JPH0221468B2 - - Google Patents

Description

[Detailed description of the invention]

The present invention is installed in a hydraulic control device of a continuously variable automatic transmission for vehicles using a V-belt continuously variable transmission operated by a hydraulic servo, and controls the reduction ratio (torque ratio) of the V-belt continuously variable transmission. This invention relates to a reduction ratio control mechanism. A V-belt type continuously variable transmission is used as a continuously variable automatic transmission for a vehicle in combination with a forward/reverse switching device and a fluid coupling. However, in this continuously variable automatic transmission for a vehicle, if the continuously variable transmission section is not at the maximum reduction ratio when the vehicle is started, sufficient torque cannot be obtained and a smooth start cannot be achieved. In a V-belt continuously variable transmission, it is generally difficult to change the reduction ratio while the vehicle is stopped, and therefore, when the vehicle suddenly stops, a rapid downshift is required. One of the main factors that determines the speed of this downshift is the operating speed (response) of the valve in the reduction ratio control mechanism. As shown in Fig. 37, the conventional reduction ratio control mechanism increases the supply of pressure oil during downshifts (decelerations) and decreases the supply during upshifts. Provided between the supply oil passage 2 and the line pressure supply oil passage 1 from the hydraulic source,
A reduction ratio control valve 80 that controls communication between the hydraulic servo and a hydraulic source or an oil drain port and has drain ports (oil drain ports) 813 and 814, and a spring 811 that is mounted on the back of a spool 812 of the reduction ratio control valve. and,
An upshift electromagnetic solenoid valve 84 and a downshift electromagnetic solenoid valve 85 control the spool.
In the downshift, the spool is moved from the spring 811 by the discharge of the hydraulic pressure applied to the spool by the downshift electromagnetic solenoid valve 85.
is compressed and moved, and the line pressure maintained by the closing operation of the upshift electromagnetic solenoid valve is released to the oil drain port 81.
The structure is such that the spool gradually leaks from the spring 811 and stops at the equilibrium position of the external force applied to the spool during downshifting.
Because the robot was moving against the load of the robot, a rapid response could not be obtained. SUMMARY OF THE INVENTION An object of the present invention is to provide a reduction ratio control mechanism for a continuously variable automatic transmission for a vehicle that provides a response of a reduction ratio control valve sufficient for rapid downshifting. The present invention provides a reduction ratio control valve that connects the hydraulic servo of the input pulley with a hydraulic source or an oil drain port, a spring that applies a spring load to the spool of the reduction ratio control valve, and a spring that applies a spring load from both of the spools. In a reduction ratio control mechanism consisting of an upshift electromagnetic solenoid valve and a downshift electromagnetic solenoid valve that discharge hydraulic pressure according to vehicle running conditions, the spring load by the spring is applied in the direction in which the spool moves during downshifting. This is the basic outline. Next, the present invention will be explained based on an embodiment shown in the drawings. FIG. 1 shows a V-belt type continuously variable automatic transmission for vehicles that is controlled by a hydraulic control device. This continuously variable automatic transmission 100 for a vehicle includes a V-belt type continuously variable transmission 200, a fluid coupling 300 such as a torque converter and a fluid coupling connected to the input side of the transmission 200, and in this embodiment, the transmission A planetary gear transmission 400 for forward/reverse switching connected to the output side of the machine 200, a reduction gear mechanism 500 connected to the output side of the planetary gear transmission 400, and a reduction gear mechanism 500 connected to the reduction gear mechanism 500. It is composed of a differential gear 600. Note that all mechanisms other than the planetary gear mechanism are used as the forward/reverse switching mechanism. 101 is an input shaft of the continuously variable automatic transmission 100 connected to the engine output shaft, and 102 is a tubular first intermediate shaft forming the input shaft of the V-belt type continuously variable transmission 200, which is connected to the input shaft 101 and the first intermediate shaft. The shaft 102 is connected via a fluid coupling 300, which is the torque converter of this embodiment. 103 is an output shaft of the continuously variable automatic transmission 100, and 104 is arranged coaxially outside the output shaft 103, and is a V-belt continuously variable transmission 200.
a tubular second intermediate shaft which is an output shaft of the second intermediate shaft;
The intermediate shaft 104 and the output shaft 103 include a forward/reverse switching planetary gear transmission 400, a third intermediate shaft 105, a reduction gear mechanism 500, and a differential gear 60.
It is connected via 0. 106 and 107 are movable flanges that are slidably fitted to the first intermediate shaft 102 and the second intermediate shaft 104, respectively, and have tubular bearing portions 6A and 7A along the intermediate shafts 102 and 104, respectively; A first cylinder 1 is provided concentrically with the respective intermediate shafts 102 and 104 using the movable flanges 106 and 107 as side walls.
08 are welded together to form a first cylinder 109. 110 and 111 are fixed flanges integrally formed with the first intermediate shaft 102 and the second intermediate shaft 104, respectively, and the movable flange 106 and the fixed flange 110 and the movable flange 107 and the fixed flange 111 respectively correspond to the V-belt 112. It forms receiving V-shaped spaces 113 and 114, and also constitutes an input pulley A and an output pulley B. 115 and 11
6 are first cylinders 108 and 109, respectively.
A first fixed wall inserted into the cylinder 1
Flange portion 15A in contact with inner walls of 08 and 109
and 16A, tubular portions 15B and 16B continuous to the flange portions 15A and 16A, and respective intermediate shafts 10 continuous to the tubular portions 15B and 16B.
It has fixing parts 15C and 16C fixed to 2 and 104. The first fixed walls 115 and 116 are the movable flange 10 which is the side wall of the cylinder.
