JPH0231264B2 - SHARYOYOMUDANJIDOHENSOKUKINORAINATSUSEIGYOKIKO - Google Patents

Description

[Detailed description of the invention]

(Industrial Application Field) The present invention is provided in a hydraulic control device of a continuously variable automatic transmission for a vehicle using a V-belt type continuously variable transmission,
The present invention relates to a line pressure control mechanism for controlling a hydraulic pressure adjustment device that outputs line pressure required by a continuously variable automatic transmission for a vehicle. (Prior Art) In a continuously variable automatic transmission for a vehicle formed by combining a V-belt type continuously variable transmission, a planetary gear transmission for forward/reverse switching, and a hydraulic torque converter, there is a gap between two sheaves. A belt having a V-shaped or trapezoidal cross section is stretched over the two sheaves, and the radial position where the belt contacts the flange of each sheave is changed by controlling the hydraulic pressure supplied to the hydraulic servo. The rotational speed ratio between the axes can be changed steplessly. Since the output oil pressure of this hydraulic control device is generated from an oil pump driven by an engine, it is desirable that the output oil pressure be as low as possible in order to improve the fuel efficiency of the vehicle. The hydraulic pressure required for continuously variable automatic transmissions for vehicles is
It changes depending on the driving conditions of the vehicle. For example, a V-belt type continuously variable transmission requires a hydraulic pressure that does not cause the V-belt to slip during transmission. Therefore, by bringing the output oil pressure of the hydraulic control device closer to the minimum required oil pressure, the discharge pressure of the oil pump is lowered, and the consumption of engine output by the oil pump is reduced, thereby improving fuel efficiency. becomes possible. (Problem to be Solved by the Invention) However, in the control device for the continuously variable automatic transmission for vehicles having the above configuration, when the rotational speed of the V-belt type continuously variable transmission increases, the Centrifugal force acts on the oil, resulting in excessive clamping force. Furthermore, if the centrifugal force becomes excessive, it acts as a force that clamps the V-belt and generates more tension than necessary in the V-belt, which may reduce the durability of the V-belt. An object of the present invention is to solve the problems of the conventional control device for a continuously variable automatic transmission for vehicles, and to solve the problem of the control device of the conventional continuously variable transmission for vehicles. The present invention provides a control device for a continuously variable automatic transmission for a vehicle that does not apply the above clamping force to the V-belt. (Means for Solving the Problems) To this end, the line pressure control mechanism of the continuously variable automatic transmission for vehicles of the present invention includes drive sheaves whose effective diameters are variable by hydraulic servos each having approximately the same pressure receiving area. and a driven sheave, a continuously variable transmission consisting of a V-belt that transmits power between both sheaves, and a planetary gear transmission that switches between forward and reverse movement using a frictional engagement element that is engaged or released by a hydraulic servo. In a hydraulic control device for a continuously variable automatic transmission for vehicles that is combined with a switching mechanism, the pressure of the pressure oil supplied from the hydraulic source is adjusted according to the input, and the pressure is adjusted according to the vehicle running conditions such as vehicle speed and throttle opening. , a first regulator valve that outputs as a first line pressure, and a second regulator that adjusts the pressure of surplus pressure oil of the first regulator valve according to vehicle running conditions and outputs it as a second line pressure lower than the first line pressure. a hydraulic pressure adjustment device including a valve; a line pressure control mechanism for inputting vehicle speed and throttle opening to produce an output oil pressure for controlling the hydraulic pressure adjustment device; and a line pressure control mechanism for controlling the hydraulic pressure control device according to vehicle running conditions. Controlled by an electronic control device, line pressure is supplied directly or after pressure adjustment to each hydraulic servo of the drive sheave and driven sheave, or the pressure is discharged from these hydraulic servos to change the reduction ratio between the two sheaves. It has a hydraulic control device consisting of a reduction ratio control device. The vehicle also includes a throttle valve that generates an output hydraulic pressure to the hydraulic pressure adjustment device, a cutback valve that generates a cutback pressure that controls the output hydraulic pressure of the throttle valve, and a valve that generates a governor pressure that controls the cutback valve in accordance with the vehicle speed. The cutback valve adjusts the output oil pressure of the throttle valve according to the output oil pressure of the governor valve and feeds it back to the throttle valve, and the throttle valve adjusts the output oil pressure of the throttle valve according to the output oil pressure of the governor valve and feeds it back to the throttle valve. Controls two regulator valves. Furthermore, when the throttle opening is below a set value and a downshift is being performed, the low cutback valve outputs an output to the first regulator valve to raise the first line pressure. (Operations and Effects of the Invention) According to the present invention, as described above, the throttle valve generates the output hydraulic pressure to the hydraulic pressure adjusting device, the cutback valve generates the cutback pressure to control the output hydraulic pressure of the throttle valve, and the throttle valve generates the cutback pressure according to the vehicle speed. a valve for generating governor pressure to control the cutback valve;
The cutback valve regulates the output oil pressure of the throttle valve according to the output oil pressure of the governor valve and provides feedback to the throttle valve, and the throttle valve adjusts the output oil pressure of the throttle valve according to the throttle opening degree and the vehicle speed.
