gears 21 and 24.
5 (third rotating element). Then, the sun gear 19s and the carrier 19c rotate in opposite directions. It is also possible to use the first rotating element as a carrier and the second rotating element as a sun gear.

In the present embodiment, as shown in the speed diagram of FIG. 26 (b), the engine rotation from the input shaft 3 is transmitted through the low clutch C _L and the chain 22 to the sun gear 19s of the dual planetary gear 19. CV
It is transmitted to the carrier 19c via T11 and the gears 133 and 132, and the both rotations are combined and taken out to the output shaft 5 from the ring gear 19r. At this time, similarly,
The torque is circulated and is changed in the forward and reverse directions to be output from the output shaft 5.

[Brief description of drawings]

1 shows an embodiment of a power transmission mechanism in a continuously variable transmission according to the present invention, (a) is a skeleton diagram, and (b) is a velocity diagram.

FIG. 2 is a diagram showing a change in output torque with respect to a torque ratio of the belt type continuously variable transmission (CVT).

FIG. 3 is a diagram showing a change in output rotational speed with respect to the CVT torque ratio.

FIG. 4 is a cross-sectional view showing a double chamber type hydraulic servo that applies an axial force to a primary pulley and a secondary pulley.

FIG. 5 is a hydraulic circuit diagram showing the hydraulic control mechanism, showing a low (L) mode state of the D range.

FIG. 6 is a diagram showing a high (H) mode state of the D range.

FIG. 7 is a diagram showing the R range.

FIG. 8 is a diagram showing its operation.

FIG. 9 is a sectional view showing a double regulator type hydraulic servo.

FIG. 10 is a circuit diagram showing the hydraulic control mechanism, showing a low (L) mode state of the D range.

FIG. 11 is a diagram showing a high (H) mode state of the D range.

FIG. 12 is a diagram showing the R range.

FIG. 13 is a diagram showing the operation thereof.

FIG. 14 is a diagram showing a partially modified hydraulic control mechanism applicable to the double-chamber type hydraulic servo.

FIG. 15 is a diagram showing a partially modified hydraulic control mechanism applicable to the double regulator type hydraulic servo.

FIG. 16 is a block diagram showing an electric control mechanism according to the present embodiment.

FIG. 17 is a diagram showing the general flow thereof.

FIG. 18 is a diagram showing the reverse (R) range control subroutine.

FIG. 19 is a diagram showing a flow of the upshift control.

FIG. 20 is a diagram showing the L _o → Hi determination subroutine.

FIG. 21 is a diagram showing a flow of the downshift control.

FIG. 22 is a diagram showing the Hi → Lo determination subroutine.

FIG. 23 is a diagram showing the axial force balance between a primary pulley and a secondary pulley.

FIG. 24 shows another embodiment of the power transmission mechanism, (a) is a skeleton diagram, and (b) is a velocity diagram.

FIG. 25 shows another embodiment of the power transmission mechanism, (a) is a skeleton diagram, and (b) is a velocity diagram.

FIG. 26 shows another embodiment of the power transmission mechanism, (a) is a skeleton diagram, and (b) is a velocity diagram.

[Explanation of symbols]

1 continuously variable transmission 2 engine output shaft (crankshaft) 3 input shaft 5 output shaft 7 first (primary) pulley 9 second (secondary) pulley 10 belt 19 planetary gear 19c carrier 19s sun gear 19r ring gear 32 axial force actuating means (primary hydraulic servo) 33 axial force operating means (secondary hydraulic servo) 45, 46 first hydraulic chamber 47, 49 the second hydraulic chamber 54, 54 ', 54 _1, 54 ₂ control unit 56 regulator valve (first Regulator valve) 57 Ratio control valve (second regulator valve) 58 (Downshift) relief valve 59, 59 'Switching means (switching valve, manual valve)

Claims (10)