6 and 107, respectively, to form first annular oil chambers 117 and 118. 119 and 1
A second fixed wall 20 is formed integrally with the second cylinders 121 and 122 that are fitted onto the first cylinders 108 and 109, respectively, and is in contact with the fixed parts 15C and 16C of the first fixed wall, and axis 1
02 and 104. The tip end (right side in the figure) of the second cylinder 121 is bent in the outer radial direction to form a flange-shaped portion 21A, and teeth 21B are formed on the outer peripheral side of the flange-shaped portion 21A. Reference numeral 123 designates an electromagnetic pickup mounted at a predetermined position corresponding to the flange-shaped portion 21A of the automatic transmission case 700. The electromagnetic pickup 123 and the flange-shaped portion 21A constitute a device for detecting the input pulley rotation speed, that is, the rotation speed of the first intermediate shaft 102. The second fixed wall 119
and 120 is an annular plate-shaped pressure receiving plate slidably inserted between the second cylinders 121 and 122 integral with the second fixed wall and the tubular parts 15B and 16B of the first fixed wall. Second annular oil chambers 126 and 127 are formed between 124 and 125. 128 and 129 are spheres inserted into axial grooves provided on both the sliding surfaces of the first intermediate shaft 102 and the movable flange 106 and the second intermediate shaft 104 and the movable flange 107, respectively. This functions to prevent relative rotation of the intermediate shafts 102 and 104. The V-belt type continuously variable transmission 200 has a V-belt 11
2, an input pulley A, an output pulley B, and hydraulic servos C and E for the pulleys A and B. This V-belt type continuously variable transmission 200 receives detection information from the input pulley rotation speed detection device, vehicle speed, throttle opening, etc., and operates the first oil chambers 117 and 118, the second oil chamber 126 and By controlling the hydraulic pressure applied to hydraulic servos C and E having 127, movable flanges 106 and 107 are driven, and the widths of V-shaped spaces 113 and 114 are increased or decreased. and V in contact with output pulley B
The rotation radius of the belt 112 increases or decreases to achieve stepless speed change depending on the running condition of the vehicle. The fluid coupling 300, which is a torque converter, has a known configuration including a pump impeller 301, a turbine runner 302, a stator 303, and a one-way clutch 304. The forward/reverse switching planetary gear transmission 400 includes a hydraulic servo 40 that operates a second intermediate shaft 104 that is the output shaft of the continuously variable transmission 200 and a multi-disc clutch 401.
A ring gear 404 is connected via a drum 403 having two cylinders 402A, and a third intermediate shaft 105, which is the output shaft of the planetary gear transmission 400, is connected by spline fitting, and the drum 40
3 and a sun gear 405 connected via the multi-plate clutch 401, a planetary gear 406 rotatably meshed between the sun gear 405 and the ring gear 404, and a planetary gear 406 rotatably supported. A planetary carrier 408 engaged with the automatic transmission case 700 via a multi-disc brake 407 , a hydraulic servo 402 that operates a multi-disc clutch 401 , and a multi-disc brake 4
It is composed of a hydraulic servo 409 that operates 07. This forward/reverse switching planetary gear transmission 400
When the multi-disc clutch 401 is engaged and the multi-disc brake 407 is released, a forward gear with a reduction ratio of 1 is obtained, and when the multi-disc clutch 401 is disengaged and the multi-disc brake 407 is engaged, a forward gear is obtained. It becomes a reverse gear with a reduction ratio of 0.7. This reduction ratio of 0.7 during reversing is smaller than the reduction ratio of a normal automobile transmission during reversing, but in this example, the reduction ratio (for example, 2.4) obtained in a V-belt continuously variable transmission and the reduction Since the speed reduction is performed in the gear mechanism 500, an appropriate speed reduction ratio can be obtained as a whole. FIG. 2 shows a hydraulic circuit of a hydraulic control device for controlling the continuously variable automatic transmission in the transmission device shown in FIG. The hydraulic control device includes a hydraulic power source 50 and a hydraulic adjustment device 6.
0, a shift control mechanism 70 that controls the timing of engagement of a multi-disc brake and a multi-disc clutch in the star-induced gear transmission 400 and alleviates the impact during N-D and N-R shifts, and the deceleration according to the present invention. It consists of a ratio control device 80. 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 6 that output detent pressure and throttle pressure according to the throttle opening θ of the engine.
5. A torque ratio valve 66 that works in conjunction with the movable flange 107 of the output pulley B and supplies line pressure to the detent valve 64 according to the amount of displacement of the movable flange 107, and discharges pressure from the output hydraulic pressure feedback oil passage 9 provided in the throttle valve 65, and the hydraulic pressure. It is comprised of a regulator valve 61 that regulates the pressure of the pressure oil supplied from the source 50 and supplies it to various parts of the hydraulic pressure adjustment device 60 as line pressure. A hydraulic source 50 supplies hydraulic oil pumped up from an oil strainer 51 by a pump 52 driven by an engine to an oil passage 1 to which a relief valve 53 is attached.
1 and is supplied to the regulator valve 61. As shown in FIG. 3, the manual valve 62 has a spool 621 that is set in positions P, R, N, D, and P, corresponding to the shift positions P, R, N, D, and L of the shift lever provided in the exercise seat.
It is set at each position D and L, and connects the oil passage 1 to which line pressure is supplied and the output oil passages 3 to 5 as shown in the table.