Since the regulator valve and the second regulator valve are controlled, the governor valve detects a change in the output side rotation speed of the V-belt type continuously variable transmission, and adjusts the belt clamping force in accordance with the output side rotation speed. The hydraulic pressure inside the annular oil chamber increases due to centrifugal force, which acts as a force to clamp the V-belt, eliminating the need for more tension on the V-belt and improving the durability of the V-belt. . In addition, since it is equipped with a low-cut back valve,
When the throttle opening is below a set value and a downshift is being performed, the output can be output to the first regulator valve to raise the level of the first line pressure. (Example) Hereinafter, an example of the present invention will be described in detail with reference to the drawings. FIG. 1 shows a cross section of an automatic continuously variable transmission for a vehicle controlled by the control device of the present invention. In the figure, the automatic continuously variable transmission 100 for a vehicle is
A V-belt continuously variable transmission 200, a torque converter 300 connected to the input side of the V-belt continuously variable transmission 200, and a V-belt continuously variable transmission 200.
A forward/reverse switching planetary gear mechanism 400 connected to the output side of the forward/reverse switching planetary gear mechanism 400; and a reduction gear mechanism 500 connected to the output side of the forward/reverse switching planetary gear mechanism 400.
and a differential gear 600 connected to the reduction gear mechanism 500. 1 is an input shaft of the continuously variable transmission 100 connected to the engine output shaft; 2 is a V-belt continuously variable transmission 20;
The input shaft 1 and the first intermediate shaft 2 are connected via a torque converter 300. 3 is the above continuously variable transmission 10
0 is an output shaft, and 4 is a tubular second intermediate shaft which is disposed coaxially on the outside of the output shaft 3 and is the output shaft of the V-belt continuously variable transmission 200, and the second intermediate shaft 4 and the output shaft are 3 is a planetary gear mechanism 400 for forward/reverse switching;
It is connected via the third intermediate shaft 5, reduction gear mechanism 500, and differential gear 600. 6,
Reference numeral 7 designates a movable flange that is slidably fitted to the first intermediate shaft 2 and the second intermediate shaft 4, and has tubular bearing portions 6A and 7A along the central axes 2 and 4, respectively. A first cylinder 8, which is provided concentrically on each intermediate shaft 2, 4 with flanges 6, 7 as side walls, is welded together, and a second cylinder 9 is integrally formed. 10 and 11 are fixed flanges integrally formed with the first intermediate shaft 2 and the second intermediate shaft 4, respectively, and the movable flange 6 and the fixed flange 10, and the movable flange 7 and the fixed flange 11.
define V-shaped spaces 13 and 14 that respectively receive the V-belt 12, and constitute sheaves A and B. 15 and 16 are first fixed walls inserted into the first cylinders 8 and 9, respectively, and have flange portions 15A and 16A in contact with the inner walls of the cylinders 8 and 9;
It has tubular parts 15B, 16B continuous to A, and fixing parts 15C, 16C continuous to the tubular part 16B and fixed to the respective intermediate shafts 2, 4. First fixed walls 15 and 16 form first annular oil chambers 17 and 18 between movable flanges 6 and 7, which are side walls of cylinders 8 and 9, respectively. 19,2
0 is a second cylinder formed integrally with the second cylinders 21 and 22 which are fitted onto the first cylinders 8 and 9, respectively.
is a fixed wall, and the fixed part 15C of the first fixed wall,
16C and is fixed to the intermediate shafts 2 and 4. Further, the distal end portion (right side in the figure) of the second cylinder 21 forms a flange-shaped portion 21A bent in the outer radial direction, and teeth 21B are formed on the outer peripheral portion of the flange-shaped portion 21A. Reference numeral 23 denotes an electromagnetic pickup inserted at a predetermined position corresponding to the flange-shaped portion 21A of the automatic transmission case 700. The electromagnetic pickup 23 and the flange-shaped portion 21A constitute a device for detecting the rotational speed of the drive sheave, that is, the rotational speed of the first intermediate shaft 2. Annular pressure receiving plates 24, 25 are slidably inserted between the second cylinders 21, 22 and the tubular portions 15B, 16B of the first fixed walls 15, 16, and Fixed walls 19, 20 and the pressure receiving plate 2
A second annular oil chamber 26, 27 is formed between the oil chambers 4, 25. 28 and 29 are spheres inserted into axial grooves formed on both the sliding surfaces between the first intermediate shaft 2 and the movable flange 6 and between the second intermediate shaft 4 and the movable flange 7, respectively. , functions to prevent relative rotation between the movable flanges 6, 7 and the intermediate shafts 2, 4. Here, oil pressure correction chambers 24', 25' are formed between the pressure receiving plates 24, 25 and the first fixed walls 15, 16, and in the oil pressure correction chambers 24', 25', an annular oil chamber 26, The oil in 27 is allowed to enter. Therefore, when the oil pressure in the annular oil chambers 17, 18, 26, 27 becomes higher than the centrifugal force as the rotation of the sheaves A and B increases, oil enters the oil pressure correction chambers 26, 27 and the oil pressure is reduced. Suppress the rise and balance it. Therefore, the clamping force of the belt 12 remains constant without being affected by centrifugal force. The V-belt type continuously variable transmission 200 has a V-belt 1
2. Consists of sheaves A and B and hydraulic servos C and D for sheaves A and B. This V-belt type continuously variable transmission 200 receives detection information from the drive sheave rotation speed detection device, vehicle speed, throttle opening sensor, etc., and operates the first oil chambers 17, 1.
