JPH0262743B2 - - Google Patents

Description

[Detailed description of the invention]

(Industrial Application Field) The present invention relates to a speed change control device for a V-belt continuously variable transmission for a vehicle. (Prior Art) The V-belt type continuously variable transmission for vehicles uses an endless metal V-belt invented by Van Doorne wrapped around two pulleys. It is characterized in that a large number of V-shaped blocks are arranged without interruption on a steel band, and power is transmitted between the side surfaces of the pulley and the V-shaped blocks and by the pressing force between the V-shaped blocks. In conventional speed change control devices for V-belt type continuously variable transmissions for vehicles, the line pressure for controlling the hydraulic servo of the pulleys is set to increase as the torque ratio or reduction ratio between both pulleys increases. There is. In addition, in the V-belt type continuously variable transmission for vehicles, the V-belt is held between two pulleys,
Power is transmitted by the frictional force between the V-belt and the pulley. Therefore, the belt clamping force generating means (hydraulic servo, etc.) that applies the belt clamping force to clamp the V belt to the pulley is required to ensure the transmission torque capacity required for the V-belt type continuously variable transmission for vehicles. A clamping force greater than the required belt clamping force must be generated. However, if the belt clamping force is too large, the V-belt will be pressed more than necessary, which will not only reduce the durability of the V-belt, but also when a hydraulic servo is used as the belt clamping force generating means. In order to generate a large clamping force on the V-belt, it is necessary to increase the driving torque of the oil pump, which reduces transmission efficiency. Therefore, it is desirable to generate a belt clamping force that does not reduce torque transmission efficiency or reduce the durability of the V-belt. However, even if the input torque is constant, the force transmitted to the belt varies depending on the wrapping radius of the belt, that is, the effective diameter, so the optimal belt clamping force depends on the torque ratio of the V-belt continuously variable transmission for vehicles. It's different. Further, the force transmitted by the V-belt also differs depending on the input torque, that is, the engine output. For this reason, a control device for a V-belt type continuously variable transmission for a vehicle as disclosed in West German Published Patent Application No. 2518496 has been provided. In the control device for the V-belt type continuously variable transmission for vehicles, the belt clamping force generated by the hydraulic servo, which is the belt clamping force generating means, is the belt clamping force adjusting unit that adjusts the hydraulic pressure supplied to the hydraulic servo. The pressure control valve is adjusted according to the torque ratio of the V-belt continuously variable transmission and the engine output. (Problems to be Solved by the Invention) However, originally, in a control device for a V-belt type continuously variable transmission for a vehicle, the belt clamping force (hereinafter referred to as
It is called "required belt clamping force." ) is approximately proportional to the rate of increase in engine output, and while it would be ideal to run with a belt clamping force that matches this required belt clamping force, in the case of the above technology, the engine output When the amount increases, the increase in engine output is directly added to the belt clamping force. Therefore, in the control device for the conventional V-belt type continuously variable transmission for vehicles, the belt clamping force generated by the belt clamping force generating means by the engine output (hereinafter referred to as "generated belt clamping force") and its An error occurs between the required belt clamping force and the required belt clamping force corresponding to the engine output. For example, when the vehicle is running at a low engine output without depressing the accelerator, if the engine output is increased in a full throttle state, the belt clamping force will be uniformly increased regardless of the torque ratio at that time. Therefore, if the torque ratio of the V-belt type continuously variable transmission at that time is small, a belt clamping force greater than the originally required clamping force will be generated. That is, the generated belt clamping force greatly exceeds the required belt clamping force, and the V-belt is pressed against the pulley more than necessary, reducing the durability of the V-belt. The present invention solves the problems of the control device of the conventional V-belt continuously variable transmission for vehicles, and combines the generated belt clamping force generated by the belt clamping force generating means with the engine output and the V-belt continuously variable transmission. A vehicle V-belt that approximates the required belt clamping force that changes depending on the torque ratio of the machine, thereby preventing the V-belt from being pressed against the pulleys more than necessary and improving the durability of the V-belt. The object of the present invention is to provide a control device for a continuously variable transmission. (Means for solving the problem) To achieve this, the present invention changes the interval between the V-shaped grooves of the input pulley and the output pulley, changes the effective diameter of both pulleys, and changes the torque ratio between both pulleys. A control device for a V-belt type continuously variable transmission for a vehicle that changes steplessly a belt clamping force adjusting means for adjusting a belt clamping force generating means for clamping a belt stretched between both pulleys. The belt clamping force adjusting means increases the belt clamping force in accordance with an increase in engine output, and increases the rate of increase in accordance with an increase in the torque ratio. (Operation and Effects of the Invention) According to the present invention, as described above, the interval between the V-shaped grooves of the input-side pulley and the output-side pulley is changed, and the effective diameter of both pulleys is changed, so that the torque ratio between both pulleys is changed. A control device for a vehicle V-belt type continuously variable transmission that changes steplessly a belt clamping force adjusting means for adjusting a belt clamping force generating means for clamping a belt stretched between both pulleys; The clamping force adjusting means increases the belt clamping force in accordance with an increase in engine output, and increases the rate of increase in accordance with an increase in the torque ratio. Therefore, as the engine output increases, the belt clamping force increases, and the rate of increase increases as the torque ratio increases, making it possible to control the belt clamping force very closely to the required belt clamping force. Adjust the pressing force of V
The durability of the belt can be improved. (Example) Hereinafter, an example of the present invention will be described in detail with reference to the drawings. FIG. 1 is a schematic diagram of a V-belt type continuously variable transmission for a vehicle. 100 is the engine, 102 is the carburetor, 20
is a transmission device provided between the engine 100 and the drive side axle, which includes a fluid coupling 21 connected to the output side 101 of the engine, and a V-belt continuously variable transmission 30 connected to the fluid coupling 21. , a continuously variable transmission comprising a forward/reverse switching planetary gear transmission 40 connected to the output shaft 26 of the continuously variable transmission 30, and a reduction gear mechanism 23 connected to the output shaft 47 of the planetary gear transmission 40. It is configured. The fluid coupling 21 is the engine 100
A pump impeller 21 connected to an output shaft 101 of
1 and a turbine runner 212 connected to a fluid coupling output shaft 214. Note that other fluid torque converters or mechanical clutches may be used in place of the fluid coupling 21. The V-belt continuously variable transmission 30 has an input pulley 31 connected to the output shaft 214 of the fluid coupling 21 and is disposed parallel to the input pulley 31. An output pulley 32 connected to the shaft 26 and both pulleys 3
It consists of a V-belt 33 stretched between 1 and 32. The input pulley 31 is connected to a fixed flange 31 connected to the fluid coupling output shaft 214.