[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. is adjusted, and line pressure is output from the output port 616 to the oil passage 1. Oil is supplied from the port 614 through the oil passage 12 to the fluid coupling, oil cooler, and parts requiring lubrication. The detent valve 64 includes a spool 641 that moves in conjunction with the throttle opening θ of the engine as shown in FIG. 4, and when 0≦θ≦ _θ1 , the oil passage 5 and the regulator valve When θ<θ≦100%, the oil passage 7 and the detent valve 64 are connected to the torque ratio valve as shown in FIG. 4B. 66. The throttle valve 65 includes a spool 651 which is connected in series with the spool 641 of the detent valve via a spring 645 and has a spring 652 on its back. By the action of the spool 651 that moves in response to fluctuations, the opening area of the port 653 communicating with the oil passage 1 is adjusted, 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. The spool 651 is
Each branches from the oil passage 8 and has an orifice 6.
The land 65 is connected to the land 65 through output oil pressure feedback oil passages 9 and 10 provided with output oil pressures 54 and 655.
6 and a land 657 which has a larger pressure receiving area than the land 656 receives feedback of the output hydraulic pressure. The torque ratio valve 66 includes a spool 662 linked to the movable flange 322 of the output pulley B 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 ≧ When T≧t ₁ ), the spool 662 is located on the left side of the figure as shown in FIG. Daytent valve 6
4 is communicated with a drain port 665 to discharge pressure. Movement amount L of movable flange 322
When l ₂ ≦L≦l ₃ (t ₃ ≧T>t ₂ ), the spool 662 is located in the middle part as shown in FIG.
6 is in communication with the oil passage 9, and the pressure in the oil passage 9 is exhausted. Movement amount L is 0
When ≦L≦l ₂ (t ₄ ≧T> ₃ ), the spool 662 is located on the right side in the figure as shown in FIG. Line pressure is supplied to the oil passage 6. Further, the spool 662 is slidably interlocked with the movable flange 322 of the output pulley B which is in a rotating state, and as shown in FIG.
2. Since the valve has a structure in which there is no spring, oil pressure, etc. that can prevent movement in the axial direction, it is not possible to prevent movement of the movable flange, and it also prevents wear of sliding parts that have large relative speeds. can do. 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 an electromagnetic solenoid valve 74 that adjusts the oil pressure in the oil chamber 713 and is controlled by an electric control circuit to be described later. When the electromagnetic solenoid valve 74 operates to open the drain port 741 and evacuate the oil chamber 713, the shift control valve 7
The spool 712 of No. 1 is moved to the right in the figure by the action of the spring 711, and the hydraulic servo 402 operates the multi-disc clutch 401 of the planetary gear transmission 400.
The oil passage 13 that communicates with the hydraulic servo 409 that operates the multi-disc brake 407 and the oil passage 14 that communicates with the hydraulic servo 409 that operates the multi-disc brake 407 are connected to the drain ports 714 and 715, respectively, to discharge pressure, and the multi-disc clutch 401 or the multi-disc brake 407 is activated. let go. Electromagnetic solenoid valve 7
4 is not operating, the drain port 741 is closed, and the spool 712 is located on the left side in the figure due to the line pressure supplied to the oil chamber 713.
and the oil passage 4 are connected to the oil passage 13 and the oil passage 14, and a multi-disc brake 407 or a multi-disc clutch 4 is connected.
01 is engaged. In this embodiment, the shift control valve 71 is provided with an oil chamber 717 and an oil chamber 716 that feed back the output oil pressure of the oil passage 13 and the oil passage 14, thereby mitigating the rise of the output oil pressure and controlling the multi-disc clutch 4.
01 and multi-disc brake 407 are engaged. The reduction ratio control mechanism 80 of the present invention includes a reduction ratio control valve 81, orifices 82 and 83, an upshift electromagnetic solenoid valve 84, and a downshift electromagnetic solenoid valve 85. The reduction ratio control valve 81 has lands 812A and 812B, and one land 81
Spool 8 with spring 811 attached to 2B
12, oil chambers 81 at both ends to which line pressure is supplied from oil passage 1 via orifices 82 and 83, respectively;
5 and 816, an input port 817 that communicates with the oil path 1 to which line pressure is supplied and whose opening area increases or decreases according to the movement of the spool 812, and a hydraulic servo C of the input pulley A of the V-belt continuously variable transmission 200. Output port 818 communicating with via oil line 2
An oil chamber 819 is provided with an oil chamber 819, and a drain port 8 drains pressure from the oil chamber 819 according to the movement of the spool 812.
14, and the oil chamber 8 according to the movement of the spool 812.
An upshift electromagnetic solenoid valve 84 and a downshift electromagnetic solenoid valve 85, which are equipped with a drain port 813 for discharging the pressure of They are operated by the output of the control circuit to evacuate the pressure in the oil chamber 815 and the oil chamber 816, respectively. FIG. 6 shows an electric control circuit for controlling the electromagnetic solenoid valve 74 of the shift control mechanism 70, the upshift electromagnetic solenoid valve 84, and the downshift electromagnetic solenoid valve 85 of the reduction ratio control mechanism 80 in the hydraulic control device shown in FIG. The configuration is shown below. 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 input pulley A; 903 is a vehicle speed sensor; 904 905 is a throttle sensor that detects the throttle opening of the engine, and 905 is a rotation speed sensor 9.
906 is a vehicle speed detection circuit that converts the output of vehicle speed sensor 903 into voltage; 907 is a throttle opening detection processing circuit that converts the output of throttle sensor 904 into voltage; 908 911 is the input interface of each sensor, 912 is the central processing unit (CPU), 913 is the electromagnetic solenoid valve 74, 84,
Read-on memory (ROM) that stores programs to control the 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; 917 is a solenoid output driver that downshifts the output of the output interface 916; The output is changed to the operating output of the solenoid valve 85, upshift electromagnetic solenoid valve 84, and shift control solenoid 74. Input interfaces 908 to 911, CPU 912, ROM 9
13, RAM914, output interface 91
6 is a data bus 918 and an address bus 91.
