JPH0762500B2 - Hydraulic control device for V-belt type continuously variable transmission - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION [Industrial field of application] The present invention provides a load thrust control oil chamber for controlling a load thrust required for torque transmission on one of a drive pulley and a driven pulley, and the other. The present invention relates to a hydraulic control device for a V-belt type continuously variable transmission provided with a gear ratio control oil chamber for controlling a gear ratio, and particularly to a structure of a control valve for controlling a load thrust control oil chamber.

[Prior art and its problems]

Conventionally, as a hydraulic control device for a V-belt type continuously variable transmission, as described in Japanese Patent Application Laid-Open No. 55-65755, a gear ratio control hydraulic pressure guided to a drive pulley is used to control a throttle opening and an engine speed. It is known that the transmission control valve that operates in a related manner adjusts the line pressure introduced to the driven pulley by the pressure adjustment valve that operates in the relationship between the actual transmission ratio and the engine speed. There is.

In the above-mentioned V-belt type continuously variable transmission, the line pressure introduced to the driven pulley is adjusted to give the belt tension necessary for torque transmission, but the torque transmitted to the V-belt depends on the engine-generated torque. Since it fluctuates, it is most efficient to regulate the line pressure according to the engine generated torque, and it is also preferable from the viewpoint of the life of the V belt. However, since the above pressure adjusting valve only mechanically adjusts the line pressure according to the gear ratio and the engine speed, it cannot adjust the line pressure according to the engine generated torque, resulting in excessive thrust or insufficient thrust. There is a problem that becomes.

Generally, the engine generated torque can be estimated from the throttle opening and the engine speed based on the engine performance curve, and the hydraulic pressure of the load thrust control side pulley is electronically controlled according to the estimated engine generated torque. If the pressure is adjusted at, the above problems can be solved. In this case, simply adjusting the hydraulic pressure of the load thrust control side pulley to a hydraulic pressure proportional to the engine generated torque is enough to reduce the thrust of the load thrust control side pulley when a negative engine generated torque acts, for example, during engine braking. Will be zero and the V-belt will slip. Therefore, it is necessary to maintain a predetermined minimum load hydraulic pressure so that the V-belt does not slip even when a negative engine-generated torque is applied.

In this case, the constant of proportionality between the load hydraulic pressure and the engine generated torque can be adjusted by the pressure receiving area of the spool of the load thrust control valve, and the minimum load hydraulic pressure can also be set by the spring load of the return spring of the spool. However, it is difficult to set the base point at which the load hydraulic pressure starts to rise in proportion to the engine-generated torque from the minimum load hydraulic pressure to a desired value.Therefore, when the engine-generated torque actually changes, the load thrust is excessive or There was a tendency for thrust to become insufficient.

[Object of the Invention]

The present invention has been made in view of the above problems, and an object thereof is to set the base point at which the load hydraulic pressure starts to rise in proportion to the engine generated torque from the minimum load hydraulic pressure to a desired value, and set the load hydraulic pressure to an optimum value. An object of the present invention is to provide a hydraulic control device for a V-belt type continuously variable transmission that can be easily set.

[Structure of Invention]

To achieve the above object, the present invention provides a load thrust control oil chamber for controlling a load thrust required for torque transmission on one of a drive side pulley and a driven side pulley, and controls a gear ratio on the other side. In a V-belt type continuously variable transmission provided with a gear ratio control oil chamber for controlling, a control valve for controlling the oil pressure of the load thrust control oil chamber and a signal oil pressure for controlling the control valve are provided. The control valve is provided with an electromagnetic valve for generating, and the control valve is formed between the input port to which the working pressure is input from the hydraulic pressure source, the drain port, and both ports, and outputs the hydraulic pressure to the load thrust control oil chamber. The output port, the main spool that selectively communicates the input port and drain port with the output port, and the output hydraulic pressure feedback to which the output hydraulic pressure is fed back to urge the main spool in the direction in which the output port and the drain port communicate with each other. Port and A sub spool arranged on the drain port side of the main spool, a first spring interposed between the spools and biasing the spools in opposite directions, and a second spring biasing the sub spool to the main spool side. , And a signal hydraulic pressure input port through which the signal hydraulic pressure of the solenoid valve is guided so as to bias the sub spool toward the main spool.