1, and a movable flange 312 provided to face the fixed flange 311 and form a V-shaped groove.
The movable flange 312 has a hydraulic servo 31
3 so as to be movable in the axial direction. The output pulley 32 also includes a fixed flange 321 connected to the output shaft 26 of the continuously variable transmission 30,
It has a movable flange 322 provided so as to face the fixed flange 321 and form a V-shaped space, and the movable flange 322 is provided so as to be movable in the axial direction by a hydraulic servo 323. The forward/reverse switching planetary gear transmission 40 includes a sun gear 41, a ring gear 43, these sun gears 41,
It is composed of a double planetary gear 44 that meshes with a ring gear 43, a carrier 46 that rotatably supports the double planetary gear 44, and a sun gear 4.
1 is connected to an output shaft 26 of a continuously variable transmission, and a carrier 46 is connected to an output shaft 47 of a planetary gear transmission 40 for forward/reverse switching. The sun gear 41 and the carrier 46 are removably connected by a multi-disc clutch 45, and the ring gear 43 is connected to the multi-disc brake 42.
It is detachably connected to the case 400 of the transmission. In the planetary gear transmission 40 for forward/reverse switching, when hydraulic pressure is supplied to the hydraulic servo 49, the multi-disc clutch 45 is engaged, and the output shaft 26 of the continuously variable transmission is engaged.
The rotation of the vehicle is directly transmitted to the output shaft 47 of the planetary gear transmission 40 for forward/reverse switching, thereby enabling a forward traveling state. Furthermore, when hydraulic pressure is supplied to the hydraulic servo 48, the multi-disc brake 42 is engaged and the ring gear 43 is fixed, so the output shaft 47 rotates in the opposite direction to the rotation of the output shaft 26 of the continuously variable transmission. to enable reverse driving. Next, the reduction gear mechanism 23 is for compensating for the fact that the speed change range obtained by the V-belt type continuously variable transmission 30 is lower than the speed change range achieved by a normal vehicle transmission. For example, the reduction gear mechanism 23 has a reduction ratio of 1.45. The torque is increased by decelerating the torque. The output shaft of the reduction gear mechanism 23 is connected to the differential gear 22, and performs final reduction at a reduction ratio of 3.727, for example. FIG. 2 shows a hydraulic control circuit for the V-belt continuously variable transmission shown in FIG. The hydraulic control circuit includes a hydraulic power source 50 and a hydraulic adjustment device 6.
The hydraulic power source 50 includes a shift control mechanism 70 that cushions the shock during 0, N-D, and N-R shifts, and a torque ratio control device 80. The hydraulic oil pumped up through the pump is supplied to the regulator valve 61 through the oil passage 11 to which the relief valve 53 is attached. The hydraulic adjustment device 60 includes a manual valve 6 that is manually operated by a shift lever (not shown).
2. A detent valve 64 and a throttle valve 65 output detent pressure and throttle pressure according to the throttle opening θ of the carburetor 102, and the detent valve 64 and the throttle valve 65 work together with the movable flange 322 of the output pulley 32 and apply line pressure to the detent valve 64 according to its changing position. from the torque ratio valve 66 which supplies the hydraulic pressure and discharges the output hydraulic pressure feedback oil passage 9 provided in the throttle valve 65, and the regulator valve 61 which regulates the hydraulic pressure supplied from the hydraulic source 50 and supplies it as line pressure to the hydraulic servo 323. configured. The manual valve 62 corresponds to shift positions P, R, N, D, and L of a shift lever provided at the driver's seat.
As shown in FIG. 3, the spool 621 has P, R,
The oil passages 1, which are set at the N, D, and L positions and are supplied with line pressure as shown in Table 1, are connected to the output oil passages 3 to 3.
Contact 5.

[Table] In Table 1, the ◯ mark indicates the state of communication with the oil passage 1, and the × mark indicates that the oil passages 3 to 5 are in a discharged pressure state. The regulator valve 61 connects the spool 611 to the spool 611 by inputting the detent pressure and the throttle pressure.
Regulator valve plunger 6 controlling 11
12, the gap area communicating with the second output port 614 is adjusted as the spool 611 is displaced, and line pressure is output from the output port 616 to the oil passage 1. The oil discharged from the port 614 flows through the oil passage 12.
The fluid is then supplied to the fluid coupling 21, oil cooler and lubrication points. The detent valve 64 includes a spool 641 that is linked and interlocked with the throttle opening θ of the butterfly valve of the carburetor 102 and moves as shown in FIG.
And when the throttle opening is 0≦θ≦θ ₁ ,
As shown in FIG. 4A, the oil passage 5 is communicated with the oil passage 7 for detent pressure output which communicates with the input port 616' provided in the regulator valve 61, and θ ₁ <θ≦
When it is 100%, the oil passage 6 connecting the detent valve 64 to the torque ratio valve 66 as shown in FIG. 4B.
The oil passage 7 is communicated with. The throttle valve 65 includes a spool 651 that is arranged in series with the spool 614 of the detent valve via a spring 645 and has a spring 652 on its back. and spool 6
41 and the spring 645, the spool 6
51 moves. By this action, the opening area of the port 653 communicating with the oil passage 1 is adjusted, and the throttle pressure is output to the oil passage 8 for throttle pressure output communicating with the input port 618 provided in the regulator valve 61. The spools 651 are branched from the oil passage 8, and the orifices 654,
The output hydraulic pressure is fed back to the land 656 and the land 657 which has a larger pressure receiving area than the land 656 through the output hydraulic pressure feedback oil passages 9 and 10 provided with the output hydraulic pressure and the output hydraulic pressure feedback oil passages 9 and 10 provided with the output hydraulic pressure. The torque ratio valve 66 includes a spool 662 linked to the movable flange 322 of the output pulley 32 via a connecting rod, and the amount of movement L of the movable flange 322 is l ₃ ≦L≦l ₄ (torque ratio T is t ₂ ≧T≧t ₁ ), as shown in FIG. 5A, the spool 662
is located on the left side of the diagram. At this time, the input port 664 connected to the output oil pressure feedback oil passage 9 provided in the throttle valve 65 is closed, and the output oil passage 6 to the detent valve 64 is communicated with the drain port 665 to discharge pressure. Further, when the moving amount L of the movable flange 322 is smaller than the first set value l 3 and l ₂ ≦L<l ₃ (t ₃ ≧T>t ₂ ), the spool 662 moves as shown in FIG _. 5B.
is located in the middle part and communicates with the oil passage 9.