By controlling the hydraulic pressure supplied to hydraulic servos C and D having 8 and second oil chambers 26 and 27, the width of the V-shaped spaces 13 and 14 is increased or decreased.
Along with this, the V-belt 12 that comes into contact with sheaves A and B
The turning radius of the wheel increases or decreases to provide stepless speed change according to the driving conditions of the vehicle. The hydraulic torque converter 300 includes a pump impeller 301, a turbine line 302, and a stator 3.
03, it has a known configuration consisting of a one-way clutch 304. In the forward/reverse switching planetary gear mechanism 400,
The second intermediate shaft 4 is the output shaft of the continuously variable transmission 200
is a drum 403 with a cylinder 402A of a hydraulic servo 402 that actuates a multi-disc clutch 401;
The third intermediate shaft 5, which is the output shaft of the planetary gear mechanism 400 for forward/reverse switching, is connected to the sun gear 405 via the drum 403 and the multi-disc clutch 401 through spline fitting. are connected at the same time. A planetary carrier 408 is disposed between the sun gear 405 and the ring gear 404, and is engaged with the automatic transmission case 700 via a multi-disc brake 407. A hydraulic servo 402 for operating the multi-disc clutch 401 and a hydraulic servo 409 for operating the multi-disc brake 407 are provided. In the forward/reverse switching planetary gear 400, when the multi-plate clutch 401 is engaged and the multi-disc brake 407 is released, a forward gear with a reduction ratio of 1 is obtained; When engaged, it becomes a reverse gear with a reduction ratio of 0.7.
This reduction ratio during reverse movement of 0.7 is smaller than the reduction ratio during reverse movement of a normal automatic transmission, but the reduction ratio (for example, 2.4) obtained in the V-belt type continuously variable transmission 200 in this embodiment, Reduction gear mechanism 50
With the reduction ratio obtained at 0, an appropriate reduction ratio can be obtained as a whole. The reduction gear mechanism 500 is intended to compensate for the fact that the speed change range obtained by the V-belt continuously variable transmission 200 is lower than the speed change range achieved by a normal vehicle transmission. A 1.45 speed change is performed to increase torque. The differential gear 600 is connected to the output shaft 3 and performs final deceleration of 3.727:1. FIG. 2 is a hydraulic circuit diagram of a hydraulic control device for an automatic continuously variable transmission for a vehicle, showing an embodiment of the present invention. The hydraulic control device of the present invention includes a hydraulic pressure generation source 50 powered by an engine, and adjusts the hydraulic pressure supplied from the hydraulic pressure generation source 50 according to vehicle running conditions such as throttle opening and vehicle speed, and outputs it as line pressure. A reduction ratio control mechanism 70 that controls the reduction ratio of the V-belt continuously variable transmission 200, a shift sequence valve 78, and a hydraulic adjustment device 60 that adjusts the throttle opening and vehicle speed according to vehicle running conditions. A line pressure control mechanism 80 that outputs and controls hydraulic pressure, and a manual valve 9 that manually switches between forward and reverse movement.
Consists of 0. The oil pressure generation source 50 is driven by an oil pump 51 driven by an engine, and is driven by an oil pump 51 driven by an engine.
2 from the oil pan 53, and discharges the oil to the oil passage 31 with the relief valve 54 at a discharge pressure according to the engine speed. The hydraulic pressure adjustment device 60 includes a first regulator valve 61 and a second regulator valve 65. The first regulator valve 61 has a pressure regulating oil chamber 610 that communicates with the oil passage 31 . The oil pressure of the oil passage 31 is supplied to the illustrated upper end oil chamber 611 via the orifice 64 and is applied to the upper end land 621 of the spool 62. Further, the oil pressure of the oil passage 32 communicating with the oil passage 31 via the orifice 75 is applied to the land 622 having a larger outer diameter than the upper land 621 . A spring 612 is attached to the lower end land 623.
is placed on the back. A plunger 63 disposed opposite the spool 62 receives the second output oil pressure of the line pressure control mechanism 80 from the lower end oil chamber 613, and presses the spool 62 upward in the drawing. The oil pressure of the oil passage 31 applied to the upper end land 621 and the oil pressure of the oil passage 32 applied to the land 622 are balanced with the pressing force of the spring 612 and the plunger 63 installed on the back, so that the spool 62 moves up and down, and the spool 62 moves up and down. The communication annular opening area between the oil chamber 610 formed by the land 624 and the valve wall 615 and the excess oil discharge port 616 is increased or decreased, and the oil pressure in the oil passage 31 is adjusted to the first line pressure related to the vehicle speed, throttle opening, etc. be pressured. Next, the second regulator valve 65 has a pressure regulating chamber 650 in the oil passage 33 that communicates with the excess oil discharge port 616 of the first regulator valve 61 . Spool 6
The oil pressure of the oil passage 33 is supplied to the upper end oil chamber 651 of the upper end land 6 through the orifice 68.