That is, the load thrust control valve is provided with a main spool, a sub spool that receives a signal hydraulic pressure, a first spring interposed between the spools, and a second spring that biases the sub spool toward the main spool. This achieves the object of the present invention. That is, the minimum load hydraulic pressure is determined by the spring load of the first spring, the slope of the proportional straight line between the load hydraulic pressure and the engine generated torque can be determined by the pressure receiving areas of the main and auxiliary spools, and the initial value of this proportional straight line is the second spring. It can be determined by the spring load. Therefore, the base point at which the load hydraulic pressure starts to rise from the minimum load hydraulic pressure in proportion to the engine generated torque can be set to a desired value, and the optimum load thrust corresponding to the change in the engine generated torque can be obtained.

[Explanation of Examples]

FIG. 1 shows a schematic configuration of a V-belt type continuously variable transmission according to the present invention.

In the drawing, a crankshaft 2 of an engine 1 is connected to an input shaft 5 via a flywheel 3 and a damper mechanism 4. A direct coupling clutch 6 and a rotatable direct coupling drive gear 7 are provided on the input shaft 5, and the direct coupling clutch 6 couples the direct coupling drive gear 7 to the input shaft 5 during direct coupling driving. An external gear 8 is fixed to the end of the input shaft 5,
The external gear 8 meshes with the internal gear 9 fixed to the drive shaft 11 of the continuously variable transmission 10, decelerates the power of the input shaft 5 and transmits it to the drive shaft 11.

The continuously variable transmission 10 includes a drive pulley 12 provided on a drive shaft 11,
The driven shaft 13 is composed of a driven pulley 14 and a V belt 15 wound between the pulleys. Drive pulley 12
Has a fixed sheave 12a and a movable sheave 12b, and a gear ratio control oil chamber 16 for controlling the gear ratio is provided behind the movable sheave 12b. On the other hand, the driven pulley 14 also has a fixed sheave 14a and a movable sheave 1 like the drive pulley 12.
4b, and behind the movable sheave 14b, a load thrust control oil chamber 17 for providing the V belt 15 with a load thrust necessary for torque transmission is provided. The hydraulic pressures of the gear ratio control oil chamber 16 and the load thrust control oil chamber 17 are controlled by a hydraulic control device described later.

A hollow shaft 19 is rotatably supported on the outer periphery of the driven shaft 13, and the driven shaft 13 and the hollow shaft 19 are disengaged by a starting clutch 20. The starting clutch 20 is engaged or loosely engaged when the belt is driven, and is disengaged when the direct coupling is driven. The forward gear 21 is rotatably supported on the driven shaft 13, and the reverse gear 22 is rotatably supported on the hollow shaft 19, and either the forward gear 21 or the reverse gear 22 is supported by the forward / reverse switching dog clutch 23. One is connected to the hollow shaft 19. A reverse drive idler gear 25 that meshes with the reverse drive gear 22 and another reverse drive idler gear 26 are fixed to the reverse drive idler shaft 24. Further, a reduction gear 28, which simultaneously meshes with the direct drive gear 7, the forward gear 21, and the reverse idler gear 26, and a final reduction gear 29 are fixed to the reduction shaft 27, and the final reduction gear 29 is a differential device 30. Meshes with the ring gear 31 of and transmits power to the output shaft 32.