4 and the drain port 666 communicate with each other, and the oil passage 9 is depressurized. Then, the moving amount L of the movable flange 322 is the second
is smaller than the set value l ₂ and 0≦L≦l ₂ (t ₄ ≧T>t ₃ ), the spool 662 is closed as shown in FIG. 5C.
is located on the right side in the figure, a port 663 connected to the oil passage 1 and the oil passage 6 communicate with each other, and line pressure is supplied to the oil passage 6. The shift control mechanism 70 includes a shift control valve 71, an orifice 72 provided in the oil passage 1 that supplies line pressure to the oil chamber 713, and the orifice 72 and the oil chamber 71.
3, and a solenoid valve 7 that is controlled by an electric control circuit to be described later and adjusts the oil pressure in the oil chamber 713.
4, the shift control valve 71 includes a spool 712 having a spring 711 disposed behind it on one end and receiving line pressure from an oil chamber 713 provided at the other end. Then, when the solenoid valve 74 is turned on and the drain port 741 is opened to evacuate the oil chamber 713, the spool 712 of the shift control valve 71 is moved to the left in the figure by the action of the spring 711. and an oil passage 13 communicating with a hydraulic servo 49 that operates a multi-plate clutch 45 of the planetary gear transmission 40;
The oil passages 14 that communicate with the hydraulic servo 48 that operates the multi-disc brake 42 are connected to drain ports 71, respectively.
4 and 715 to release pressure, and multi-disc clutch 4
5 or release the multi-disc brake 42. On the other hand, when the solenoid valve 74 is off, the drain port 741 is closed and the spool 712 is moved to the right in the figure by the line pressure supplied to the oil chamber 713. At that time, the oil passages 3 and 4 communicate with the oil passages 13 and 14, respectively, and the multi-disc brake 4
2 or multiple plate clutch 45 is engaged. Also, oil chambers 717 and 716 feed back the output oil pressure of the oil passages 13 and 14 to the shift control valve 71.
is provided to moderate the rise of the output hydraulic pressure and prevent shock when the multi-disc clutch 45 and the multi-disc brake 42 are engaged. The torque ratio control device 80 includes a torque ratio control valve 81, orifices 82, 83, a downshift solenoid valve 84, and an upshift solenoid valve 85. The torque ratio control valve 81 has a spool 812 with a spring 811 on one side,
Orifices 8 are installed in the oil chambers 815 and 816 at both ends.
Line pressure is supplied from the oil passage 1 via 2 and 83, respectively. It communicates with the oil passage 1 to which line pressure is supplied, and opens into an input port 817 whose opening area increases or decreases according to the movement of the spool 812, an oil chamber 819, and an input port to the V-belt continuously variable transmission 30. Hydraulic servo 313 of side pulley 31
An output port 818 that communicates with the oil passage 2 via an oil passage 2, a drain port 814 that evacuates the oil chamber 819 according to the movement of the spool 812, and a drain port 813 that evacuates the oil chamber 815 according to the movement of the spool 812 are provided. . Downshift solenoid valve 84 and upshift solenoid valve 85
are the oil chamber 81 of the torque ratio control valve 81, respectively.
5,816, and both are operated by the output of the electric control circuit described later, and the oil chambers 815, 81
6 is exhausted. FIG. 6 shows an electric control circuit for controlling the shift control solenoid valve 74 of the shift control mechanism 70, the downshift solenoid valve 84 and the upshift solenoid valve 85 of the torque ratio control device 80 in the hydraulic control circuit shown in FIG. 90. FIG. In the figure, 901 indicates that the shift lever is P, R,
A shift lever switch detects which position of N, D, or L it is shifted to; 902 is a rotational speed sensor that detects the rotational speed of the input pulley 31;
903 is a vehicle speed sensor, 904 is a throttle sensor that detects the throttle opening of the carburetor or the amount of depression of the accelerator pedal, and 905 is a rotation speed sensor 9
906, a vehicle speed detection processing circuit that converts the output of the vehicle speed sensor 903 into voltage; 907, a throttle opening detection processing circuit that converts the output of the throttle sensor 904 into voltage; 908 to 911 are input interfaces for each sensor, 912 is a central processing unit (CPU), and 913 is a solenoid valve 74, 84, 8
A read-only memory (ROM) that stores programs to control the 5 and data necessary for control;
914 is a random access memory (RAM) for temporarily storing input data and parameters necessary for control; 915 is a clock; 916 is an output interface; and 917 is a solenoid output driver. The driver 917 converts the output of the output interface 916 into operating outputs of the upshift solenoid valve 85, the downshift solenoid valve 84, and the shift control solenoid valve 74. Input interface 908-911
A data bus 91 is connected between the CPU 912, ROM 913, RAM 914, and output interface 916.