It is communicated with 9. Next, torque ratio valve 66, detent valve 6
4. The operation of the hydraulic pressure adjusting device 60 of this embodiment, which is composed of 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. Servo is V belt 11
It is defined by the hydraulic pressure that allows torque to be transmitted without causing slippage. The solid line in FIG. 7 shows the minimum line pressure necessary for a change in the torque ratio T between the input and output shafts when the engine is operated in a state that provides the best fuel efficiency, 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 indicated by the broken line, which is about % larger, and it is desirable to set the line pressure characteristic to be higher than the line pressure indicated 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 the torque ratio of both pulleys (torque ratio between input and output shafts) are adjusted as follows. D position Only oil passage 3 is oil passage 1 in manual valve 62
The oil passage 4 and the oil passage 5 are in communication with each other, and the oil passage 4 and the oil passage 5 are depressurized. At this time, in the shift control mechanism 70,
When the shift control electromagnetic 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 passage 13 are connected, and the oil passage 3 acts on the hydraulic servo 402 of the forward multi-disc clutch 401 through the oil passage 13,
The vehicle is now ready to move forward. (1) When the torque ratio T is t ₁ ≦T≦t ₂ . As shown in FIG. 5A, the torque ratio valve 66 is
Close port 663 connected to oil passage 1, and close oil passage 6.
is communicated with the drain port 665 to exhaust pressure. As a result, detent pressure (equal to line pressure) is not generated in the oil passage 7 regardless of the throttle opening degree θ. Further, the throttle valve 65 is connected to a port 664 of a torque ratio valve that communicates with the oil passage 9.
is closed and spool 651 is connected to land 6.
In addition to 56, the land 657 also receives feedback pressure, so that the throttle pressure having the characteristics shown in FIG.
Output to 13. As a result, the line pressure output from the regulating valve 61 becomes as shown in area H of FIG. 9 and area E of FIG. 10. (2) When the torque ratio T is t ₂ <T≦t ₃ . 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. Further, the pressure in the oil passage 6 is exhausted through the port 665. 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 657 of the spool 651.
This is expressed by the characteristic curve shown in FIG. 8D. The line pressure at this time has the characteristics shown in area R in FIG. 9 and area G in FIG. (3) When the torque ratio T is t ₃ <T≦t ₄ . As shown in FIG. 5C, the pressure in the oil passage 9 is exhausted from the drain port 666, so that the throttle pressure is expressed by D in FIG. 8 as in 2 above. However, since the port 663 opens and the oil passage 1 and the oil passage 6 communicate with each other, the throttle opening degree θ is within the range of 0≦θ≦θ ₁ %, and the spool 6 of the detent valve 64
41 is on the left side in the figure as shown in FIG.
However, the throttle opening θ is θ ₁
When %<θ≦100%, the spool 641 moves 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 has a characteristic that changes stepwise at θ=θ ₁ %, as shown in area 2 of FIG. 9 and area 1 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 the same as the D position. (1) When the torque ratio T is t ₁ ≦T≦t ₂ . 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 valve 64, and detent pressure is generated to push up the throttle plunger and generate high line pressure.
When 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. 4A, and detent pressure is not generated, and the throttle pressure is is the same as for the D position. Therefore, the line pressure has the characteristics shown in FIG. (2) When the torque ratio T is t ₂ <T≦t ₃ . The difference from the above (1) is that 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 adjustment valve 61 via the oil passage 8 increases. In particular, as a result, the line pressure is represented by a characteristic curve as shown in FIG. (3) When the torque ratio T is t ₃ <T≦t ₄ . Oil passage 6 and oil passage 1 are connected by the torque ratio valve 66.
and the oil passage 9 is connected to the drain port 666.
Pressure is being released from the 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 inputs the detent pressure and the same throttle pressure as in (2) above. The regulating valve 61 outputs the line pressure shown in FIG. R Position As shown in the table, the oil passage 4 and the oil passage 5 communicate with the oil passage 1 in the manual valve 62, and the oil passage 3 is depressurized. At this time, in the shift control mechanism 70, if the shift control electromagnetic solenoid valve 74 is in the OFF state and line pressure is being supplied to the oil chamber 713, the spool 712 is positioned to the left, so that the oil passage 4 and the oil passage 14, and the line pressure supplied to the oil passage 4 is supplied to the hydraulic servo 409 of the reverse multi-disc brake 407 through the oil passage 14.
