JPH0255665B2 - - Google Patents

Description

[Detailed description of the invention]

[Industrial Application Field] The present invention relates to a V-belt continuously variable transmission for adjusting the oil pressure (line pressure) supplied to a hydraulic servo mechanism that generates a clamping force for a drive belt of a V-belt continuously variable transmission. Regarding the hydraulic adjustment device of the machine. [Prior Art] A V-belt continuously variable transmission uses an endless metal V-belt invented by Van Doorne wrapped around two pulleys. The band is characterized in that a large number of V-shaped blocks are arranged without interruption from one another, 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. FIG. 33 shows a conventional continuously variable transmission for vehicles proposed in Japanese Unexamined Patent Publication No. 157930/1983. The input shaft a is provided with an input pulley consisting of a fixed flange b and a movable flange c, and the output shaft d is provided with an output pulley consisting of a fixed flange e and a movable flange f. A belt g is stretched between the pulleys and transmits power from the input side a to the output side d. The control device of this continuously variable transmission for vehicles includes a regulator valve k that regulates the hydraulic pressure generated by the pump j to a predetermined line pressure.
is provided, the line pressure regulated by the regulator valve is supplied to the movable flange f of the output pulley via the oil path i, giving the output pulley a force to clamp the belt g, and the input pulley Movable flange c
A torque ratio control valve l selectively supplies and discharges hydraulic pressure to an oil passage h connected to the movable flanges c and f, thereby moving the movable flanges c and f. Furthermore, fluid pressure proportional to the rotational speed of the input shaft a acts on one end of the spool m of the torque ratio control valve l through a pitot tube n, while the other end of the spool m acts in conjunction with the movement of the throttle pedal. cam p
Pressure due to the rotation of is applied via lever q and spring r. Therefore, hydraulic pressure is supplied to and discharged from the movable flange c of the input pulley by the torque ratio control valve l in accordance with the rotational speed of the input side a and the movement of the throttle pedal. Furthermore, fluid pressure proportional to the rotational speed of the input shaft a acts on one end of the spool s of the regulator valve k through the pitot tube n, while the input pulley is affected by the axial movement of its movable flange c. A detection rod t is provided which is displaced in conjunction with the spool s, and the pressure of the detection rod t acts on the other end of the spool s via a lever u and a spring v. In the above configuration, the line pressure generated in the oil passage i is composed of the fluid pressure generated by the pitot tube n depending on the rotation speed of the input shaft, and the load of the spring v that changes due to the rotation of the lever u linked to the movement of the detection rod t. The hydraulic pressure increases as the torque ratio increases, and increases as the input shaft rotational speed decreases. [Problem to be solved by the invention] In a V-belt type continuously variable transmission for a vehicle, the minimum required belt clamping force to prevent slipping between the belt and the pulley is determined by the line pressure in Fig. 7. It is the product of the pressure-receiving area of the hydraulic servo of the driven pulley, and therefore has the same characteristics as the curve shown by the solid line in FIG. That is, the rate of increase in the minimum necessary belt clamping force increases as the torque ratio (corresponding to the pulley position) increases. However, in the above-mentioned conventional belt type continuously variable transmission, the signal corresponding to the torque ratio is set by the load of the spring v linked to the movement of the detection rod t, so the line pressure is linearly adjusted according to the torque ratio. If the line pressure is set low, the clamping force will be insufficient in the high torque ratio range.
There is a problem in that the belt slips when driving in a city, where the vehicle repeatedly starts and runs at low speeds, causing scuffing between the belt and the pulleys, which impairs the durability of the belt.In addition, the belt slips when the vehicle starts suddenly, making it impossible to start. If the engine goes into overdrive or kicks down while running in overdrive, the belt slips during the transition to underdrive, causing the engine to overrun, as well as scuffing between the belt and pulley, causing damage to the belt. On the other hand, if the line pressure is set high, the clamping force will be excessive in a small torque ratio range, and oil pump loss will increase during frequent high-speed driving, resulting in poor fuel efficiency. ,Also,
At high speeds, the linear velocity of the belt is high and the contact radius between the output pulley and the belt is small, so the belt bends significantly. If an excessive clamping force is applied to the belt in this state, the steel band that makes up the belt will break. However, there is a problem in that stress is concentrated on the blocks and the durability of the belt is impaired. In order to solve the above-mentioned problems, the present invention aims to reduce the line pressure supplied to the hydraulic servo mechanism that generates the clamping force of the drive belt of the V-belt type continuously variable transmission to V
Controls according to the torque ratio of the belt type continuously variable transmission,
This ensures the minimum necessary belt clamping force at a predetermined torque ratio. [Means for Solving the Problems] The hydraulic pressure adjustment device for a vehicle V-belt type continuously variable transmission of the present invention has an input shaft 214 and an output shaft 26, respectively.