8 and address bus 919. Next, the torque ratio valve 66, the detent valve 64,
The operation of the hydraulic pressure adjusting device 60 of this embodiment, which is composed of a throttle valve 65, a manual valve 62, and a regulator valve 61, will be explained. The hydraulic oil supplied to the hydraulic control circuit is supplied from a pump 52 driven by the engine, and if the line pressure is high, power consumption by the pump 52 increases accordingly. Therefore, in order to drive the vehicle with low fuel consumption, it is necessary to bring the line pressure supplied to the hydraulic control circuit close to the minimum necessary value. Therefore, in the continuously variable transmission, the line pressure is defined as a hydraulic pressure that allows each hydraulic servo of the input pulley 31 and the output pulley 32 to transmit torque without causing the V-belt 33 to slip. When the engine is operated in a state that provides the best fuel efficiency, the minimum necessary line pressure for changes in the torque ratio T between the input and output shafts is shown by the solid line in FIG. 7 using the throttle opening θ as a parameter. When the vehicle is started, it is impossible to operate the engine at the best fuel efficiency within the range of torque ratios that can be achieved by both pulleys 31 and 32. 20% from the characteristic curve of
It is desirable to set the line pressure to a certain level. Furthermore, during engine braking, it is desirable to have a higher line pressure characteristic even when the throttle opening degree θ=0, as shown by the dashed line. In this embodiment, the hydraulic pressure adjustment device 60 adjusts the line pressure, which is the output of the regulator valve 61, to the shift position (L, D, N, R, P) of the manual valve 62, the throttle opening θ, and the torque ratio between the input and output shafts. It is adjusted as follows depending on the change in . (Position D) As shown in Table 1 above, in the manual valve 62, only the oil passage 3 communicates with the oil passage 1, and the oil passages 4 and 5 are depressurized. At this time, in the shift control mechanism 70, the shift control solenoid valve 74
When line pressure is being supplied to the oil chamber 713 with This acts on the hydraulic servo 49 of the multi-plate clutch 45 for forward movement, and the vehicle becomes ready to move forward. When the torque ratio T is t1 _≦ T≦ _t2 . As shown in FIG. 5A, the torque ratio valve 66 closes the port 663 communicating with the oil passage 1, and communicates the oil passage 6 with the drain port 665 to exhaust pressure. As a result, regardless of the throttle opening degree θ, the oil path 7
No detent pressure occurs. Also, since the port 664 of the torque ratio valve 66 that communicates with the oil passage 9 is closed, the spool 664 of the throttle valve 65
1 receives feedback pressure not only from land 656 but also from land 657. Therefore, the throttle pressure having the characteristic shown in the characteristic curve C in FIG. As a result, the line pressure output from the regulator 61 becomes as shown in area H of FIG. 9 and characteristic curve H of FIG. 10. When the torque ratio T is _t2 <T≦ _t3 . As shown in FIG. 5B, the torque ratio valve 66 closes the port 663, and the oil passage 9 and the drain port 666 communicate with each other. Also, oil passage 6 is port 665
is discharged through. Therefore, detent pressure is not generated, and the throttle pressure increases by the amount that the pressure in the oil passage 9 is exhausted and the feedback pressure is no longer applied to the land 656 of the spool 651, and the throttle pressure has the characteristics shown in the characteristic curve D in FIG. . The line pressure at this time has the characteristics shown by the area R in FIG. 9 and the characteristic curve G in FIG. When the torque ratio T is t ₃ <T≦t ₄ . As shown in FIG. 5C, the oil passage 9 is discharged from the drain port 666, so that the throttle pressure has the characteristic shown in the characteristic curve D of FIG. 8, as described above.
Here, the throttle opening θ is within the range of 0≦θ≦θ ₁ %, and the spool 641 of the detent valve 64 is
As shown in FIG. 4A, while on the left side of the figure, the oil passage 6 is closed by the spool 641, and the oil passage 7 is closed by the spool 641.
is exhausted from the manual valve 62 via the oil passage 5. When the throttle opening θ is θ ₁ %<θ≦100%, the spool 641 moves to the right in the figure as shown in FIG. 4B, and the oil passage 6 and the oil passage 7 communicate with each other. Detent pressure is generated. As a result, the line pressure has a characteristic that changes stepwise at θ=θ ₁ %, as shown in the area of FIG. 9 and the characteristic curve of FIG. 10. (L position) The oil passage 5 communicates with the oil passage 1 at the manual valve 62. Oil passage 3 and oil passage 4 are equivalent to the D position. When the torque ratio T is t1 _≦ T≦ _t2 . When the throttle opening θ is 0≦θ≦θ ₁ %, the oil passage 5 and the oil passage 7 communicate with each other in the detent 64, and detent pressure is generated to push up the throttle plunger and generate high line pressure. When the throttle opening θ is θ ₁ %<θ≦100%, the oil passage 7 is connected to the oil passage 6 and the torque ratio valve 6 as shown in FIG. 4B.
Since the pressure is exhausted through the drain port 665 of No. 6, no detent pressure is generated, and the throttle pressure is the same as in the case of the D position. Therefore, the line pressure has the characteristics shown in the characteristic curve 1 in FIG. When the torque ratio T is _t2 <T≦ _t3 . Unlike the above, in the torque ratio valve 66, the oil passage 9 communicates with the drain port 666 to exhaust pressure, and the throttle pressure output from the throttle valve 65 to the regulator valve 61 via the oil passage 8 increases.