In the continuously variable transmission with the direct coupling mechanism, the input shaft 5, the direct coupling clutch 6, the direct coupling drive gear 7, the reduction gear 28, the final reduction gear 2
9, the differential device 30, the output shaft 32 constitutes a direct drive path, and the input shaft 5, the external gear 8, the internal gear 9, the continuously variable transmission 10, the starting clutch 20, the forward gear 21, the reduction gear 2
8, final reduction gear 29, differential device 30, output shaft 3
Reference numeral 2 constitutes a continuously variable transmission path (when moving forward). And
The direct connection transmission ratio i _D between the input shaft 5 and the output shaft 32 in the direct drive path is set near the maximum speed ratio i _min between the input shaft 5 and the output shaft 32 in the continuously variable transmission path.

FIG. 2 shows a hydraulic control device, which is roughly the first pressure regulating valve 100,
Manual valve 110, second pressure regulating valve 120, direct connection control valve 130, three-way valve 140, forward / reverse switching piston 150, clutch control valve 16
0, hydraulic switching valve 180, gear ratio control valve 190, coast down control valve 200, load thrust control valve 210, controller 300, gear ratio control solenoid valve 301, load thrust control solenoid valve 302, direct connection control solenoid valve 303 It is composed of.

The oil pump 33 is the first pressure regulating valve for the oil sucked up from the oil sump 34.
It discharges to the right end port 101 and the intermediate port 102 of 100,
The spool 103 is spring 10 by the hydraulic pressure of the right end port 101.
Move to the left against 4, and when the land 103a of the spool 103 reaches the position shown in the drawing, the intermediate port 102 and drain port 1
The oil is returned to the suction side of the oil pump 33 by communicating with 05. Therefore, the spool 103 is balanced at this position, and the discharge hydraulic pressure of the oil pump 33 is adjusted to a predetermined line pressure P _L. The ports 107 and 108 provided at the corresponding positions of the intermediate port 102 are lubrication ports. The back pressure chamber 109 accommodating the spring 104 has a load thrust control valve described later.
The load hydraulic control hydraulic pressure P ₈ is guided from 210, and the line pressure P _L is adjusted to a hydraulic pressure that is balanced with the sum of the load thrust control hydraulic pressure P ₈ and the load of the spring 104. Therefore, when the load thrust control oil pressure P ₈ rises, the line pressure P _L also rises, and along with this, the load thrust control oil pressure P ₈ synergistically rises, and it is possible to speed up the shift to the low speed ratio.

The manual valve 110 works in conjunction with the shift lever to select P, R, N, D, L
It has a spool 111 which is operated at each position of, and this spool 111 allows two output ports from an input port 112.
The oil passages are selectively switched to 113 and 114. For example, in the D range, port 11 as shown
The line pressure is output from 3, and the port 114 is drained.
L range is the same as D range, and R range is port
113 is drained and the line pressure is output from the port 114. Further, in the P range, the land 111a closes the input port 112, and in the N range, the input port 112 and the output ports 113, 114 are cut off by the lands 111a, 111b, so that both output ports are drained.

The spool 121 of the second pressure regulating valve 120 is biased to the left by the spring 122, and the spool 121 has an input port 123 to which the line pressure is input from the first pressure regulating valve 100 and a drain port 124.
And are selectively opened and closed to output the hydraulic pressure P ₀ from the output port 125. The output hydraulic pressure P ₀ is fed back to the left end chamber 126 via the communication hole 121a provided inside the spool 121, and thereby the output hydraulic pressure P ₀ is adjusted to a constant pressure balanced only by the spring load. The output hydraulic pressure P ₀ is input to the three solenoid valves 301, 302, 303.