61. In addition, the lower end land 662 has
A spring 652 is provided behind it. A plunger 67 disposed opposite the spool 66 has a lower end oil chamber 6
63 receives the first output oil pressure of the line pressure control mechanism 80, and presses the spool 66 upward in the drawing.
The hydraulic pressure of the oil passage 33 applied to the upper end land 661 and the pressing force of the spring 652 and the plunger 67 installed on the back are in balance, and the spool 66 moves up and down, and the spool 66 is formed by the intermediate land 664 and the valve wall 655. The communication annular opening area between the oil chamber 650 and the excess oil discharge port 656 or drain port 657 is increased or decreased, and the oil pressure in the oil passage 33 is regulated to a second line pressure related to vehicle speed, throttle opening, etc. The surplus oil discharged from the surplus oil discharge port 656 is transferred from the oil passage 34 provided with the cooler bypass valve 55 to the torque converter 1 and the oil cooler 5.
6 and parts that require lubrication, and the drain from the drain port 657 flows out to the suction side of the oil pump 51. In addition, in the reduction ratio control mechanism 70, a drain port 7 is attached to the upper end land 721 of the spool 72.
An inclined surface 71a for adjusting the area of communication with 10 is provided, and a spring 711 is provided on the back, thereby forming a reduction ratio control valve 71. The reduction ratio control valve 7
1, the downshift solenoid valve 73 controls the oil pressure in the lower end oil chamber 713, so the upshift solenoid valve 7
4 is provided. The spool 72 is connected to a land 712 from an oil chamber 712 connected to the oil passage 33 via an orifice 76.
The lower end land 722 receives a downward force due to the hydraulic pressure and the spring load of the spring 711, and from the oil chamber 713 connected to the oil passage 33 via the orifice 77.
controlled by the equilibrium with the upward force exerted on the
Then, the opening area where the first oil chamber 724 communicating with the intermediate output port 715 communicates with the first line pressure supply port 717 and the drain port 714 is adjusted,
Hydraulic pressure is output from the output port 715 to the drive sheave hydraulic servo D of the V-belt continuously variable transmission via the oil path 35. At this time, from the oil passage 32 to the port 71
Drain port 71 for the hydraulic pressure supplied through 8
4 to regulate the oil pressure in the oil passage 32. In the shift sequence valve 78, a spring 781 is installed behind the left end land 791 of the spool 79, and the oil pressure of the oil passage 35 outputted by the reduction ratio control mechanism 70 that controls the torque ratio is applied to the right end land 792 in the figure. An intermediate land 793 is provided at the center of the spool 79. Then, the spool 79 is moved in balance between the spring load of the spring 781 and the hydraulic pressure applied to the land 792.
The oil passage 31 and the second line pressure are supplied with the first line pressure.
An oil passage 36 that switches between the oil passage 33 to which line pressure is supplied and communicates with the driven sheave hydraulic servo C of the V-belt continuously variable transmission and the manual valve 90.
At the same time, the oil passage 33 and the oil passage 37 connected to the line pressure control mechanism 80 are connected to or disconnected from each other. The line pressure control mechanism 80 includes a throttle valve 81,
It consists of a cutback valve 84, a low cutback valve 86, and a governor valve 88. The throttle valve 81 has a spool 82 and a slot plunger 83 that is disposed opposite to the spool 82 via a spring 811 and is moved according to the throttle opening. Spool 82
is provided between the left end land 821 on which the spring 811 is placed on its back, the small outer diameter land 822 at the right end, and lands 823 and 822 which are provided in the middle and have the same pressure receiving area as the land 821; , 822. An oil chamber 825 is provided between the lands 821 and 823, and the passage area with the port 812 communicating with the oil passage 31 is adjusted, and the first control pressure is outputted via the output port 814 and the oil passage 38. Also, land 82
The oil chamber 826 between 3,824 and 3,824 has an orifice 81.
The output hydraulic pressure of the oil passage 38 is fed back through the oil passage 39, and the cutback pressure output from the cutback valve 84 is supplied from the oil passage 39 to the oil chamber 827 between the lands 824 and 822. Here, when the spool 82 is moved due to the balance between the spring load from the spring 811 and the oil pressure applied to the oil chambers 826 and 827, the port 8
The opening area of 12 is adjusted and the first control pressure is output. On the other hand, the throttle plunger 83 moves according to the throttle opening degree, and outputs the output hydraulic pressure of the low cutback valve 86 supplied from the oil passage 40 from the oil passage 41. Reference numeral 89 denotes a check valve, which connects either one of the oil passages 38 and 41 with the oil passage 42, and outputs the second output oil pressure to the oil chamber 6 of the first regulator valve.
Output to 13. The cutback valve 84 also has a spool 85, which adjusts the opening area of a port 841 that communicates with the oil passage 38, and controls the opening area of the oil chamber 84 in which the port 841 is provided.