As a result, the line pressure has a characteristic as shown in characteristic curve H in FIG. When the torque ratio T is t ₃ <T≦t ₄ . The oil passage 6 and the oil passage 1 are communicated with each other by the torque ratio valve 66, and the pressure of the oil passage 9 is exhausted from the drain port 666. Since line pressure is supplied to both the oil passage 6 and the oil passage 5, the detent valve 64 outputs the detent pressure regardless of the throttle opening, and the regulator valve 61 inputs the detent pressure and the same throttle pressure as above. outputs the line pressure shown in FIG. (R position) At the manual valve 62, the oil passage 4 and the oil passage 5 communicate with the oil passage 1, and the oil passage 3 is depressurized. At this time, in the shift control mechanism 70, if the shift control solenoid 74 is in the OFF state and line pressure is being supplied to the oil chamber 713, the spool 712 is positioned to the right, and the oil passage 4 and the oil passage 14 are communicated with each other. The line pressure supplied to the oil passage 4 is supplied to the hydraulic servo 48 of the multi-disc brake 42 for reverse movement through the oil passage 14, and the vehicle enters the reverse movement state. (P position and N position) Since the oil passages 3, 4, and 5 are all exhausted in the manual valve 62, the line pressure that is the output of the regulator valve 61 is the same as in the D position. In this line pressure adjustment, the manual valve 62
When the engine is shifted to the D, N, and P shift positions, the line pressure when the torque ratio T is in the range of t ₃ < T ≤ t ₄ is as shown in the characteristic curve R of Fig. 10. The opening degree θ is set low at ₁ % or less. If the line pressure is set high when the throttle opening θ is small and the pump discharge is low, such as when idling, it will be difficult to maintain the line pressure when the oil temperature is high and there is large oil leakage from various parts of the hydraulic circuit. This is because the oil temperature further increases due to a decrease in the amount of oil supplied to the oil cooler, which is likely to cause trouble. Further, when the manual valve 62 is shifted to the L and R shift positions, the torque ratio T is in the range of t ₁ ≦T≦t ₂ and the throttle The line pressure is set high under operating conditions where the opening degree θ is θ ₁ % or less. This is because relatively high oil pressure is required during engine braking even when the throttle opening is low. As shown in FIG. 9, by bringing the line pressure close to the minimum necessary oil pressure shown in FIG. 7, power loss caused by the pump 52 can be reduced, and fuel efficiency can be improved. Next, the operations of the shift control mechanism 70 and the torque ratio control device 80 controlled by the electric control circuit 90 shown in FIG. 6 will be explained with reference to the program flowchart shown in FIG. 12. After the throttle opening θ is read by the throttle sensor 904 (step 921), the shift lever position is determined by the shift lever switch 901 (step 922). As a result of the determination,
When the shift lever is in the P position or N position, the solenoid valves 84 and 85 are turned off by the P position or N position processing subroutine shown in FIG. 13 (step 931), and the P or N state is stored in the RAM 914 (step 932). As a result, a neutral state of the input pulley 31 is obtained. When the shift lever changes from the P position or the N position to the R position, and when it changes from the N position to the D position, a shift shock is activated to relieve the shift shock associated with the N-R shift and N-D shift, respectively. Perform control processing (step 9)
40,950). This shift shock control process will be described in detail below. First, the shift control mechanism 70 adjusts the timing of supplying and discharging hydraulic pressure to the hydraulic servos 48 and 49 of the planetary gear transmission 40 by the action of the solenoid valve 74 that is controlled by the output of the electric control circuit 90 described above. The hydraulic servo 4 is prevented by the action of the pressure limiting valve 73.
It has the function of keeping the upper limit of the oil pressure supplied to 8 and 49 below a set value, and limits the engagement pressure of the clutch and brake. In this embodiment, as shown in FIG.
The pressure-receiving areas of the lands provided on the spool 712 of the shift control valve 71 are S ₁ , S ₁ , and S ₂ in order from the left side in the figure, the elastic force of the spring 711 is F _s1 , and the oil chamber 71 is
3, the hydraulic pressure P _c is supplied to the hydraulic servo 49 of the multi-disc clutch 45 that is engaged when moving forward, and the hydraulic pressure P _b is supplied to the hydraulic servo 48 of the multi-disc brake 42 that is engaged when moving _backward. are given as follows from the equations 1 and 2, which are hydraulic equilibrium equations of the shift control valve 71, respectively. When moving forward P _s・S ₁ = P _c・S ₂ + F _s1 P _c = (S ₁ / S ₂ ) P _s − (F _s1 / S ₂ ) When moving backward P _s・S ₁ = P _b (S ₁ − S ₂ ) + F _s1 P _b = [S ₁ / (S ₁ - S ₂ )] P _s - [F _s1 / (S ₁ - S ₂ )] In addition, the valve inserted in the pressure limiting valve 73 The pressure receiving area of the body 731 is S ₃ , and the valve body 73 is
The elastic force of the spring 732 installed behind 1 is F _s2
Then, the pressure limiting valve 73 operates at the maximum pressure P limit of P _s according to the hydraulically balanced equation. P limit×S ₃ =F _s2 P limit=F _s2 /S ₃ At this time, the maximum pressures P _c limit and P _b limit of P _c and P _b are limited according to the formulas 1 and 2. When moving forward, P _c limit = (S ₁ / S ₂ ) P limit - F _s1 / S ₂ When moving backward, P _b limit = [S ₁ / (S ₁ - S ₂ )] P limit - [F _s1 / (S ₁ - S ₂ )] The solenoid valve 74 generates a solenoid pressure P _s in the oil chamber 713 with a duty (%) given by the following equation. Duty = (solenoid on time in one cycle/solenoid operating cycle) x 100 (%) As shown in Fig. 15, this duty control has a pulse width of L ^* - in one cycle K ^* .
This is done by applying pulses whose nM ^* (n=1, 2, 3, . . . ) gradually decreases to the shift control solenoid valve 74 shown in FIG. By controlling the duty of the shift control solenoid valve 74 in this way, the hydraulic pressure P _s adjusted according to the duty is applied to the shift control valve 71.
occurs in the oil chamber 713. The change in the solenoid pressure P _s shown in FIG. 17 is amplified in the shift control valve 71, and the hydraulic pressures P _c and P _b supplied to the hydraulic servos 48 and 49 shown in FIG. 18 are obtained. When relaxing the engagement shock during N-D shift and N-R shift, hydraulic servos 48 and 49 are used.
The rise of the hydraulic pressures P _b and P _c supplied to the AC is controlled as shown in the hydraulic characteristic curve shown in FIG. 16, and the multi-disc clutch 45 or the multi-disc brake 42 between AC is engaged in the diagram. In this way, the solenoid valve 74
A program flowchart of shift shock control processing 940, 950 for controlling the hydraulic pressure supplied to the hydraulic servos 48, 49 is shown in FIG. That is, FIG. 19 is a program flowchart when control is performed using the parameters K ^* , L ^* , and M ^* of the waveform diagram shown in FIG. 15. In step 941, it is determined whether or not the shock control processing flag is on. If the flag is on, the shift shock control process is in progress, so the process goes to step 946. If the flag is not on, the shift lever position and current shift lever stored in the RAM 914 are used to start the shift shock control process. By comparing the positions, it is determined whether the shift lever has changed from the P position or the N position to the R position (step 942).