The direct connection control valve 130 is a valve for controlling the direct connection clutch 6, and has a pool 131 biased to the left by a spring 132. During direct connection drive (shown in the lower half of the drawing), the spring load and the signal hydraulic pressure P _{3 of the} direct connection control solenoid valve 303 guided to the left end chamber 133 face each other, and the spool 131 moves from the manual valve 110 to the line pressure. Input port 134
And drain port 135 selectively open and close, and output port 136
The hydraulic pressure P ₅ is directly connected to the clutch 6 and the clutch control valve described later
It is outputting to port 162 of 160. The output oil pressure P ₅ is transmitted through the communication hole 131 a provided inside the spool 131 to the spring 132.
Is fed back to the right end chamber 137 in which the signal oil pressure P ₃ is balanced with the sum of the output oil pressure P ₅ and the spring load. Since the line pressure P _L is input to the input port 134 from the output port 113 of the manual valve 110 only when the vehicle is moving forward (D, L), regardless of the signal hydraulic pressure P ₃ in other ranges, particularly in the R range. Output hydraulic pressure P ₅ becomes zero, and direct drive is not possible. When the belt is driven (shown in the upper half of the drawing), the output hydraulic pressure P ₄ of the clutch control valve 160 is input to the port 138, so the spool 131 is forcibly moved to the left end position, and the direct connection control solenoid valve 203. The hydraulic pressure P ₅ of the direct coupling clutch 6 is drained regardless of the operation of.

The direct connection control valve 130 is a solenoid valve for direct connection control 30 if the V-belt 15 is damaged and normal starting is impossible.
By controlling the duty of 3, the hydraulic pressure P ₅ to the direct coupling clutch 6 can be gently raised and an emergency start can be performed via the direct coupling drive path.

The three-way valve 140 is connected to the output port 113 of the manual valve 110 via the input port 134 of the direct connection control valve 130 to advance the line pressure during forward movement.
An input port 141 P _L is guided, an input port 142 where the line pressure during backward through the hydraulic switching valve 180 P _L is guided to be described later from the output port 114 of the manual valve 110, an input port 163 of the clutch control valve 160 Output port 143 to output hydraulic pressure to
The ball 144 reverses by the hydraulic pressure selectively input to the input ports 141 and 142, and selectively outputs the hydraulic pressure.

The forward / reverse switching piston 150 has a piston member 153 which is movable by hydraulic pressure acting on the left and right oil chambers 151, 152, and a fork shaft 154 for operating the forward / reverse switching dog clutch 23 is connected to the piston member 153. . The fork shaft 154 is constantly urged by the spring 155 to the forward position. When the line P _L is guided from the output port 113 of the manual valve 110 to the right chamber 152 through the input port 134 of the direct connection control valve 130 during forward movement, the state shown in the upper half of the drawing is reached, and the left chamber 151 of the manual valve 110 Output port
When the line pressure P _L is introduced from 114 through a hydraulic switching valve 180 described later at the time of retreat, the state shown in the lower half of the drawing is obtained. The right chamber 152 is connected to the outside oil through the one-way valve 156, and when the R range is switched to the P or N range, the operating time when the piston member 153 moves to the forward position by the spring 155. Has been shortened.

The clutch control valve 160 is a valve for controlling the starting clutch 20 and has a sleeve 161 having a total of six ports 162 to 167. Inside the sleeve 161, a spool 168 is freely arranged. The spool 168 is biased by the first spring 169 from the left side, and is biased by the second spring 170 and the third spring 171 from the left and right, and the sum of the loads of the second and third springs 170, 171 is equal to that of the first spring 169. Greater than load. The hydraulic pressure P ₅ is input to the port 162 from the direct connection control valve 130, the forward hydraulic pressure or the reverse hydraulic pressure is input to the input port 163 from the three-way valve 140, and the port 164 is the starting clutch 20 and the signal hydraulic port 138 of the direct connection control valve 130. Output hydraulic pressure P ₄ to
Port 165 is a drain port. The port 166 is connected to an output port 206 of a coast down control valve 200, which will be described later, and a signal hydraulic pressure P ₂ is led to a port 167 at the right end from a load thrust control solenoid valve 302. An orifice 35 and a one-way valve 36 are connected in parallel between the solenoid valve 302 and the port 167 to reduce the pulsation of the signal oil pressure P ₂ by the orifice 3 and to reduce the signal oil pressure P when the solenoid valve 302 is turned off. _2-way valve 36
Since it drains through, it is possible to eliminate a defective disconnection of the starting clutch 20. The output hydraulic pressure P ₄ is fed back to the chamber 172 accommodating the first spring 169 through a communication hole 168a formed inside the spool 168. Also spool 168
At the right end of the piston 174 with the fourth spring 173 in between.
Is slidably arranged, and the signal oil pressure P ₂ input to the right end port 167 pushes the spool 168 to the left via the piston 174 and the fourth spring 173, and the pulsation of the signal oil pressure P ₂ is caused by a kind of accumulator mechanism. Absorbing. The fourth
The chamber accommodating the sprig 173 communicates with the drain port 165 through the communication hole 168b of the spool 168.