The oil pressure of 1 is regulated and the throttle valve 8 is connected from the oil passage 39.
Output to 1. The spool 85 is an orifice 84
Oil passage 39 supplied to the left end oil chamber 852 via 2
output oil pressure, right end land 854 and intermediate land 8
53 and the governor pressure supplied to the oil chamber 856. The governor pressure is determined by orifice 4.
The oil is supplied through an oil passage 44 that communicates with the oil passage 31 via a passage 3. The low cutback valve 86 includes an oil chamber 871 at the upper end that inputs the governor pressure of the oil passage 44 and an orifice 86.
The oil chamber 872 has an oil chamber 872 at the lower end to which the low cutback pressure of the oil passage 40 is fed back through the oil passage 1 . The spool 87 is moved by receiving governor pressure in a land 873 and receiving low cutback pressure in a land 874, thereby forming a pressure regulating oil chamber 875 between both lands. The spool 87 is connected to the shift sequence valve 7.
8 through the oil passage 37, the opening area of the supply boat 803 and the drain port 86
5 and outputs low cutback pressure to the oil passage 40 via the output port 867. The governor valve 88 has a known configuration, in which governor weights 881 and 882 expand in the radial direction according to the rotational speed of the output shaft of the continuously variable automatic transmission for vehicles, and the oil passage 4
The hydraulic pressure of No. 4 is adjusted according to the vehicle speed. The manual valve 90 is manually operated.
(low), D (drive), N (neutral), R
(reverse) and P (park) ranges. When shifted to the L or D range, the oil passage 36 and the oil passage 45 are communicated to supply hydraulic pressure to the hydraulic servo 402 of the multi-disc clutch 401 of the planetary gear transmission mechanism 400 for forward/reverse switching, and the multi-disc brake 402 Oil passage 46 that connects to the hydraulic servo 409 of
Exhaust pressure. Also, when in R range, oil passage 45
is discharged and the oil passages 46 and 36 are connected, and when in the N and P ranges, the oil passages 45 and 46 are connected.
are both exhausted. Next, the operation of the hydraulic control device will be explained. This is necessary to prevent slipping between each sheave A, B and the V-belt 12 in response to the input torque transmitted from the engine to the drive sheave A of the V-belt continuously variable transmission via the torque converter. Oil pressure is normally applied to the hydraulic servo C of the driven sheave B, and the required oil pressure is as shown in the characteristic curve shown in FIG. 3 with respect to the driven sheave rotation speed. In contrast, in the hydraulic control device of the present invention, the hydraulic pressure supplied to the driven sheave hydraulic servo C is controlled as follows. That is, the governor valve 88 adjusts the oil pressure in the oil passage 44 to the governor pressure shown in FIG. 4 in accordance with the output shaft rotation speed. A low cutback valve 86 and a cutback valve 84 receive the governor pressure as input.
The low cutback pressure and the cutback pressure output by the motor change as shown in FIGS. 5 and 6, respectively. On the other hand, in the throttle valve 81, the throttle plunger 83 strokes as shown in FIG. The spool 82 is controlled by the cutback pressure, and a throttle pressure whose gradient becomes discontinuous at the cutback point is outputted from the oil passage 38 as shown in FIG. The first regulator valve 61 and the second regulator valve 65 are controlled by the slot pressure. Here, when the throttle opening degree θ becomes less than the set value, the oil passage 40 communicates with the oil passage 41, the check valve 89, and the oil passage 42, and the plunger 83 outputs the low cut back pressure to the first regulator valve 61 as throttle pressure. do. FIG. 9 is a characteristic curve showing the relationship between the throttle pressure outputted through the oil passage 38 and the throttle opening degree θ. Next, in the first regulator valve 61, the throttle pressure shown in FIG. 8 is input, and the first line pressure shown in FIG. 10 is outputted via the oil passage 31. Further, the output oil pressure of the first regulator valve 61 is fed back to the lands 621 and 622. Among them, when only land 621 receives feedback pressure, the 1st stage and land 62
1,622 is the 2nd stage, which is divided and displayed in Figure 10. On the other hand, the second regulator valve 65 receives the throttle pressure shown in FIG. 8, and outputs the second line pressure shown in FIG. 11 through the oil passage 33. In the second regulator valve 65, the spring 652 and the plunger 67 are in series,
When the throttle pressure is below a certain value determined by the spring force and the pressure-receiving area of the plunger, the second line pressure becomes a constant value regardless of the throttle pressure as shown in FIG. When the throttle pressure is higher than this, the second line pressure exhibits pressure characteristics corresponding to the throttle pressure. This second line pressure is set to be higher than the required hydraulic pressure after the shift start of the hydraulic servo of the driven sheave. Next, the operating state of the reduction ratio control mechanism 70 is changed to the 12th state.