Input side and output side pulleys 31, 3 with variable effective diameters are attached to a fixed flange, a movable flange, and a hydraulic servo provided on the movable flange.
2, and a drive belt 3 stretched between these pulleys.
The clamping force for the drive belt is generated by supplying hydraulic pressure to the hydraulic servo 323 on either one of the input side or output side pulleys, and the hydraulic servo 3 on the other of the input side or output side pulleys
In a hydraulic pressure adjustment device for a vehicle V-belt type continuously variable transmission, the hydraulic pressure adjustment device for a vehicle V-belt type continuously variable transmission controls the torque ratio between the input and output shafts by hydraulically adjusting the effective diameters of the two pulleys by selectively supplying hydraulic pressure to the hydraulic power source 13. 50, a regulator valve 61 that regulates the pressure of hydraulic fluid from the hydraulic source to a line pressure that supplies the one hydraulic servo;
control means 66 for controlling the line pressure to increase in accordance with an increase in the torque ratio between the input and output shafts;
The control means 66 is characterized in that it includes means for generating oil pressure in accordance with the torque ratio and ensuring the minimum required belt clamping force at a predetermined torque ratio. Note that the numbers added to the above configurations are for comparison with the drawings, and the configurations of the present invention are not limited thereby. [Operations and Effects of the Invention] According to the present invention, line pressure is supplied to a hydraulic servo that changes the torque ratio between both pulleys as a parameter and generates a clamping force for a drive belt of a V-belt continuously variable transmission. By approximating the hydraulic pressure to the characteristic curve shown in Figure 7 to ensure the minimum necessary belt clamping force at a predetermined torque ratio, the above-mentioned conventional problem can be solved, and the V-belt will not slip. It is possible to prevent the V-belt from receiving more clamping force than necessary and damage its durability, and to obtain the minimum necessary clamping force of the V-belt. Since generation of a line pressure higher than the oil pressure can be prevented, it is possible to prevent a decrease in fuel efficiency due to an increase in torque loss of the oil pump. [Examples] Examples of the present invention will be described below 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 type continuously variable transmission connected to the fluid coupling 21. 30, 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 made up of. The fluid coupling 21 is well known and consists of a pump impeller 211 connected to the output shaft 101 of the engine and a turbine runner 212 connected to the fluid coupling output shaft 214. Note that other fluid torque converters or mechanical clutches may be used in place of the fluid coupling. The V-belt continuously variable transmission 30 includes an input pulley 31 connected to an output shaft 214 of a fluid coupling, and an output shaft 26 of the V-belt continuously variable transmission arranged parallel to the input pulley 31. The output pulley 32 is connected to the output pulley 32, and the V-belt 33 is stretched between the two pulleys. The input pulley 31 includes a fixed flange 311 connected to the output side 214, and a fixed flange 311 connected to the output side 214.
It has a movable flange 312 provided so as to form a V-shaped space facing the movable flange 31.
2 is provided so as to be movable in the axial direction by a hydraulic servo 313. The output pulley 32 includes a fixed flange 321 connected to the output shaft 26 of the continuously variable transmission 30, and a movable flange 322 provided opposite to the fixed flange 321 to form a V-shaped space. , 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 side 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 4.
2, it is removably connected to the transmission case 400. In this planetary gear transmission 40 for forward/reverse switching, when hydraulic pressure is supplied to the hydraulic servo 49, the multi-plate clutch 45 engages and the rotation of the output side 26 of the continuously variable transmission remains unchanged. It is transmitted to the output side 47 of the machine 40 to enable forward driving conditions. Furthermore, when hydraulic pressure is supplied to the hydraulic servo 48, the multi-disc brake 42 is engaged and the ring gear is fixed, so that the output side 47 is connected to the output side 2 of the continuously variable transmission.
It rotates in the opposite direction to the rotation of No. 6 to enable a backward running state. The reduction gear mechanism 23 is a V-belt continuously variable transmission 3
This is to compensate for the fact that the shift range obtained with 0 is lower than the shift range achieved by a normal vehicle transmission, and for example, the torque is increased by reducing the speed by a reduction ratio of 1.45. 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.
A shift control mechanism 70 that alleviates shock during 0, N-D, and N-R shifts, and a torque ratio control device 80
Consisting of The oil pressure source 50 is connected to an oil strainer 51 from an oil sump.