Then, it is determined whether there is a change from the N position to the D position (step 943). If any change has occurred, the corresponding parameters K ^* , L ^* , M ^* are set in steps 944 and 945, and the parameter K is set to 0, and the shock control processing is performed. A flag indicating this is turned on (step 955). If no change has occurred, the process returns and no shock control processing is performed. In step 946, it is determined whether the parameter K that determines the end of one period K ^* is greater than 0, and if K is less than 0, K is set to K-1.
Then, L is set to L ^* and L ^* is set to L ^* -M ^* (step 947), and it is determined whether L is less than or equal to 0 in step 948. If L is less than or equal to 0, step 9 is executed.
The process proceeds to step 51, and if L is less than 0, it is assumed that all shock control processing has been completed and the flag is turned off. In step 946, if the parameter K for determining the end of one period K ^* is greater than 0, then K-1
is set to K (step 950), and then one period
It is determined whether the parameter L, which determines the end of the on-time at K ^* , is 0 (step 951).
When L is 0, an OFF command is issued for the solenoid valve 74 (step 952), and when L is not 0, an on command is issued (step 953), and then L-1 is set to L.
and return. Similar shift shock control processing can also be performed using programmable timer 920 (see FIG. 6). Next, referring back to FIG. 12, the transmission control device, which is a feature of the present invention, will be explained. N-D shift shock control processing 950
Next, the input side pulley rotation speed sensor 902
The actual input pulley rotation speed, that is, the input pulley actual rotation speed N is read (step 923).
It is determined whether or not the throttle opening degree θ read in the next step 921 is 0 (step 924).
When θ≠0, a subroutine 960 is executed to set the input pulley target rotation speed N ^* to the input pulley rotation speed with the best fuel efficiency. Furthermore, when the throttle is fully closed at θ=0, in order to determine the necessity of engine braking, it is determined whether the shift lever is set to the D position or the L position (step 926), and the shift lever is When the shift lever is set to the D position, the D position engine brake processing subroutine 970 is executed, and when the shift lever is set to the L position, the L position engine brake processing subroutine 980 is executed, and the input side pulley target Set the rotation speed N ^* to an appropriate value for each. Next, subroutine 9 sets the input pulley target rotation speed N ^* to the input pulley rotation speed for the best fuel efficiency.
60 will be explained. Generally, in order to operate the engine at the best fuel efficiency, it is preferable to operate the engine along the best fuel efficiency power line shown by the broken line in FIG. In FIG. 20, the horizontal axis shows the engine rotation speed (rpm) and the vertical axis shows the torque (Kg·m) of the engine output shaft, and the best fuel efficiency power line is obtained as follows. In other words, from the engine's equal fuel consumption rate curve (in g/ps/h) shown by the solid line in Figure 20 and the equal horsepower curve (in ps) shown by the two-dot chain line, the fuel at point A in the figure can be calculated. Assuming that the consumption rate is Q (g/ps·h) and the horsepower is P (ps), at point A, fuel will be consumed per hour by S=Q×P (g/h). By determining the fuel consumption per hour S at all points on each equal horsepower curve, the point where S is the minimum on each equal horsepower curve can be found, and by connecting these points, the best fuel consumption for each horsepower can be found. The best fuel efficiency power line indicating the engine operating state is obtained. Here, when the engine 100 and the fluid coupling 21, which is a fluid transmission mechanism, are combined as in this embodiment, the engine output performance curve at the throttle opening θ shown in FIG. The best fuel efficiency fluid coupling output line can be found on the fluid coupling output performance curve shown in FIG. 24 from the fluid coupling performance curve shown in FIG. 22 and the engine equivalent fuel consumption rate curve shown in FIG. 23. Figure 25 is the 24th
The optimum fuel efficiency fluid coupling output line shown in the figure is replaced with the relationship between throttle opening and fluid coupling output rotation speed. This fluid coupling output rotation speed directly becomes the input pulley rotation speed N _B in the continuously variable transmission of this embodiment. For this purpose, in a subroutine 960 shown in FIG. 26 that sets the input pulley target rotation speed N ^* to the input pulley rotation speed N _B for the best fuel economy, the throttle opening θ is pre-stored in the ROM 913 as data. The address of the best fuel economy input side pulley rotation speed N _B corresponding to the throttle opening θ is set (step 961), and the best fuel efficiency input side pulley rotation speed N _B is read from the set address (step 962). Best fuel efficiency input pulley rotation speed corresponding to throttle opening θ
The data of N _B is set to the input pulley target rotation speed N ^* (step 963). Next, the engine brake processing subroutines 970 and 980 shown in FIG. 12 will be explained. D position engine brake processing subroutine 9
70 is a vehicle speed sensor 90 as shown in FIG.
3 reads the vehicle speed V (step 971),
At that point, the acceleration α is calculated (step 972),
Next, it is determined whether the acceleration α is an appropriate acceleration A for the vehicle speed (step 973). When the acceleration α is larger than the acceleration A, the input pulley target rotation speed N ^* is set to a value larger than the current input pulley rotation speed N in order to downshift (step 974), and when the acceleration α is smaller than the acceleration A , the target input pulley rotation speed N ^* is set to the best fuel efficiency input pulley rotation speed N _B corresponding to the throttle opening degree θ (step 975), and the process returns. The relationship between vehicle speed and appropriate acceleration A is determined in advance through experiments or calculations for each vehicle, as shown in FIG. Engine brake processing subroutine 9 for L position
70 is a vehicle speed sensor 90 as shown in FIG.
3 reads the vehicle speed V (step 981),
Next, the current torque ratio T is calculated from the vehicle speed V and the input pulley rotation speed N using the following formula (step 98).
2). T=(N/V)×k k is the reduction gear mechanism 2 inside the transmission
3, the final reduction ratio of the vehicle, the tire radius, etc. Next, the current torque ratio T is the torque ratio that provides safe and appropriate engine braking for that vehicle speed.
Determine whether it is smaller than T ^* (step 98).
3) When the torque ratio T is smaller than the torque ratio T ^* , set the input pulley target rotation speed N ^* to a value larger than the current input pulley rotation speed N in order to downshift (step 984), and return. . When the torque ratio T is larger than the torque ratio T ^* , the input pulley target rotation speed N ^* is set to the current input pulley actual rotation speed N (step 985);
Return. The torque ratio T ^* that provides safe and appropriate engine braking for the vehicle speed is determined in advance through experiments or calculations for each vehicle, as shown in FIG. As above, input pulley target rotation speed N ^*
Once set, as shown in FIG. 12, the current input pulley actual rotation speed N is compared with the input pulley target rotation speed N ^* for the best fuel economy (step 9).