The back of the first spring 169 is supported by a plug 175 slidably fitted to the left end of the sleeve 161, and the back of the plug 175 is supported by a creep adjusting bolt 176. Then, by rotating the bolt 176, the load of the first spring 169 is increased or decreased, and the creep torque of the starting clutch 20 can be finely adjusted.

The upper half of FIG. 2 of the clutch control valve 160 shows the start of the belt drive, and the diameter control solenoid valve 303 is turned off when the belt is driven.
(P ₃ = 0), the load thrust control solenoid valve 302 is operating (P ₂ =
Since it is 0 to ³ kg / cm ² ), the starting clutch 20 has a signal hydraulic pressure P ₂ of the load thrust control solenoid valve 302 and the second and third springs 1
The sum of the loads of 70,171, the load of the first spring 169 and the chamber 172
The sum of the output hydraulic pressure P ₄ fed back to is opposed, and the output hydraulic pressure P ₄ is regulated. Since the output hydraulic pressure P ₄ is input to the port 138 of the direct coupling control valve 130 in addition to the starting clutch 20,
The direct connection control valve 130 moves to the left end position as shown in the upper half of FIG. 2, and the direct connection clutch 6 is always disengaged.

When idling in the travel range (D, L, R), the signal oil pressure P ₂ input to the right end port 167 becomes 0, but the oil pressure P according to the load difference between the first spring 169 and the second and third springs 170, 171. _{Since 4} is output to the starting clutch 20, the starting clutch 20 generates a constant creep torque. Especially the second
Since the spring 170 is made of a shape memory alloy and the spring load is constant during warm-up, the spring load is set to be lower than during warm-up during cold conditions, so the output of the clutch control valve 160 The hydraulic pressure P _{4 is} also adjusted low. Therefore, an increase in creep torque due to an increase in friction coefficient or oil viscosity during cold is corrected by reducing the transmission capacity of the starting clutch 20 and controlled to a substantially constant creep torque regardless of cold or warm-up. it can.

Further, the lower half of the clutch control valve 160 in FIG. 2 shows a direct connection drive, the load thrust control solenoid valve 302 is OFF (P ₂ = 0), and the direct connection control solenoid valve 303 is ON (P ₃ = 3 kg / cm 2). ² ) because you are
The diameter control valve 130 operates as shown in the lower half of FIG. 2, the output hydraulic pressure P ₅ of the direct connection control valve 130 is input to the signal hydraulic port 162, and the spool 168 is forcibly moved to the right end position. Therefore, the hydraulic pressure P ₅ is output from the direct coupling control valve 130 to the direct coupling clutch 6, the direct coupling clutch 6 is engaged, and the output hydraulic pressure P ₄ of the clutch control valve 160 is drained, so that the starting clutch 20 is disengaged. In this way, the direct connection control valve 130 and the clutch control valve 160 change the output oil pressures P ₄ and P ₅ to the ports 138 and
By guiding to 162, hydraulic pressure can be supplied to only one of the starting clutch 20 and the direct coupling clutch 6, and so-called double clutch can be prevented.