The vehicle goes into reverse. Further, since the line pressure is guided to the oil passage 5, the line pressure has the same characteristics as in the L position. At the R position, the torque ratio T in the V-belt continuously variable transmission 200 is set to the maximum, T= _t4 . Therefore, there is no need to perform a speed change (deceleration) within the planetary gear transmission 400, but according to this embodiment, even when the torque ratio T is changed at the R position, the line pressure is maintained at the same level as at the L position. Control is possible. P position and N position Oil passages 3, 4 and 5 in manual valve 62
Both are exhausted, and since the oil passage 5 is exhausted, the line pressure, which is the output of the regulator valve 61, 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 why the throttle opening θ is set low at ₁ % or less as shown in curve 2 is because if the line pressure is set high during operating conditions such as idling where the throttle opening θ is small and the pump discharge amount is small, the oil temperature will rise. If there is a large amount of oil leaking 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. . Manual valve 62
is shifted to L and R shift positions,
Torque ratio T
is in the range of t ₁ ≦T≦t ₂ and the throttle opening θ is θ ₁
The reason why the line pressure is set high under operating conditions below % is that relatively high oil pressure is required during engine braking even when the throttle opening is low. The required oil pressure at that time is shown by a dashed line in FIG. In this way, as shown in FIG. 9, by bringing the line pressure close to the minimum required oil pressure shown in FIG. 7, the power loss caused by the pump 52 can be reduced, thereby improving fuel efficiency and fuel consumption rate. Next, the electric control circuit 90 and the electric control circuit 9
The operation of the shift control mechanism 70 controlled by 0 and the reduction ratio control mechanism 80 of the present invention will be explained with reference to the program flowcharts shown in FIGS. 18 to 26. In this embodiment, an example is shown in which the electric control circuit 90 controls the input pulley rotation speed N to achieve the best fuel efficiency at each throttle opening θ. Generally, when operating an engine in a state with the best fuel efficiency, the engine is operated according to the best fuel efficiency power line, which is a broken line shown in the graph of FIG. In FIG. 12, the horizontal axis shows the engine rotation speed (nm), the vertical axis shows the torque of the engine output shaft (Kg·m), and the best fuel efficiency power line is obtained as follows. Equal fuel consumption rate curve of the engine shown by the solid line in Figure 12 (unit: g/ps・h)
From the equal horsepower curve (unit: ps) shown by the two-dot chain line, the fuel consumption rate Q (g/ps・ps・
h) If horsepower is P (ps), then at point A, fuel will be consumed per hour by S=Q×P (g/h). By calculating the fuel consumption per hour S at all points on each equal horsepower curve, the point where S is the minimum can be determined on each equal horsepower curve, and by connecting these points, the best fuel consumption for each horsepower can be determined. The best fuel efficiency power line indicating the engine operating state is obtained. However, when the engine and the fluid coupling 21, which is a fluid coupling, are combined as in this embodiment, the engine output performance curve at each throttle opening θ shown in FIG. The best fuel consumption fluid coupling output line can be determined on the fluid coupling output performance curve as shown in FIG. 16 from the fluid coupling performance curve shown in the figure and the engine equivalent fuel consumption rate curve shown in FIG. 15. FIG. 17 shows the best fuel efficiency fluid coupling output line shown in FIG. 16 replaced by the relationship between throttle opening and fluid coupling output rotation speed. This fluid coupling output rotation speed directly becomes the input pulley rotation speed in the continuously variable transmission of this embodiment. In the continuously variable transmission of this embodiment, the gear ratio (reduction ratio) between input pulley A and output pulley B is determined based on the best fuel efficiency input pulley rotation speed obtained as described above and the detected actual input pulley rotation speed. Control. The reduction ratio control mechanism 80 controls the increase or decrease of the gear ratio between the input and output pulleys by comparing the best fuel efficiency input pulley rotation speed determined in FIG. 17 with the actual input pulley rotation speed. This is done by the action of two electromagnetic solenoid valves 84 and 85 provided at 80, and the actual input pulley rotation speed is made to match the input pulley rotation speed for the best fuel efficiency. FIG. 18 shows an overall flowchart of input pulley rotation speed control. After the throttle opening θ is read 921 by the throttle sensor 904, the shift lever position is determined 92 by the shift lever switch 901.
Do step 2. As a result of the determination, if the shift lever is in the P position or the N position, both the electromagnetic solenoid valves 84 and 85 are turned off by the P position and N position processing subroutine 930 shown in FIG.
31) Store the P or N state in RAM 914. (932) As a result, the neutral state of the input pulley A is obtained. When the shift lever changes from the P position or N position to the R position, and from the N position to the D position, in order to relieve the shift shock caused by the N-R shift and N-D shift, respectively. Shift shock control processing 940 and 950 are performed. As shown in Fig. 20, the shift shock control has ^a pulse width of L ^* −nM ^* (n=1・
The shift control mechanism 7 shown in FIG.
This is done by adding the shift control electromagnetic solenoid valve 74 of 0 to 0. (This will be referred to as duty control hereinafter.) By performing duty control on the shift control electromagnetic solenoid valve 74 in this manner, a hydraulic pressure ps adjusted in accordance with the duty is generated in the oil chamber 713 of the shift control valve 71. The shift control mechanism 70 controls the hydraulic servo 402 of the planetary gear transmission 400 by the action of an electromagnetic solenoid valve 74 controlled by the output of the electric control circuit 90.
and 409 to prevent shocks during shifting, and to keep the upper limit of the hydraulic pressure supplied to the hydraulic servos 402 and 409 below a set value by the action of the pressure limiting valve 73. This limits the engagement pressure of the clutch and brake. In this embodiment, as shown in FIG. 28, 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 ,
When the elastic force of the spring 711 is Fs ₁ and the oil pressure of the oil pressure 713 is ps, the hydraulic servo 402 of the multi-disc clutch 401 is engaged when moving forward, and the hydraulic servo 409 of the multi-disc brake 407 is engaged when moving backward.
The hydraulic pressures Pc and Pb supplied to the shift control valve 71 are given as follows from equations (1) and (2), which are oil pressure balance equations for the shift control valve 71, respectively. When moving forward Ps x S ₁ = Pc x S ₂ + Fs ₁ (1) Pc = S ₁ /S ₂ Ps - Fs ₁ /S ₂ When moving backward Ps x S ₁ = Pb x (S ₁ - S ₂ + Fs ₁ ) + Fs ₁ (2) Pb=S ₁ /S ₁ −S ₂ Ps−Fs ₁ /S ₁ −S ₂ In addition, the pressure receiving area of the valve body 731 inserted in the pressurized valve 73 is S 3 , and the pressure receiving area of the valve body 731 is S ₃ . Assuming that the elastic force of the spring 732 provided behind the spring 731 is Fs ₂ , the pressure limit valve 73 operates at the maximum pressure Plimit of Ps according to the hydraulic equilibrium equation (3). Plimit×S ₃ =Fs ₂ (3) Plimit=Fs ₂ /S ₃ At this time, the maximum pressures Pclimit and Pblimit of Pc and Pb are limited according to equations (4) and (5). When moving forward Pclimit=S ₁ /S ₂ Plimit−sFs ₁ /S ₂ (4) When moving backward Pblimit=S ₁ /S ₁ −S ₂ Plimit−Fs ₁ /S ₁ −S ₂ (5
) FIG. 22 shows a program flowchart when control is performed using the parameters K ^* , L ^* , and M ^* of the waveform diagram shown in FIG. 10. FLUG is determined 941 to determine whether shock control processing is in progress, and if processing is in progress, the processing is continued; if processing is not in progress, shift lever switch 901 changes from P position or N position to R position. 942 and 943 to determine whether there is a change from the N position to the D position. If any change has occurred, the corresponding parameters of K ^* , L ^* , and M ^* are set 944 or 945, and turns on FLUG, which indicates that the shift shock control process is to be performed (955). Returns if no change has occurred,
No shift shock control processing is performed.