Hydraulic oil pumped up by a pump 52 driven by the engine is supplied to the regulator valve 61 through an oil passage 11 to which a relief valve 53 is attached. The hydraulic adjustment device 60 includes a manual valve 62 that is manually operated by a shift lever (not shown), a detent valve 64 and a throttle valve 65 that output detent pressure and throttle pressure according to the throttle opening θ of the carburetor 102, and an output side. A torque ratio valve 6 that operates in conjunction with the movable flange 322 of the pulley 32 and supplies line pressure to the detent valve 64 according to the amount of displacement thereof, and also discharges pressure from the output hydraulic pressure feedback oil passage 9 provided in the throttle valve 65.
6, and a regulator valve 61 that regulates the hydraulic pressure supplied from the hydraulic source 50 and supplies it to the hydraulic servo 323 as line pressure.

[Table] In the table, 〇 indicates the connection status with oil path 1,
× indicates that the oil passages 3 to 5 are in a discharged pressure state. The regulator valve 61 includes a spool 611 and a regulator valve plunger 612 that controls the spool 611 by inputting detent pressure and throttle pressure. The line pressure is output from the output port 616 to the oil passage 1. Oil discharged from port 614 is supplied via oil line 12 to the fluid coupling, oil cooler, and lubrication points. The detent valve 64 is linked to and interlocks with the throttle opening θ of the rear valve of the carburetor 102.
It is equipped with a spool 641 that moves as shown in the figure, and when the throttle opening degree is 0≦θ≦ _θ1 , the detent pressure is connected to the oil passage 5 and the input port 616' provided in the regulator valve 61 as shown in FIG. 4A. When θ ₁ <θ≦100%, the oil passage 7 and the oil passage 6 which connects the detent valve 64 to the torque ratio valve 66 are communicated as shown in FIG. 4B. The throttle valve 65 includes a spool 651 which is arranged in series with the spool 641 of the detent valve via a spring 645 and has a spring 652 on its back. The opening area of the port 653 communicating with the oil passage 1 is adjusted by the action of the spool 651 that moves according to the fluctuation of the degree θ, and the throttle pressure output oil passage 8 communicating with the input port 618 provided in the regulator valve 61 is adjusted. Outputs throttle pressure. Spool 6
51 branches from the oil passage 8, and supplies the output oil pressure to a land 656 and a land 657 having a larger pressure receiving area than the land 656 via output oil pressure feedback oil passages 9 and 10 provided with orifices 654 and 655, respectively. We are receiving feedback from 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 movable flange 322
The amount of movement L is l ₃ ≦L≦l ₄ (torque ratio T is t ₂ ≧T≧t ₁ )
In this case, as shown in FIG. 5A, the spool 662 is located on the left side in the drawing, closes the input port 664 connected to the output oil pressure feedback oil passage 9 provided in the throttle valve 65, and also closes the input port 664 connected to the output oil pressure feedback oil passage 9 provided in the throttle valve 65. The output oil passage 6 is communicated with the drain port 665 to discharge pressure. The amount of movement L of the movable flange 322 is smaller than the first set value _l3 , and _l2 ≦L< _l3 (t3 _≧
When T>t ₂ ), the spool 662 is located at the intermediate portion as shown in FIG. 5B, the port 664 connected to the oil passage 9 and the drain port 666 communicate with each other, and the oil passage 9 is depressurized. Movement amount L of movable flange 322
is smaller than the second set value l ₂ and 0≦L≦l ₂ (t ₄ ≧T
>t ₃ ), the spool 6 is closed as shown in Fig. 5C.
62 is located on the right side in the figure, and a port 663 connected to the oil passage 1 and the oil passage 6 are interlocked to supply line pressure to the oil passage 6. The shift control mechanism 70 has a spring 71 on one side.
A shift control valve 71 equipped with a spool 712 that receives line pressure from an oil chamber 713 provided at the other end, an orifice 72 provided in the oil passage 1 that supplies line pressure to the oil chamber 713, and the orifice 7
2 and an oil chamber 713, and a solenoid valve 74 that is controlled by an electric control circuit to be described later and adjusts the oil pressure in the oil chamber 713. The solenoid valve 74 is turned on and the drain port 7
41 is opened to discharge pressure from the oil chamber 713, the spool 712 of the shift control valve 71 is connected to the spring 7.