The hydraulic pressure switching valve 180 is a valve for switching between the backward oil passage and the lubricating oil passage, and has a spool 182 biased to the left by a spring 181. The hydraulic pressure switching valve 180 has six ports 183 to 188, and the signal hydraulic pressure P ₁ of the solenoid valve 301 for gear ratio control is input to the left end port 183. Until the signal oil pressure P _{1 of the} gear ratio control solenoid valve 301 exceeds a certain value (for example, 1.3 kg / cm ² ), the spool 182 is at the left end position as shown in the upper half of the figure. That is, the signal hydraulic pressure P ₁ of the solenoid valve for proportional shift control 301 is lower than the spring load in the vicinity of the low speed ratio. Hydraulic pressure is supplied to the three-way valve 140 and the forward / reverse switching piston 150. At the same time, since the ports 186 and 187 are in communication, the lubricating oil output from the first pressure regulating valve 100 is directly guided to the starting clutch 20 via the ports 186 and 187 and the oil cooler 37. That is, in the low speed ratio position, the starting clutch 20 is often in the anti-clutch state as when starting, so a large amount of lubricating oil is introduced to the starting clutch 20 to prevent heat generation.

When the signal hydraulic pressure P _{1 of} the gear ratio controlling solenoid valve 301 exceeds a certain value (for example, 1.3 kg / cm ² ), the spool 182 moves to the right end position as shown in the lower half of the figure. That is, when the gear ratio is changed to the high speed ratio, the signal oil pressure P of the gear ratio control solenoid valve 301 is changed.
_{Since 1} becomes larger than the spring load, the spool 182 moves to the right end position and closes the port 184 so that the port 185 communicates with the drain port 188. Therefore, even if the change lever is mistakenly switched from the D range to the R range during high-speed traveling, the reverse hydraulic system is completely shut off as described above, and the forward / reverse switching piston 150 does not change to the reverse position.
At the same time, the port 186 is also closed, so that the lubricating oil is guided to the starting clutch 20 via the orifice 38. That is, since the starting clutch 20 is engaged at the high speed ratio position, the necessity of lubrication is low, and the lubricating oil is gradually supplied through the orifice 38. Note that the direct coupling clutch 6 does not exist in the half-clutch state except when an emergency starts, so that the lubricating oil is constantly supplied to the orifice 39.
Supply through.

The gear ratio control valve 190 is an oil chamber 1 for gear ratio control of the drive pulley 12.
The valve 6 controls the hydraulic pressure, and has a spool 192 biased to the left by a spring 191. In the left end chamber 193, the signal hydraulic pressure P ₁ of the solenoid valve 301 for gear ratio control is the hydraulic switching valve 1
Input via port 183 of 80, signal hydraulic P ₁
Is less than a fixed value (for example, 1.0 kg / cm ² ), the spring load is larger than the signal hydraulic pressure P ₁ , so the spool 192 is at the left end position as shown in the upper half of the figure. Therefore, the input port 194 to which the line pressure P _L is input is closed, and the output port 195 communicates with the drain port 196, so that the gear ratio control hydraulic pressure P ₆ is drained. When the signal hydraulic pressure P _{1 of the} gear ratio control solenoid valve 301 exceeds a certain value, the spool 192 overcomes the spring load and moves to the right as shown in the lower half of the figure, and the spool 192 moves to the input port 194 and the drain port. 196 and 196 are maintained near the position where they are selectively opened and closed. And
Since the output hydraulic pressure P ₆ is fed back to the signal hydraulic pressure port 197, the sum of the output hydraulic pressure P ₆ and the spring load is adjusted so that the signal hydraulic pressure P ₁ is balanced.