27), when N<N ^* , a command is issued to operate the downshift solenoid valve 84 (step 928), and N
>N ^* , a command to operate the upshift solenoid valve 85 is issued (step 929), and when N=N ^* , a command to turn off both solenoid valves 84 and 85 is issued (step 920). The torque ratio control device 80 controls the speed change between the input and output pulleys by comparing the target rotational speed N ^* of the input pulley with the best fuel economy obtained from the subrunan of FIG. 27 and the actual input pulley rotational speed N. The ratio is increased or decreased by operating two solenoid valves 84 and 85 provided in the torque ratio control device 80, so that the actual input pulley rotation speed N matches the input pulley target rotation speed N ^* for best fuel economy. be exposed. (During constant torque ratio running) As shown in FIG. 31A, the solenoid valves 84 and 85 controlled by the output of the electric control circuit are turned off. As a result, the oil pressure P ₁ in the oil chamber 816 is the line pressure, and the oil pressure P ₂ in the oil chamber 815 is also the line pressure when the spool 812 is on the right side in the figure. The spool 812 is moved to the left in the figure because of the pressing force _P3 exerted by the spring 811. When the spool 812 is moved to the left and the oil chamber 815 and the drain port 813 communicate with each other, the pressure P ₂ is exhausted, so the spool 812 is moved to the right in the figure by the oil pressure P ₁ in the oil chamber 816. When spool 812 is moved to the right, drain port 813 is closed. In this case, a flat notch 8 is formed on the land edge of the drain port 813 and the spool 812.
12b makes it possible to maintain the spool 812 more stably at the equilibrium point in the intermediate position of FIG. 31A. In this state, the oil passage 2 is closed,
The hydraulic pressure of the hydraulic servo 313 of the input pulley 31 is
The line pressure applied to the hydraulic servo 323 of the output pulley 32 causes it to be compressed via the V-belt 33, resulting in equilibrium with the hydraulic pressure of the hydraulic servo 323. Actually, since there is oil leakage in the oil passage 2 as well, the input pulley 31 is gradually expanded and the torque ratio T changes in the direction of increasing. Therefore, as shown in FIG. 31A, the spool 8
12 is balanced, the drain port 814 is closed, and a flat notch 812a is provided at the land edge with the spool 812 so that the oil passage 1 is in a slightly open state to compensate for oil leakage in the oil passage 2. There is. (During upshift) As shown in FIG. 31B, the solenoid valve 85 is turned on by the output of the electric control circuit. As a result, the pressure in the oil chamber 816 is evacuated, so the spool 812 is moved to the left in the figure.As the spool 812 moves, the pressure in the oil chamber 815 is also evacuated from the drain port 813, but due to the action of the spring 811, the spool 812
12 is set at the left end in the figure. In this state, the line pressure of oil passage 1 is at port 818.
Since the oil is supplied to the oil path 2 via the hydraulic servo 3
13 increases, the input pulley 31 operates in the closing direction, and the torque ratio T decreases. Therefore, by controlling the ON time of the solenoid valve 85 as necessary, the upshift is performed by reducing the torque ratio by a desired amount. (During downshift) As shown in FIG. 31C, the solenoid valve 84 is turned on by the output of the electric control circuit, and the oil chamber 815 is evacuated. The spool 812 is moved to the right in the figure by the line pressure of the oil chamber 816, the oil passage 2 is communicated with the drain port 814 and the pressure is discharged, and the input pulley 31 is operated in the expanding direction to increase the torque ratio. In this way, by controlling the on-time of the solenoid valve 84, the torque ratio is increased and downshifted. As mentioned above, input side (drive side) pulley 3
The output hydraulic pressure of the torque ratio control valve 81 is supplied to the hydraulic servo 313 of No. 1, and line pressure is led to the hydraulic servo 323 of the output side (driven side) pulley 32, which controls the hydraulic pressure of the input side hydraulic servo 313. When P _i is the hydraulic pressure of the output side hydraulic servo 323, P _p /P _i has a characteristic as _shown in the graph of FIG. 32 with respect to the torque ratio T. For example, when driving with throttle opening θ = 50% and torque ratio T = 1.5 (point a in the figure), loosen the accelerator and θ = 30%.
In this case, when P _p /P _i is maintained as it is, it moves to point b in the figure and the torque ratio T = 0.87,
Conversely, when maintaining the torque ratio T=1.5, the value of P _p /P _i is increased by the output of the torque ratio control mechanism 80 that controls the input pulley and changed to the value at point c in the figure. By controlling the value of P _p /P _i as necessary in this manner, it is possible to set an arbitrary torque ratio corresponding to any load condition. Note that the present invention is not limited to the above-described embodiments, and various modifications can be made based on the spirit of the present invention, and these are not excluded from the scope of the present invention. For example, in the above embodiment, the actual rotational speed signal corresponding to the engine rotational speed and the engine output signal corresponding to the engine load are detected, but instead of detecting the engine output signal corresponding to the engine load, the vehicle speed is detected. A corresponding signal may also be detected. In addition, in the above-mentioned speed change control, the target rotation speed of the input pulley is set to obtain the best fuel efficiency.
The input pulley target rotation speed can be set to obtain the maximum torque.