The oil passage 13 is moved to the left in the figure by the action of 11 and communicates with the hydraulic servo 49 that operates the multi-disc clutch 45 of the planetary gear transmission 40, and the oil passage 14 that communicates with the hydraulic servo 48 that operates the multi-disc brake 42. are connected to drain ports 714 and 715, respectively, to discharge pressure and release the multi-disc clutch 45 or the multi-disc brake 42. When the solenoid valve 74 is off, the drain port 741 is closed, and the spool 712 is located on the right side in the figure due to the line pressure supplied to the oil chamber 713, and connects the oil passages 3 and 4 to the oil passage 13, respectively.
and the oil passage 14 to engage the multi-disc brake 42 or the multi-disc clutch 45. In this embodiment, the shift control valve 71 includes an oil passage 13 and an oil passage 1.
An oil chamber 717 and an oil chamber 716 are provided to feed back the output hydraulic pressure of the multi-disc clutch 45 and the multi-disc brake 42, thereby alleviating the rise in the output hydraulic pressure and preventing 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 and 83, a downshift solenoid valve 84, and an upshift solenoid valve 85. The torque ratio control valve 81 has a spring 8 on one side.
a spool 812 with a spool 11 installed behind it, oil chambers 815 and 816 at both ends to which line pressure is supplied from the oil passage 1 through orifices 82 and 83, respectively;
An input port 817 that communicates with the oil passage 1 to which line pressure is supplied and whose opening area increases or decreases according to the movement of the spool 812 and the V-belt continuously variable transmission 3
An oil chamber 819 is provided with an output port 818 that communicates with the hydraulic servo 313 of the input pulley 31 of No. A drain port 813 is provided to evacuate the oil chamber 815 in accordance with the movement of the oil chamber 815. The downshift solenoid valve 84 and the upshift solenoid valve 85 are the torque ratio control valve 8, respectively.
1 oil chamber 815 and oil chamber 816,
Both are operated by the output of an electric control circuit to be described later, and evacuate the pressure in the oil chamber 815 and the oil chamber 816, respectively. FIG. 6 shows the solenoid valve 74 of the foot control mechanism 70 and the downshift solenoid valve 84 of the torque ratio control device 80 in the hydraulic control circuit shown in FIG.
The configuration of an electric control circuit 90 that controls the upshift solenoid valve 85 is also shown. 901 is a shift lever switch that detects whether the shift lever is shifted to P, R, N, D, or L; 902 is a rotation speed sensor that detects the rotation 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; 905 is a speed detection processing circuit that converts the output of the rotational speed sensor 902 into voltage; and 906 is the converter that converts the output of the vehicle speed sensor 903 into voltage. a vehicle speed detection circuit; 907 is 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 read-on memory (ROM) that stores programs for controlling the solenoid valves 74, 84, and 85 and data necessary for control. , 914 is a random access memory RAM that temporarily stores input data and parameters necessary for control, 915 is a clock, 916 is an output interface, and 917 is a solenoid output driver, which connects the output of the output interface 916 to an upshift solenoid. The output is changed to the operating output of the valve 85, the downshift solenoid valve 84, and the shift control solenoid valve 74. Input interface 908-911
A data bus 9 is connected between the CPU 912, ROM 913, RAM 914, and output interface 916.
18 and an address bus 919. Next, the torque ratio valve 6 which is a feature of the present invention
6. The operation of the hydraulic pressure adjustment device 60 composed of the detent valve 64, the throttle valve 65, the manual valve 62, and the 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 run a vehicle with low fuel consumption, it is necessary to bring the line pressure supplied to the hydraulic control circuit close to the minimum necessary, and in a continuously variable transmission, this line pressure is applied to the input pulley 3.
Each hydraulic servo of the V-belt 1 and the output pulley 32 is defined by a hydraulic pressure that allows torque to be transmitted 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 ratio that can be achieved by both pulleys, so as shown by the dotted line, 20 It is desirable to set the line pressure to the line pressure as shown by the broken line which is about % larger, and it is desirable to set the line pressure to a higher line pressure characteristic as shown by the dashed line even when the throttle opening θ=0 during engine braking. In this embodiment, the line pressure, which is the output of the regulator valve 61, is controlled by the hydraulic pressure adjustment device 60.
Shift position of manual valve 62 (L, D, N,
R, P), throttle opening θ, and changes in the torque ratio between the input and output shafts as follows. (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 passage 4 and the oil passage 5 are depressurized. At this time, in the shift control mechanism 70, if the shift control solenoid valve 74 is in the OFF state and line pressure is supplied to the oil chamber 713, the spool 712 is positioned to the right, so that the oil passage 3 and the oil The line pressure supplied to the oil passage 3 is connected to the hydraulic servo 4 of the multi-plate clutch 45 for forward movement through the oil passage 13.