This will be explained using figures. During constant shift driving, the solenoid valves 73 and 74 controlled by the output of the electric control circuit 95 are turned off as shown in FIG. 12A. As a result, oil chamber 7
The oil pressure P ₁ in the oil chamber 712 is the line pressure, and the oil pressure P ₂ in the oil chamber 712 is also the line pressure when the spool 72 is on the right side in the figure. Here, the spool 72 is moved to the left due to the pressing force _P3 exerted by the spring 711. When the spool 72 is moved to the left and the oil chamber 712 and the drain port 710 communicate with each other, the oil pressure P ₂ is exhausted, and the spool 72 is moved to the right by the oil pressure P ₁ in the oil chamber. When the spool 72 is moved to the right, the drain port 710
will be closed. By providing a flat plane 71a at the land edge of the drain port 710 and the spool 72, the spool 72 is more stably maintained at an intermediate equilibrium point as shown in FIG. 12A. In this state, the oil passage 31 is closed, and the hydraulic pressure of the hydraulic servo D of the driving sheave A is pressurized via the V-belt 12 by the line pressure applied to the hydraulic servo C of the driven sheave B. As a result, the oil pressure of the hydraulic servo C is balanced. Actually, since there is oil leakage in the oil passage 35,
The drive sheave A is gradually expanded, and the torque ratio T changes in the increasing direction. Therefore, the 12th
As shown in Figure A, the drain port 714 closes at the position where the spool 72 is balanced, and the
However, a flat surface 71a is provided on the land edge of the spool 72 so that the spool 72 is in a slightly open state to compensate for oil leakage in the oil passage 35. Also, instead of forming the surface 71a as shown in FIG. 14, an orifice 47 is formed between the oil passage 33 and the oil passage 35.
A similar function can be achieved by communicating through an oil passage 48 having a. During upshift, the output of the electric control circuit 95 causes the solenoid valve 74 to shift as shown in FIG. 12B.
is turned on. As a result, the pressure in the oil chamber 713 is exhausted, and the spool 72 is moved to the left. As the spool 72 moves, the pressure in the oil chamber 712 and the drain port 710 is exhausted, but the action of the spring 711 causes the spool 72 to be set to the left. In this state, the line pressure of the oil passage 31 is
15 to the oil passage 35, the oil pressure of the hydraulic servo D increases, the drive sheave A operates in the direction of closing, and the torque ratio T decreases. Therefore, by controlling the ON time of the solenoid valve 74 as necessary, the upshift is performed by reducing the desired torque. In addition, when downshifting, the output of the electric control circuit 95 causes the solenoid valve 7 to move as shown in FIG. 12C.
3 is turned on, and the oil chamber 712 is evacuated. The spool 72 is moved to the right by the line pressure of the oil chamber 713, the oil passage 2 is communicated with the drain port 714 and the pressure is discharged, and the drive sheave A operates in the direction of expansion, increasing the torque ratio. In this way, solenoid valve 7
By controlling the on-time of 3, the torque ratio is increased and downshifted. In this way, the hydraulic servo D of the drive sheave A has
The output hydraulic pressure of the reduction ratio control valve 71 is supplied, and the line pressure is guided to the hydraulic servo C of the driven sheave B. If the hydraulic pressure of the hydraulic servo D is Pi and the hydraulic pressure of the output side hydraulic servo C is Po, then Po /Pi has the characteristics shown in Figure 13 with respect to the torque ratio T. For example, when the accelerator is pressed from a state where the vehicle is running with the throttle opening θ = 50% and the torque ratio T = 1.5 (point a in the diagram). When loosening and setting θ = 30%, if Po/Pi is maintained as it is, the operating state will shift to the operating state shown at point b in the figure with torque ratio T = 0.87, and conversely, if the torque ratio T = 1.5 is maintained. , the value of Po/Pi is increased by the output of the reduction ratio control mechanism 70 that controls the input sheave, and changed to the value at point c in the figure. Like this Po/
By controlling the value of Pi as necessary, it is possible to set any torque ratio corresponding to any load condition. The reduction ratio control valve 71 also regulates the first line pressure output from the first regulator valve 61 to form the oil pressure in the oil passage 32, which is one of the feedback oil passages. Oil passage 32 communicates with oil chamber 725 between land 721 and land 723 via port 718. The oil chamber 725 is evacuated from the drain port 714 during an upshift and from the drain port 710 during a downshift due to the movement of the spool 72, whereby the first regulator valve 61
Since the feedback hydraulic pressure applied to the land 622 is exhausted, the first line pressure is as shown in Figure 10.
This is the 1st stage pressure. The shift sequence valve 78 receives the hydraulic pressure supplied to the hydraulic servo D of the driving sheave as an input signal, and changes the hydraulic pressure supplied to the hydraulic servo C of the driven sheave to the first line pressure (1st line pressure shown in FIG. 10) when downshifting.
stage pressure), otherwise set to the second line pressure. Therefore, in the upshift state and no-shift state, the required capacity is secured by the second line pressure, and at the time of downshift, the required capacity is secured by the first line pressure of the 1st stage (downshift against hydraulic pressure due to centrifugal force applied to the drive sheave). possible hydraulic pressure) is ensured. Further, during a downshift, second line pressure is supplied to a low cut valve 86, which will be described later. The manual valve 90 inputs the supply pressure to the driven sheave C via the oil passage 36, and operates the multi-plate clutch of the planetary gear transmission mechanism 400 for forward/reverse switching according to the shift range of P to L as shown in Table 1 below. 401 and the multi-disc brake 407.