The shock control determines whether or not the parameter K, which determines the end of one cycle K ^*, is a positive value 946.
, and when K is not a positive value, let K be ^* and L ^* be L ^* −
Set M ^* and L as L ^* (947), determine whether L is less than or equal to 0 (948), and if L is less than or equal to 0,
FLUGOFF949 and return. In this case, L≦0 and turning OFF FLUG indicates that all shift shock control processing has been completed. When the parameter K for determining the end of one cycle K ^* is a positive value in determination 946, K-1 is set as K (950);
In addition to determining whether L≦0 in determination 948, determination 951 is performed to determine whether the parameter L that determines the end of the ON time in one cycle K is L=0. L
When = 0, OFF command 95 for solenoid valve 74
2, and when L is other than 0, the solenoid valve 74
After issuing the ON command 953, L-1 is set to L (954) and the process returns. Similar shift shock control processing can also be performed using a programmable timer shown at 920 in FIG. N-D shift shock control processing 950
Next, the input pulley rotation speed sensor 902
Read the actual input pulley rotation speed N by
Do step 3. Next, it is determined 924 whether the throttle opening degree θ is 0 or not, and if θ≠0, the throttle opening degree θ is set as data in advance according to the subroutine shown in FIG.
The best fuel efficiency input pulley rotation speed corresponding to the throttle opening degree θ in Fig. 17 stored in the ROM913
In order to set N ^* (960), set 961 the function address of the input pulley rotation speed N ^* data corresponding to the throttle opening, and then start from the set address.
Read data of N ^* (962) Read N ^*
The data is temporarily stored in the data storage RAM 914 (963). Next, a comparison 927 is made between the actual input pulley rotation speed N and the best fuel economy input pulley rotation speed N ^* . N<
When N ^* , upshift electromagnetic solenoid valve 84
When N>N ^* , an operation command 928 is issued for the downshift electromagnetic solenoid valve 85.
When N=N ^* , both electromagnetic solenoid valves 8
4 and 85 OFF commands 920 are issued. θ=
When the throttle is fully closed at 0, a determination 926 is made as to whether the shift lever is set at the D position or the L position in order to determine the necessity of engine braking, and engine brake control 970 or engine brake control is performed as necessary. Do 980. As shown in FIG.
71 and calculate the acceleration α at that point (97
2) Next, it is determined 973 whether the acceleration α is an appropriate acceleration A for the vehicle speed. α＞A
In order to perform DOWNSHIFT control 974, set N ^* to a value larger than N, then
When α≦A, the best fuel consumption input pulley rotation speed N ^* corresponding to the throttle opening degree of 0 is set to N ^* (975), and then the process returns. The relationship between vehicle speed and appropriate acceleration A is determined by experiment or calculation for each vehicle, and is shown in the graph of FIG. 25. In the engine brake processing 980 for the L position, as shown in FIG.
is calculated from the following equation. (982) T=N/V×k k is the reduction gear mechanism 5 inside the transmission
This is a constant determined from the 00 reduction ratio, the final reduction ratio of the vehicle, the tire radius, etc. Next, a determination 983 is made as to whether or not the current torque ratio T is larger than the torque ratio T ^* that provides safe and appropriate engine braking for the vehicle speed V, and if T<T ^* , DOWNSHIFT is performed. N ^* is set 984 to a value larger than N, and when T≧T ^* , N ^*
A value equal to N is set 985 and the process returns. The torque ratio T ^* that provides safe and appropriate engine braking for each vehicle speed is determined by experiment or calculation for each vehicle, and is shown in the graph of FIG. 27. The shift control mechanism 70 controls the hydraulic servo 402 of the planetary gear transmission 400 by the action of an electromagnetic solenoid valve 74 controlled by the output of the electric control circuit 90.
and 409 to prevent shocks during shifting, and to keep the upper limit of the hydraulic pressure supplied to the hydraulic servos 402 and 409 below a set value by the action of the pressure limiting valve 73. This limits the engagement pressure of the clutch and brake. When loosening the engagement shock during the N-D shift and N-R shift, the rise of the hydraulic pressure Pc or Pb supplied to the hydraulic servo 402 or the hydraulic servo 409 is controlled as shown in the hydraulic characteristic curve shown in FIG. Multi-disc clutch 401 between AC in the figure
Alternatively, the engagement of the multi-disc brake 407 is completed. FIG. 27 shows the relationship between the duty (%) of the electromagnetic solenoid valve 74 for controlling the hydraulic pressure supplied to the hydraulic servo 402 or 409 and the solenoid pressure Ps generated in the oil chamber 713 by the operation of the electromagnetic solenoid valve 74. Shown below. Duty (%) is given by the following formula. Duty =
Solenoid ON time in one cycle/Solenoid operating cycle × 100 (%) The solenoid pressure shown in Fig. 30 is the shift control valve 7
The hydraulic servo 4 shown in FIG.