9, 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 a port 663 communicating with the oil passage 1, and communicates the oil passage 6 with a drain boat 665 to discharge pressure. As a result, no detent pressure is generated in the oil passage 7 regardless of the throttle opening degree θ. The throttle valve 65 also has a torque ratio valve 66 connected to the oil passage 9.
port 664 is closed, and the spool 65
1 receives feedback pressure from the land 657 in addition to the land 656, so the oil passage 8 maintains the throttle pressure having the characteristic shown in FIG. 8C for the throttle opening θ.
It is output to the regulator valve plunger 612 of the regulating valve 61 through the. As a result, the line pressure output from the regulator 61 is reduced to the area shown in FIG.
The result will be as shown in Figure 0 (E). When the torque ratio T is _t2 <T≦ _t3 . As shown in FIG. 5B, the torque ratio valve 66 closes the port 663 and allows the oil passage 9 and the drain port 666 to communicate with each other. Also, the oil passage 6 is the port 66
The pressure is exhausted through 5. Therefore, no detent pressure is 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 is expressed by the characteristic curve shown in FIG. 8D. Ru. The line pressure at this time has the characteristics shown in area R in FIG. 9 and area G in FIG. When the torque ratio T is t ₃ <T≦t ₄ . As shown in Fig. 5C, the oil passage 9 is connected to the drain port 6.
66, and thus the throttle pressure is shown in FIG. 8D as above. However, port 66
3 is opened and oil passage 1 and oil passage 6 communicate with each other, the throttle opening θ is within the range of 0≦θ≦θ ₁ %,
The spool 641 of the detent valve 64 is shown in FIG. 4A.
As shown in the left side of the figure, the spool 6
41 closes the oil passage 6, and the oil passage 7 closes the oil passage 5.
However, when the throttle opening degree θ is θ ₁ %<θ≦100%, the spool 641 moves to the right side in the figure as shown in FIG. 4B, and the oil passage 6 and the oil passage 7 communicate with each other, and detent pressure is generated in the oil passage 7. As a result, the line pressure is reduced to θ as shown in area 9 of FIG.
= θ ₁ %, the characteristic changes stepwise. (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 at the detent valve 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 pressure in the oil passage 7 is exhausted through the oil passage 6 and the drain port 665 of the torque ratio valve as shown in FIG. 4B, and no detent pressure is generated. First, the throttle pressure is the same as in the D position. Therefore, the line pressure is 11th
The characteristics are shown in the figure below. When the torque ratio T is _t2 <T≦ _t3 . The difference from the above is that in the torque ratio valve 66, the oil passage 9 communicates with the drain port 666 and the pressure is discharged, and the throttle pressure that the throttle valve 65 outputs to the adjustment valve 61 via the oil passage 8 increases. As a result, the line pressure is expressed by a characteristic curve as shown 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 detent pressure regardless of the throttle opening.
The regulator valve 61, which receives the detent pressure and the same throttle pressure as above, outputs the line pressure shown in FIG. (R position) Oil passage 4 and oil passage 5 at manual valve 62
is in communication 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 located on the right side, oil passage 4 and oil passage 14
The line pressure supplied to the oil passage 4 is supplied to the hydraulic servo 48 of the reverse multi-disc brake 42 through the oil passage 14, and the vehicle enters the reverse traveling state. (P position and N position) Oil passages 3, 4 and 5 in manual valve 62
Since both are exhausted, the regulator valve 61
The line pressure that is the output of is the same as at the D position. When the manual valve 62 is shifted to the D, N, and P shift positions in this line pressure adjustment, the line pressure when the torque ratio T is in the range of t ₃ < T ≦ t ₄ is determined by the characteristics shown in Fig. 10. The reason for setting the throttle opening θ as low as ₁ % or less as shown in curve 2 is because if the line pressure is set high in operating conditions such as idling where the throttle opening θ is small and the pump discharge amount is low, the oil temperature will be high. If there is a large oil leak from various parts of the hydraulic circuit, it will be difficult to maintain line pressure, and the oil temperature will further rise due to a decrease in the amount of oil supplied to the oil cooler, which can easily cause trouble. It is designed to prevent this. 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 opening is The reason why the line pressure is set high under the operating condition where θ is θ ₁ % or less is because relatively high oil pressure is required during engine braking even when the throttle opening is low. In this way, as shown in Fig. 9, by bringing the line pressure close to the minimum necessary oil pressure shown in Fig. 7,
Since the power loss caused by the pump 52 can be reduced, 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 explained in FIG. 6 will be explained with reference to the program flowchart 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 P position or N position,
Both solenoid valves 84 and 85 are controlled by the P position or N position processing subroutine shown in FIG.
OFF (step 931) and set to P or N state.