[Table] The second line pressure is supplied to the low-cut valve 86 via the shift sequence valve 78 only during downshifting, and is regulated by the governor pressure. Assuming that the governor pressure is P _G , the second line pressure is P ₂ , and the low-cut valve output pressure is P _M , when P _G ≧ P ₂ , P _M = P ₂ When P _G < P ₂ , P _M = P _G A low-cut output pressure P _M expressed by the relationship is output to the throttle valve plunger 83. The throttle valve plunger 83 inputs the output pressure of the low-cut valve 86 and checks the output pressure via the check valve 89 to the first regulator valve 61 only when the throttle opening is small (for example, θ≦10%).
input to the plunger 63. When the throttle opening is large (for example, 10% < θ), the output pressure of the low cut valve 86 is not output to the check valve 89, and therefore the throttle pressure is lower than the first regulator valve 61.
added to. In this way, when downshifting when the throttle opening is small, as shown in FIG. 15, the first
Line pressure is obtained. Next, FIGS. 16 to 19 show the communication state of the oil passages of the hydraulic control device when the manual valve 90 is set to the D range or the L range. Fig. 16 shows the state during an upshift, and Fig. 17 shows the state during a constant shift, both of which show when the oil passages 40 and 41 are communicated by the throttle valve plunger 83 and when they are not. It includes. FIG. 18 shows the state at the time of downshifting, in which the communication between oil passages 40 and 41 is stopped, and FIG. 19 shows the state at the time of downshifting.
0 and 41 are connected. FIG. 20 is a hydraulic circuit diagram of a hydraulic control device for a continuously variable automatic transmission for a vehicle, showing a second embodiment of the present invention. In the embodiment shown in FIG. 2, the first line pressure is supplied directly to the feedback oil passage 32 to the first regulator valve 61, whereas in this embodiment, it is supplied to the oil passage 35. . The supply pressure of the driving sheave D is fed back via the oil passage 35, and in this embodiment as well, the first regulator valve 61 generates the first line pressure shown in FIGS. 10 and 15. FIG. 21 shows a third embodiment in which the low cutback valve 86 in FIG. 2 has been removed. In this case, the hydraulic pressure supplied to the throttle valve plunger 83 during downshifting is always the second line pressure. Therefore, the throttle opening is small,
Further, during a downshift, a hydraulic pressure as shown in FIG. 22 is generated as the first line pressure. FIG. 23 shows a fourth embodiment in which, like the embodiment shown in FIG. 21, the low cutback valve 86 has been removed. The governor pressure in the oil passage 44 is then applied to the throttle plunger 67 via the shift sequence valve 78. In this case, the hydraulic pressure supplied to the throttle valve plunger 67 during downshifting is governor pressure. Therefore, when the throttle opening is small and a downshift is performed, a hydraulic pressure as shown in FIG. 22 is generated as the first line pressure. Note that the present invention is not limited to the above-mentioned embodiments, and various modifications can be made based on the spirit of the present invention, and these are not excluded from the scope of the present invention.

[Brief explanation of drawings]

FIG. 1 is a sectional view of a continuously variable transmission for a vehicle controlled by a control device of the present invention, FIG. 2 is a hydraulic circuit diagram of a control device for a continuously variable transmission for a vehicle showing an embodiment of the present invention, and FIG. is a diagram showing the relationship between driven sheave rotation speed and required oil pressure, Figure 4 is a diagram showing governor pressure characteristics,
Figure 5 is a diagram showing low cutback pressure characteristics, Figure 6 is a diagram showing low cutback pressure characteristics.
Figure 7 shows the cutback pressure characteristics, Figure 7 shows the stroke amount of the throttle valve plunger, Figure 8 shows the relationship between the throttle pressure and driven sheave rotation speed, and Figure 9 shows the relationship between the throttle pressure and the throttle opening. The relationship diagram, Figure 10 is a diagram showing the first line pressure characteristics, Figure 11 is a diagram showing the second line pressure characteristics, the first
Fig. 2 is an operating state diagram of the reduction ratio control mechanism, Fig. 13 is a diagram for explaining the operation of the reduction ratio control mechanism, Fig. 14 is a hydraulic circuit diagram showing another embodiment of the reduction ratio control device, and Fig. 15. 16 is a diagram showing the first line pressure characteristics during downshift, and FIG. 16 is an upshift state diagram of the hydraulic circuit of the second embodiment of the hydraulic control device of the present invention.