[Brief explanation of drawings]

Fig. 1 is a schematic diagram of a V-belt continuously variable transmission for vehicles, Fig. 2 is a diagram showing an embodiment of a hydraulic control circuit in the V-belt continuously variable transmission for vehicles of the present invention, and Fig. 3 is a manual diagram. A diagram to explain the operation of the valve,
Fig. 4 is a diagram for explaining the operation of the detent valve and the throttle valve, Fig. 5 is a diagram for explaining the operation of the torque ratio valve, and Fig. 6 is a diagram for explaining the operation of the torque ratio valve. The configuration diagram, FIG. 7 is a diagram showing the required line pressure characteristics of the hydraulic control circuit, FIG. 8 is a diagram showing the throttle pressure characteristics, and FIGS. Figures 12 and 13 are diagrams for explaining the flow of processing in the electric control circuit, Figure 14 is a diagram for explaining the operation of the shift control mechanism, and Figure 15 is a diagram for explaining the operation of the shift control mechanism. A waveform diagram of control pulses, Figure 16 is a diagram showing the characteristics of the oil pressure supplied to the input and output side hydraulic servos, Figure 17 is a diagram showing the characteristics of the solenoid pressure, and Figure 18 is the output oil pressure of the shift control valve. 19 is a diagram for explaining the shift shock control process, FIG. 20 is a diagram showing the engine's best fuel consumption power line, and FIG. 21 is a diagram showing the characteristics of the engine's output performance. Fig. 22 is a diagram showing the performance curve of the fluid transmission mechanism, Fig. 23 is a diagram showing the engine equivalent fuel consumption rate curve, Fig. 24 is a diagram showing the best fuel efficiency fluid coupling output curve,
Figure 5 is a diagram showing the characteristics of the best fuel efficiency fluid coupling output rotation speed, Figures 26, 27, and 29 are diagrams for explaining the flow of processing in the electric control circuit, and Figures 28 and 30. 31 is a diagram for explaining the control setting data, FIG. 31 is a diagram for explaining the operation of the torque ratio control device, and FIG. 32 is a diagram showing the relationship between the torque ratio and the pressure ratio of the input/output side hydraulic servo. be. 30...Continuously variable transmission, 31...Input side pulley,
32...Output side pulley, 33...V belt, 60
... Hydraulic adjustment device (belt clamping force adjustment means), 6
1...Regulator valve, 64...Detent valve,
65... Throttle valve (engine output signal generating means), 66... Torque ratio valve (torque ratio signal generating means), 70... Shift control mechanism, 80... Torque ratio control device, 81... Torque ratio control valve,
84... Solenoid valve for downshift, 85...
Upshift solenoid valve, 313, 323...
...Hydraulic servo (belt clamping force generation means), 904
...Throttle sensor.

Claims (1)

[Scope of Claims] 1. A V-belt for a vehicle that changes the interval between the V-shaped grooves of the input pulley and the output pulley, changes the effective diameter of both pulleys, and continuously changes the torque ratio between both pulleys. A control device for a continuously variable transmission includes a belt clamping force adjusting means for adjusting a belt clamping force generating means for clamping a belt stretched between both pulleys, and the belt clamping force adjusting means adjusts the belt clamping force. A control device for a V-belt type continuously variable transmission for a vehicle, characterized in that the torque ratio increases in response to an increase in engine output, and the rate of increase increases in response to an increase in the torque ratio. 2. The belt clamping force adjustment means includes an engine output signal generation means that generates an engine output signal according to the engine output, and a torque ratio signal generation means that generates a signal according to the torque ratio between the input side and output side pulleys. The output signal of the engine output signal is controlled according to the output signal of the torque ratio signal generation means, and the belt clamping force adjustment means is controlled by the output signal of the engine output signal generation means. Claim 1 characterized by
A control device for a V-belt continuously variable transmission for a vehicle according to item 1. 3. The belt clamping force adjusting means increases the rate of increase in belt clamping force in accordance with an increase in torque ratio, increases the belt clamping force at a predetermined torque ratio in accordance with an increase in engine output, and adjusts the rate of increase of the belt clamping force in accordance with an increase in engine output. 2. The control device for a V-belt continuously variable transmission for a vehicle according to claim 1, wherein the increase rate with respect to the torque ratio is increased in accordance with an increase in engine output. 4. A V-belt continuously variable transmission for vehicles that changes the interval between the V-shaped grooves on the input and output pulleys and changes the effective diameter of both pulleys to steplessly change the torque ratio between both pulleys. The control device includes a hydraulic adjustment device that adjusts the hydraulic pressure supplied to a hydraulic servo that is provided on at least one of the pulleys and generates a belt clamping force using hydraulic pressure supplied from a hydraulic source, the hydraulic pressure adjustment device includes a throttle valve that generates hydraulic pressure according to the engine output, a regulator valve that regulates the hydraulic pressure supplied from the hydraulic source to the hydraulic servo, and a hydraulic pressure regulated by the regulator valve that is connected to the input side and the output side. and a torque ratio valve that is supplied to the regulator valve to act according to the torque ratio between the side pulleys, the hydraulic pressure that is applied to the regulator valve in accordance with the engine output, and the hydraulic pressure that is supplied from the hydraulic source to the hydraulic servo. The hydraulic pressure is regulated to increase in response to an increase in engine output, and the hydraulic pressure regulated by the regulator valve is applied to the regulator valve by the torque ratio valve in accordance with the torque ratio. . A control device for a V-belt type continuously variable transmission for a vehicle, characterized in that an increase rate of the oil pressure regulated by the regulator valve with respect to an increase in engine output is changed in accordance with the torque ratio. 5 A V-belt type continuously variable transmission for vehicles that changes the interval between the V-shaped grooves on the input and output pulleys, changes the effective diameter of both pulleys, and continuously changes the torque ratio between both pulleys. The control device includes a hydraulic adjustment device that adjusts the hydraulic pressure supplied to a hydraulic servo that is provided on at least one of the pulleys and generates a belt clamping force using hydraulic pressure supplied from a hydraulic source, the hydraulic pressure adjustment device includes a throttle valve that generates hydraulic pressure according to the engine output, a regulator valve that regulates the hydraulic pressure supplied from the hydraulic source to the hydraulic servo, and a hydraulic pressure that corresponds to the engine output between the input and output side pulleys. a torque ratio valve that feeds back to the throttle valve in accordance with the torque ratio of and changes a rate of change in oil pressure corresponding to the engine output with respect to a change in engine output in accordance with the torque ratio,
A hydraulic pressure whose rate of change with respect to a change in engine output is changed in accordance with the torque ratio is applied to the regulator valve, and the hydraulic pressure supplied from the hydraulic source to the hydraulic servo is increased in response to an increase in engine output. 1. A control device for a V-belt type continuously variable transmission for a vehicle, characterized in that the increase rate with respect to an increase in engine output is changed in accordance with the torque ratio.