It is stored in 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, the shift shock is activated to relieve the shift shock associated with the N-R shift and N-D shift, respectively. Control processing is performed (steps 940 and 950). This shift shock control process will be described in detail below. First, the shift control mechanism 70 includes a solenoid valve 7 controlled by the output of the electric control circuit 90 described above.
The action of 4 adjusts the timing of supplying and discharging hydraulic pressure to the hydraulic servos 48 and 49 of the planetary gear transmission 40 to prevent shock during shifting, and the action of the pressure limiting valve 73 adjusts the timing of supplying and discharging hydraulic pressure to the hydraulic servos 48 and 49. It has the function of keeping the upper limit of the hydraulic pressure supplied to the engine below a set value, and limits the engagement pressure of the clutch and brake. In this embodiment, as shown in FIG. 14, the pressure-receiving areas of the lands provided on the spool 712 of the shift control valve 71 are S ₁ , S ₁ , S ₁ , S ₂ and the elastic force of the spring 711 in order from the left in the figure. F _S1 , and assuming that the oil pressure in the oil chamber 713 is P _s , the oil 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 servo 48 of the multi-disc brake 42 is engaged when moving backward. The supplied hydraulic pressure P _b is given as follows from the equations and equations, which are oil pressure balance equations for 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 pressure limiter inserted in the pressure limiting valve 73 The pressure receiving area of the return body 731 is S ₃ , and the valve body 73
1, the pressure limiting valve ₇₃ operates at the maximum pressure Plimit of _Ps according to a 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 /S2)]...The} solenoid valve 74 generates a solenoid pressure _Ps in the oil chamber 713 according to the duty (%) given by the following equation. Duty = (Solenoid ON time in one cycle/Solenoid operating cycle) x 100 (% ^{) This duty control has a pulse width of L * -nM *} (n=
1, 2, 3, . . . ), and the pulse width gradually decreases, is applied to the shift control solenoid valve 74 shown in FIG. By duty-controlling the shift control solenoid valve 74 in this manner, a hydraulic pressure P _s adjusted according to the duty is generated in the oil chamber 713 of the shift control valve 71. The solenoid pressure P _s shown in FIG. 17 is amplified by the shift control valve 71 to obtain the hydraulic pressure P _c or P _b shown in FIG. 18 to be supplied to the hydraulic servo 48 or 49. When relaxing the engagement shock during the N-D shift and N-R shift, the rise of the hydraulic pressure P _b or P _c supplied to the hydraulic servo 48 or the hydraulic servo 49 is adjusted by rolling as shown in the hydraulic characteristic curve shown in Fig. 16. , In the figure, multi-plate clutch 4 between AC
5 or complete the engagement of the multi-disc brake 42. The solenoid valve 7 for controlling the hydraulic pressure supplied to the hydraulic servo 48 or 49 in this way
Shift shock control processing 9 to control 4
40,950 program flowcharts first
Shown in Figure 9. FIG. 19 shows 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 FLUG during shock control processing is on. If FLUG is on, it means that shift shock control processing is in progress and the process proceeds to step 946. If FLUG is not on, shift shock control processing is started. By comparing the shift lever position stored in the RAM 914 with the current shift lever position, it is determined whether or not the shift lever has changed from the P position or the N position to the R position (step 942). Determining whether there is a change from the N position to the D position (step 943)
I do. If any changes occur,
In steps 944 and 945, the corresponding parameters K ^* , L ^* , and M ^* are set, parameter K is set to 0, and FLUG is turned on, indicating that the shock control process is performed (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 not greater than 0, K is
K ^* -1, L is set as L ^* , and L ^* is set as L ^* -M ^* (step 947), and in step 948 it is determined whether L is less than or equal to 0. If L is less than or equal to 0, step 951 is performed. If L is less than 0, it is assumed that all shock control processing has been completed.
Turn off FLUG. One cycle K ^* in step 946
When the parameter K for determining the end of is greater than 0, K-1 is set to K (step 950),
Next, it is determined whether the parameter L, which determines the end of the on-time in one cycle K, is 0 (step 951). When L is 0, a command to turn off the solenoid valve 74 is issued (step 952), and when L is not 0, a command to turn on is issued (step 953), after which L-1 is set to L and the process returns. Similar shift shock control processing can also be performed using programmable timer 920 in FIG. Next, we will discuss the first aspect of shift control, which is a feature of the present invention.
This will be explained by returning to FIG. 2. N-D shift shock control processing 950
Next, the input side pulley rotation speed sensor 902
The actual input side pulley rotation speed N is read (step 923), and then it is determined whether or not the throttle opening degree θ read in step 921 is 0 (step 924). When θ≠0, the input side Execute subroutine 960 to set pulley target rotation speed N ^* to the best fuel economy input side pulley rotation speed, and set θ=
0 and the throttle is fully closed, 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 set to the D position. When set, 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 target rotation speed N ^* of the input pulley is executed.