[Claims] 1. An input shaft that interlocks with an engine output shaft, an output shaft that interlocks with wheels, a first pulley, a second pulley that interlocks with the input shaft, a belt wound around these pulleys, and the above. A belt type continuously variable transmission having an axial force actuating means for exerting an axial force on both pulleys in order to change the pulley ratios of the first and second pulleys; and at least the first, second and third rotating elements. A planetary gear having the first rotary element linked to the input shaft, the second rotary element linked to the second pulley, and the third rotary element linked to the output shaft, and at least a vehicle Selecting means for selecting the forward drive state and the reverse drive state of the, and, based on the switching of the forward drive state and the reverse drive state of the selecting means,
A continuously variable transmission in which the torque transmission direction between the first and second pulleys of the belt type continuously variable transmission is changed and the output torque direction of the output shaft is changed. Based on control means for controlling the axial force actuating means so that the axial force acting on the second pulley produces a difference corresponding to the pulley ratio, and the control means based on the switching between the forward drive state and the reverse drive state by the selecting means. And a switching means for changing the magnitude relationship of the axial forces acting on the first and second pulleys to be reversed. 2. The axial force actuating means has a first hydraulic servo acting on the first pulley and a second hydraulic servo acting on the second pulley, and an hydraulic pressure based on an oil pump. Is supplied to the first and second hydraulic servos to apply an axial force to the first and second pulleys, and the switching unit is configured to apply the hydraulic pressure acting on the first hydraulic servo and the The continuously variable transmission according to claim 1, wherein the continuously variable transmission is a switching valve that switches a hydraulic pressure that acts on a second hydraulic servo. 3. The continuously variable transmission according to claim 2, wherein the switching valve is a manual valve that is interlocked with the selecting means. 4. The first and second hydraulic servos each have a plurality of hydraulic chambers, and the switching valve selectively switches the hydraulic chambers for supplying hydraulic pressure to the first and second hydraulic servos. The continuously variable transmission according to claim 2, wherein the effective pressure receiving area of the second hydraulic servo is reversed. 5. The first and second hydraulic servos each have at least a first hydraulic chamber and a second hydraulic chamber, respectively.
Further, the first hydraulic chambers of both the hydraulic servos have the same effective pressure receiving area, the control means has a regulator valve and a ratio control valve, and the hydraulic pressure from the regulator valve is always the first hydraulic chamber of the both hydraulic servos. And a hydraulic pressure by the ratio control valve is supplied to a second hydraulic chamber of either one of the hydraulic servos, and the switching valve is connected to the ratio control valve and the second hydraulic chamber.
The continuously variable transmission according to claim 4, wherein the communication of the hydraulic chamber is reversed to the other. 6. The continuously variable transmission according to claim 5, wherein the regulator valve and the ratio control valve are arranged between the oil pump and the switching valve. 7. The control means has first and second regulator valves for adjusting the hydraulic pressure based on the oil pump to respective hydraulic pressures, and one of the first and second regulator valves is provided for the control means. One of the first and second hydraulic servos and the other of the regulator valves are respectively in communication with the other of the hydraulic servos, and the switching valve includes the first and second regulator valves and the first and second hydraulic valves. The continuously variable transmission according to claim 2 or 3, wherein communication with the hydraulic servo of 2 is reversed. 8. The continuously variable transmission according to claim 7, wherein the first and second regulator valves are arranged between the oil pump and the switching valve. 9. The hydraulic pressure of a hydraulic servo, on which a high hydraulic pressure is applied, based on the judgment of the coast state of the vehicle,
3. The continuously variable transmission according to claim 2, further comprising a relief valve that makes the hydraulic pressure lower than the hydraulic pressure of the hydraulic servo operating. 10. The input shaft has a torque direction of the torque transmitted from the first pulley in the input shaft and a torque direction of the torque transmitted from the first rotating element in the input shaft. The continuously variable transmission according to claim 1, wherein the continuously variable transmission is connected to the first pulley and the first rotating element so as to be in mutually opposite directions.