The hydraulic pressure Pc or Pb supplied to 02 or 409 is obtained. The operation of the reduction ratio control mechanism 80 according to the present invention will be explained with reference to FIG. 32. When traveling at constant speed, electromagnetic solenoid valves 84 and 8 are controlled by the output of the electric control circuit 90 as shown in FIG. 32A.
5 is turned off. As a result, the oil pressure P ₁ in the oil chamber 816 becomes line pressure, and the oil pressure P ₂ in the oil chamber 815 becomes the line pressure.
When the spool 812 is on the right side in the figure, the pressure is line pressure. Spool 812 is spring 8
Since there is a pressing force P ₃ due to the spring load of 11, it is moved to the left in the figure. The spool 812 is moved to the left and communicates with the oil chamber 815 and the drain port 813, and the pressure _P2 is exhausted, so the spool 812 is moved to the right in the figure by the oil pressure _P1 of the oil chamber 816.
When spool 812 is moved to the right, drain port 813 is closed. Yotsute spool 812
In this case, by providing a flat plane (tapered surface) 812b at the edge of the drain port 813 of the land 812B of the spool 812,
In a more stable state, move the spool 812 to Figure 32A.
It is possible to maintain an equilibrium point at an intermediate position as shown in FIG. When the oil passage 2 is maintained at the intermediate equilibrium point as shown in FIG. 32A, the oil passage 2 is closed, and the oil pressure of the input pulley A hydraulic servo C
It is compressed by the hydraulic servo E via the V-belt 112, and as a result, it is balanced with the hydraulic pressure of the hydraulic servo E. Actually, since there is oil leakage in the oil passage 2 as well, the input pulley A is gradually expanded and the torque ratio T changes in the direction of increasing. Therefore, in the position where the spool 812 is in equilibrium as shown in FIG. (Tapered surface) 812a is provided to compensate for oil leakage in the oil passage 2. Furthermore, the drain port 81 of land 812A
4-edge flat surface (tapered surface) 81
By providing 2C, changes such as the rise of the oil pressure change in the oil passage 2 can be made smooth. Also, as shown in FIG. 34, an oil passage 822 having an orifice 821 between oil passage 1 and oil passage 2 instead of the tapered surface 812a
It is clear that the same function can be achieved by contacting with . During UP-SHIFT, the up-shift electromagnetic solenoid valve 84 is turned ON by the output of the electric control circuit 90 as shown in Fig. 32C.
be done. As a result, the pressure in the oil chamber 815 is exhausted, and the spool 812 is moved to the right in the drawing, the spring 811 is compressed, and the spool 812 is set to the right end in the drawing. In this state, the line pressure of oil passage 1 is at port 818.
Hydraulic servo 31
The oil pressure at No. 3 increases, the input pulley A operates in the direction of closing, and the torque ratio T decreases. Therefore, by controlling the ON time of the solenoid valve 84 as necessary, the upshift is performed by reducing the torque ratio by a desired amount. At the time of DOWN-SHIFT, the solenoid valve 85 is turned on by the output of the electric control circuit 90 as shown in FIG. 32B, and the oil chamber 816
is exhausted. The spool 812 is the spring 81
1 and the line pressure of the oil chamber 815, the oil passage 2 is communicated with the drain port 813 and the pressure is exhausted, and the input pulley A
operates in the direction of rapid expansion, and the torque ratio T increases. By controlling the ON time of the solenoid valve 85 in this manner, the torque ratio is increased and a downshift is performed. In this way, the hydraulic servo C of the input (drive side) pulley A is supplied with the output hydraulic pressure of the reduction ratio control valve 81, and the line pressure is guided to the hydraulic servo E of the output (driven side) pulley B. Letting Pi be the oil pressure of hydraulic servo C of pulley A, and Po of oil pressure of hydraulic servo E of output pulley B, Po/Pi has a characteristic with respect to torque ratio T as shown in the graph of FIG. 33,
For example, throttle opening θ = 50%, torque ratio T =
1.5 (point a in the figure) and then release the accelerator and set θ = 30%. If Po/Pi is maintained as it is, the driving state will be as shown at point b in the figure with torque ratio T = 0.87. and vice versa, torque ratio T = 1.5
When maintaining the state, the value of Po/Pi is increased by the output of the reduction ratio control mechanism 80 that controls the input pulley and changed to the value at point C in the figure. Like this Po/
By controlling the value of Pi as necessary, any torque ratio can be set to correspond to any load condition. FIG. 35 shows still another embodiment of the reduction ratio control mechanism for a continuously variable automatic transmission for a vehicle according to the present invention.
In this embodiment, the port 817 of the reduction ratio control valve 81,
Drain port 813 and drain port 81
V-shaped 817a, 813a, and 814 in 4
In the previous embodiment, the land edges of the spool 812 function as valves 812b, 812a, and 812c, thereby obtaining the same effect as in the previous embodiment. FIG. 36 shows still another embodiment of the reduction ratio control mechanism for a continuously variable automatic transmission for a vehicle according to the present invention. In this embodiment, the spool 8 of the reduction ratio control valve 81
12 three lands 812A, 812B, 812
A spring 811 is provided behind the land 812C, and an oil chamber 810 is provided between the lands 812B and 812C to communicate with the oil passage 1 through an orifice 82 and with a drain port 813 as the spool moves. , when the spring 811 is installed, the upshift electromagnetic solenoid valve 8
The oil chamber 815 in which the oil chamber 4 is provided and the oil chamber 810 are connected through an oil passage 1A. In this embodiment, in the reduction ratio control mechanism of a continuously variable automatic transmission for vehicles, the only place where line pressure leaks during upshifts and downshifts is the electromagnetic solenoid valve, which allows the line pressure to be maintained high. When it is desired to quickly complete a downshift to the point of maximum torque ratio, such as during a sudden stop, a higher line pressure can be supplied to the output pulley's hydraulic servo, which has the advantage of increasing downshift speed. Furthermore, as in this embodiment, the lands 812A and 812C at both ends of the spool do not intersect with the port at the edges on both ends, so interference between the edges of the lands and the edges of the port can be prevented, and interference due to eccentricity of the spool can be prevented. There is less occurrence of valve stickiness. As described above, the reduction ratio control mechanism of the continuously variable automatic transmission for vehicles of the present invention includes the hydraulic servo of the input pulley, the reduction ratio control valve that communicates with the hydraulic pressure source or oil drain, and the spool of the reduction ratio control valve. In a reduction ratio control mechanism consisting of a spring that applies a spring load, and an upshift electromagnetic solenoid valve and a downshift electromagnetic solenoid valve that discharge hydraulic pressure applied from both of the spools according to vehicle running conditions, the spring load due to the spring is , since it is applied in the direction in which the spool moves during downshifting, a response of the reduction ratio control valve sufficient for rapid downshifting can be obtained.