Set to appropriate values for each. A subroutine 960 for setting the input pulley target rotation speed N ^* described above to the input pulley rotation speed for the best fuel efficiency 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 determined, and by connecting these points, the best value for each horsepower can be determined. The best fuel efficiency power line indicating the engine operating state resulting in fuel efficiency can be obtained. However, 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. To this end, 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 preliminarily stored in the ROM 913 as data. Set the address of the best fuel consumption input side pulley rotation speed N _B data corresponding to the throttle opening θ (step 961), read out the best fuel efficiency input side pulley rotation speed N _B from the set address (step 962), Best fuel efficiency input pulley rotation speed corresponding to the 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 971, calculates the acceleration α at that point 972, and then determines whether the acceleration α is an appropriate acceleration A for the vehicle speed 9
73. When the acceleration α is greater than the acceleration A, the target rotation speed of the input pulley is set in order to downshift.
N ^* is set to a value larger than the current input pulley rotation speed N, and when the acceleration α is not larger than the acceleration A, the input pulley target rotation speed N ^* is set to the best fuel efficiency input pulley corresponding to the throttle opening θ. Set the rotation speed to N _B and return to 975. The relationship between vehicle speed and appropriate acceleration A is determined in advance by experiment or calculation 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, read the vehicle speed V in 981, then read the vehicle speed V
982, the current torque ratio T is calculated from the input pulley rotation speed N using the following equation. 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, it is determined whether the current torque ratio T is smaller than the torque ratio T ^* that provides safe and appropriate engine braking for the vehicle speed, and if the torque ratio T is smaller than the torque ratio T ^* , the engine is shifted down. Input side pulley target rotation speed N ^* is set to a value larger than the current input side pulley rotation speed N, and the process returns to step 984. If the torque ratio T is not smaller than the torque ratio T ^* , the input pulley target rotation speed N ^* is set to the current input pulley rotation speed N, 985, and the process returns. 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, in FIG. 12, the actual input pulley rotation speed N and the best fuel efficiency input pulley rotation speed N ^* are compared (step 927), and N<
When N ^* , a command to operate the downshift solenoid valve 84 is issued (step 928), when N>N ^* , a command to operate the upshift solenoid valve 85 is issued (step 929), and when N=N ^* , a command to operate the upshift solenoid valve 85 is issued (step 929). A command to turn off the solenoid valves 84 and 85 is issued (step 920). The torque ratio control device 80 controls the increase or decrease in the gear ratio between the input and output pulleys by comparing the rotational speed of the input pulley with the best fuel efficiency obtained in FIG. 17 and the actual rotational speed of the input pulley. Control device 80
This is done by actuating two solenoid valves 84 and 85 provided in the input pulley to match the actual input pulley rotation speed to the input pulley rotation speed for the best fuel efficiency. (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
It will be turned off. As a result, the oil pressure P ₁ in the oil chamber 816
is the line pressure, and the oil pressure _P2 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, _P2 is exhausted, so the spool 812
It is moved to the right in the diagram by P ₁ . Spool 81
2 is moved to the right, the drain boat 813 is closed. In this case, by providing a flat notch 812b at the land edge between the drain port 813 and the spool 812, it becomes possible to more stably hold the spool 812 at an equilibrium point at an intermediate position as shown in 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, in the position where the spool 812 is balanced, the drain port 814
A flat notch 812a is provided at the land edge with the spool 812 so that the oil passage 1 is closed and the oil passage 1 is slightly open, thereby compensating for oil leakage in the oil passage 2. (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.
Hydraulic servo 31
3 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.
By controlling the ON time of the solenoid valve 84 in this way, the torque ratio is increased and a downshift is performed. 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. Pi, and the hydraulic pressure of the output side hydraulic servo 323 is Po, then Po/Pi has characteristics as shown in the graph of FIG. 32 with respect to the torque ratio T. = 1.5 (point a in the diagram), then release the accelerator and change θ = 30%
In this case, if Po/Pi is maintained as it is, the transition will be to point b in the figure where the torque ratio T = 0.87, and conversely, if the torque ratio T = 1.5 is maintained, the torque controlling the input pulley will be The value of Po/Pi is increased by the output of the ratio control mechanism 80 and changed to the value at point c in the figure. By controlling the value of Po/Pi as necessary in this way, it is possible to set an arbitrary torque ratio corresponding to any load condition.