Fig. 17 is a constant shift state diagram, Fig. 18 is a downshift state diagram, Fig. 19 is another downshift state diagram, Fig. 20 is a hydraulic circuit diagram of the second embodiment of the present invention, and Fig. 21 is a hydraulic circuit diagram of the third embodiment of the present invention, FIG. 22 is a diagram showing the first line pressure characteristics during downshifting, FIG. 23 is a hydraulic circuit diagram of the fourth embodiment, and FIG. 24 is its diagram. FIG. 3 is a first line pressure characteristic diagram during downshift. 2...Input shaft (first intermediate shaft), 4...Output shaft (second
intermediate shaft), 12...V belt, 50...hydraulic pressure generation source,
51... Oil pump, 60... Hydraulic adjustment device, 61
...First regulator valve, 65...Second regulator valve, 70...Reduction ratio control mechanism, 72...Reduction ratio control valve, 73...Downshift solenoid, 74...Upshift solenoid, 78...Shift sequence valve, 80...Line pressure control Device, 81... Throttle valve, 83... Throttle plunger, 84... Cutback valve, 86... Low cutback valve, 88... Governor valve, 90... Manual valve, 200... V-belt type continuously variable transmission mechanism, 300... Fluid type torque converter, 400...Planetary gear mechanism for forward/reverse switching, A...
Drive sheave, B...Driven sheave, C, D...Sheave A
Hydraulic servo.

Claims (1)

[Scope of Claims] 1. (a) A continuously variable transmission consisting of a driving sheave and a driven sheave whose effective diameters are variable by a hydraulic servo, and a V-belt that transmits power between the two sheaves; A hydraulic control device for a continuously variable automatic transmission for a vehicle, which is combined with a forward/reverse switching mechanism having a planetary gear transmission that switches between forward and reverse by a frictional engagement element that is engaged or released, and (a) a first regulator valve that regulates the pressure of pressure oil supplied from a hydraulic source according to an input, adjusts the pressure according to vehicle running conditions such as vehicle speed and throttle opening, and outputs it as a first line pressure; a second regulator valve that regulates the pressure of excess pressure oil according to the vehicle running conditions and outputs it as a second line pressure lower than the first line pressure, and (b) inputs vehicle speed and throttle opening. , a line pressure control mechanism that generates an output hydraulic pressure that controls the hydraulic pressure adjustment device, and c) a line pressure control mechanism that is controlled by an electronic control device that controls the hydraulic control device according to vehicle running conditions, and that is controlled by an electronic control device that controls each of the drive sheave and the driven sheave. A hydraulic control device comprising a reduction ratio control device that supplies line pressure directly or in a regulated manner to hydraulic servos, or exhausts pressure from these hydraulic servos to change the reduction ratio between the two sheaves, the hydraulic pressure adjustment device The cutback valve includes a throttle valve that generates an output hydraulic pressure to the throttle valve, a cutback valve that generates a cutback pressure that controls the output hydraulic pressure of the throttle valve, and a valve that generates a governor pressure that controls the cutback valve according to the vehicle speed. The output oil pressure of the valve is regulated according to the output oil pressure of the governor valve and fed back to the throttle valve, and the throttle valve controls the first regulator valve and the second regulator valve of the oil pressure adjustment device according to the throttle opening degree and vehicle speed. A line pressure control mechanism for a continuously variable automatic transmission for vehicles, which is characterized by: 2 (a) A continuously variable transmission consisting of a driving sheave and a driven sheave whose effective diameters are variable by a hydraulic servo, and a V-belt that transmits power between the two sheaves; (b) Friction that is engaged or released by a hydraulic servo This is a hydraulic control device for a continuously variable automatic transmission for a vehicle, which is a combination of a planetary gear transmission that switches between forward and reverse modes using an engagement element, and (a) controls pressure oil supplied from a hydraulic source according to an input. A first regulator valve that regulates pressure according to vehicle running conditions such as vehicle speed and throttle opening and outputs it as a first line pressure, and adjusts excess pressure oil of the first regulator valve according to vehicle running conditions,
a second regulator valve that outputs a second line pressure that is lower than the first line pressure; and (b) a line that receives vehicle speed and throttle opening and generates an output hydraulic pressure that controls the hydraulic pressure regulator. a pressure control mechanism, and c) controlled by an electronic control device for controlling the hydraulic control device according to vehicle running conditions, and supplying line pressure directly or after adjusting the pressure to each hydraulic servo of the drive sheave and driven sheave. Or, by exhausting those hydraulic servos,
In the hydraulic pressure adjustment device comprising a reduction ratio control device that changes the reduction ratio between the two sheaves, the line pressure control mechanism controls a throttle valve that generates an output hydraulic pressure to the hydraulic pressure adjustment device, and an output hydraulic pressure of the throttle valve. a cutback valve that generates a cutback pressure that is output to the hydraulic pressure adjusting device when the throttle opening is less than a set value; and a low cutback valve that generates a low cutback pressure that is output to the hydraulic pressure adjustment device via the throttle valve when the throttle opening is less than a set value; The cutback valve adjusts the output oil pressure of the throttle valve according to the output oil pressure of the governor valve and inputs it to the throttle valve, and the throttle valve adjusts the oil pressure according to the throttle opening and vehicle speed. The first regulator valve and the second regulator valve of the adjusting device are controlled, and the low cutback valve outputs an output to the first regulator valve to level up the first line pressure when the throttle opening is less than a set value and a downshift is performed. Features a line pressure control mechanism for continuously variable automatic transmissions for vehicles.