[Brief explanation of drawings]

Fig. 1 is a sectional view of a continuously variable automatic transmission for a vehicle, Fig. 2 is a hydraulic circuit diagram of a hydraulic control device of the continuously variable automatic transmission, Fig. 3 is an explanatory diagram of the operating state of the manual valve, and Fig. 4 is a diagram of the hydraulic control device of the continuously variable automatic transmission. Fig. 5 is an explanatory diagram of the operating state of the detent valve and throttle valve, Fig. 5 is an explanatory diagram of the operating state of the torque ratio valve, Fig. 6 is a configuration diagram of the electric control circuit, and Fig. 7 is a graph showing the required line pressure characteristics of the hydraulic control device. , FIG. 8 is a graph showing the characteristics of the throttle pressure, FIGS. 9, 10, and 11 are graphs showing the line pressure characteristics obtained by the hydraulic pressure adjustment device.
Figure 2 is a graph showing the engine's best fuel efficiency power line.
FIG. 13 is a graph showing the output performance characteristics of the engine, FIG. 14 is a fluid transmission mechanism performance curve, FIG. 15 is an equal fuel consumption rate curve of the engine, and FIG. 16 is a graph showing the best fuel consumption fluid coupling output curve. Fig. 17 is a graph showing the characteristics of the best fuel efficiency fluid coupling output rotation speed, Figs. 18, 19, 22 to 24, and 26 are program flowcharts of the electric control circuit, and Fig. 20 21 is a waveform diagram for explaining the duty ratio, FIG. 21 is an explanatory diagram of the operation of the shift control electromagnetic solenoid valve, FIG. 25 is a graph showing the set acceleration, FIG. 27 is a graph showing the set torque ratio, and FIG. 25 is a graph showing the set torque ratio. An operating diagram of the shift control electromagnetic solenoid valve, Fig. 28 is an explanatory diagram of the operation of the shift control mechanism, Fig. 29 is a graph showing the characteristics of the oil pressure supplied to the input pulley hydraulic servo and output pulley hydraulic servo, and Fig. 30 is the solenoid. A graph showing the characteristics of the pressure Ps, Fig. 31 is a graph showing the characteristics of the output oil pressure of the shift control valve, Fig. 32 is an operating state diagram of the reduction ratio control mechanism of the present invention, and Fig. 33 is a V-belt type continuously variable transmission. Graph showing the relationship between the torque ratio T between the input and output shafts of the machine and the pressure ratio between the input and output side hydraulic servos, No. 34
FIG. 35 is a block diagram showing another embodiment of the reduction ratio control mechanism of the present invention, FIG. 35 is a block diagram of still another embodiment of the present invention, and FIG. 36 is a block diagram of still another embodiment of the present invention. , FIG. 37 is a configuration diagram of a conventional reduction ratio control mechanism. In the diagram, 200...V-belt continuously variable automatic transmission,
A...Input pulley, B...Output pulley, 112...
...V belt, C...Input side hydraulic servo, E...Output side hydraulic servo, 400...Planetary gear transmission for forward/reverse switching, 60...Hydraulic control circuit, 50...
... Hydraulic power source, 52 ... Pump, 61 ... Regulator valve, 613 ... Regulator valve plunger, 62 ... Manual valve, 64 ... Detent valve, 65 ... Throttle valve, 66 ... Torque ratio valve, 70... Shift control mechanism, 71... Shift control valve, 72... Orifice, 73... Pressure limiting valve, 74... Shift electromagnetic control solenoid valve, 80... Reduction ratio control device, 81...
Reduction ratio control valve, 82, 83... Orifice, 84
...Solenoid solenoid valve for upshift, 85...
Electromagnetic solenoid valve for downshift.

Claims (1)

[Claims] 1. A continuously variable automatic transmission for a vehicle, which includes an input pulley and an output pulley, each of which has an effective diameter variable by a hydraulic servo, and a V-belt that transmits power between these pulleys. A reduction ratio control mechanism that is provided in a hydraulic control device and controls a reduction ratio of the continuously variable automatic transmission by supplying and discharging hydraulic oil to and from a hydraulic servo of the input pulley; , a reduction ratio control valve that communicates with a hydraulic pressure source or an oil drain port, a spring that applies a spring load to the spool of the reduction ratio control valve, and a system that discharges the hydraulic pressure applied from both of the spools according to vehicle running conditions. A continuously variable automatic transmission for a vehicle, wherein the reduction ratio control mechanism includes an upshift electromagnetic solenoid valve and a downshift electromagnetic solenoid valve, wherein the spring load by the spring is applied in the direction in which the spool moves during downshifting. reduction ratio control mechanism.