[Brief explanation of drawings]

Fig. 1 is a schematic diagram of a V-belt type continuously variable transmission for vehicles, Fig. 2 is a diagram showing one embodiment of a hydraulic control circuit in the continuously variable transmission of the present invention, and Fig. 3 explains the operation of the manual valve. FIG. 4 is a diagram for explaining the operation of the detent valve and throttle valve. FIG. 5 is a diagram for explaining the operation of the torque ratio valve. FIG. 6 is a diagram for explaining one embodiment of the present invention. 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. 9, 10, and 11 are the diagrams shown in this book. FIGS. 12 and 13 are diagrams showing line pressure characteristics obtained by the control device of the invention, and FIGS.
Figure 4 is a diagram for explaining the operation of the shift control mechanism, Figure 15 is a waveform diagram of control pulses, Figure 16 is a diagram showing the characteristics of the hydraulic pressure supplied to the input and output side hydraulic servos, and Figure 17. is a diagram showing the characteristics of the solenoid pressure, Fig. 18 is a diagram showing the characteristics of the output oil pressure of the shift control valve, Fig. 19 is a diagram to explain the shift shock control processing, and Fig. 20 is the best fuel efficiency power line of the engine. FIG. 21 is a diagram showing the output performance characteristics of the engine, FIG. 22 is a diagram showing the performance curve of the fluid transmission mechanism, FIG. 23 is a diagram showing the equal fuel consumption rate curve of the engine, and FIG. Figure 25 is a diagram showing the best fuel efficiency fluid coupling output curve, Figure 25 is a diagram showing the characteristics of the output rotation speed of the fluid coupling with the best fuel efficiency, Figures 26, 27, and 29 explain the flow of processing in the electric control circuit. Figures 28 and 30 are diagrams for explaining the control setting data, Figure 31 is a diagram for explaining the operation of the torque ratio control device, and Figure 32 is a diagram for explaining the torque ratio and input/output. Diagram showing the relationship with the pressure ratio of the side hydraulic servo,
FIG. 33 is a schematic diagram of a conventional continuously variable transmission. 30...V belt type continuously variable transmission, 31...Input side pulley, 32...Output side pulley, 33...V belt, 6
0...Hydraulic pressure adjustment device, 80...Torque ratio control device, 3
11,321...Fixed flange, 312,322...
Movable flange, 313, 323...hydraulic servo.

Claims (1)

[Scope of Claims] 1. Input and output pulleys with variable effective diameters, each having a fixed flange, a movable flange, and a hydraulic servo installed on the movable flange, which are attached to the input shaft and the output shaft, respectively, and between these pulleys. It consists of a tensioned drive belt, and generates a clamping force for the drive belt by supplying hydraulic pressure to a hydraulic servo on either the input side or output side pulley,
A V-belt for a vehicle that controls the torque ratio between the input and output shafts by selectively supplying hydraulic pressure to a hydraulic servo on the other of the input-side and output-side pulleys to hydraulically adjust the effective diameters of both the pulleys. A hydraulic pressure adjustment device for a continuously variable transmission includes a hydraulic pressure source, a regulator valve that regulates hydraulic oil from the hydraulic source to a line pressure that is supplied to one of the hydraulic servos, and a torque ratio between the input and output shafts to convert the line pressure. control means for controlling the hydraulic pressure to increase in accordance with an increase in the torque ratio, the control means generating hydraulic pressure according to the torque ratio;
1. A hydraulic pressure adjustment device for a V-belt type continuously variable transmission for a vehicle, comprising means for ensuring a minimum necessary belt clamping force at a predetermined torque ratio. 2. The control means is a torque ratio valve that generates oil pressure according to the torque ratio between the input and output shafts, and the regulator valve inputs the output oil pressure of the torque ratio valve and adjusts the torque ratio between the input and output shafts. A hydraulic pressure adjustment device for a V-belt type continuously variable transmission for a vehicle according to claim 1, wherein the hydraulic pressure adjustment device outputs a corresponding line pressure. 3. The torque ratio valve has a plurality of supply oil passages to which hydraulic pressure is supplied according to the torque ratio between the input and output shafts, and the supply oil passages are connected to the oil chamber on the pressure increasing side of the regulator valve, and the input and output 3. The hydraulic pressure adjustment device for a V-belt type continuously variable transmission for a vehicle according to claim 2, wherein the line pressure is controlled in accordance with the torque ratio between the shafts. 4. Claim 2, wherein the torque ratio valve detects the amount of displacement of the movable pulley and acts to supply line pressure to the supply oil passage in accordance with the amount of displacement. Hydraulic pressure adjustment device for a vehicle V-belt type continuously variable transmission as described in 2.