The coast down control valve 200 is a valve for forcibly shifting to a low speed ratio during coast down (when the throttle is fully closed while running to perform rapid deceleration), and is biased to the right by a spring 201. It has a spool 202. The signal hydraulic pressure P ₁ of the solenoid valve 301 for gear ratio control is input to the signal hydraulic pressure port 203 via the port 183 of the hydraulic pressure switching valve 180, and the signal hydraulic pressure P ₁ is a constant value (for example, 1.0 kg / cm ² ) In the following cases, as shown in the lower half of the figure, the spool 202
It is maintained near the position where the input port 204 to which _L is input and the drain port 205 are selectively opened and closed. Then, the output hydraulic pressure P ₇ output from the output port 206 is fed back to the right end port 207 via the communication hole 202a of the spool 202, so that the sum of the output hydraulic pressure P ₇ and the signal hydraulic pressure P ₁ and the spring load are The pressure is adjusted to balance. When the signal oil pressure P ₁ exceeds a certain value, the spool 202 overcomes the spring load and moves to the left end position as shown in the upper half of the figure, the input port 204 is closed, and the output port 207 communicates with the drain port 205. Therefore, the output oil pressure P ₇ becomes zero.

The output hydraulic pressure P ₇ is input to the port 166 of the clutch control valve 160 and the port 219 of the load thrust control valve 210 described later. Particularly, the output hydraulic pressure P ₇ input to the clutch control 160 causes the spool 168 to move to the right. The clutch hydraulic pressure of the starting clutch 20 is lowered by a constant value (for example, 0.5 kg / cm ² ) in order to accelerate the coast down.

The load thrust control valve 210 is a valve for controlling the hydraulic pressure P ₈ of the load thrust control oil chamber 17 of the driven pulley 14, and has a main spool 211 and a sub spool 212 as shown in FIG. 2A. A first spring 213 is interposed between the spools 211 and 212. The load thrust control valve 210 outputs an oil pressure P ₈ to the input port 214 to which the line pressure P _L is constantly input, the load thrust control oil chamber 17 and the back pressure chamber 109 of the first pressure regulating valve 100.
And a drain port 216 are formed, and the output hydraulic pressure P ₈ is fed back to the right end chamber 217 via the communication hole 211a of the main spool 211. Therefore, the output hydraulic pressure P ₈ is adjusted so as to balance with the rightward load acting on the spool 211 from the sub spool 212 via the first spring 213. The sub spool 212 has a second spring having a smaller spring force than the first spring 213.
The spring 218 is urged to the right, and the output hydraulic pressure P ₇ of the coast down control valve 200 is input to the port 219 accommodating the spring 218, and the port 219 is further accommodated.
The signal hydraulic pressure P ₂ is input from the load thrust control solenoid valve 302 to the port 220 adjacent to the port 220. Therefore, the main spool 211
The rightward load acting on is determined only by the load of the first spring 213 when the sum of the load of the second spring 218 and the signal hydraulic pressures P ₂ and P ₇ is smaller than the load of the first spring 213. When the sum of the load of the second spring 218 and the signal hydraulic pressures P ₂ and P ₇ is larger than the load of the first spring 213, the sum of the load of the second spring 218 and the signal hydraulic pressures P ₂ and P ₇ is calculated. Determined (shown in the bottom half of the drawing). That is, the hydraulic pressure P ₈ of the load thrust control oil chamber 17 is equal to the hydraulic pressure P _{4 of the} starting clutch 20.
It goes up together with it, and further goes up when the course is down to speed up shifting to the low speed ratio.

FIG. 3 shows the relationship between the signal oil pressure P ₂ input to the load thrust control valve 210 and the output oil pressure P _8, and the load thrust control solenoid valve 302 that generates the signal oil pressure P ₂ is the load thrust control valve 210 only. Not only as a solenoid valve for controlling the clutch control valve 160, but the clutch torque is in a proportional relationship with the engine generated torque, and the clutch torque and the signal oil pressure P ₂ of the solenoid valve 302 are also in a proportional relationship. The horizontal axis in the figure may be replaced with the engine generated torque. Further, since the output hydraulic pressure (load hydraulic pressure) P _{8 on the} vertical axis of FIG. 3 is proportional to the load thrust, the vertical axis may be replaced by the load thrust.

The spring loads F ₁ and F ₂ of the first and second springs 213 and 218 of the load thrust control valve 210, the pressure receiving area A ₁ of the right end portion 211b of the main spool 211,
The pressure receiving areas A ₃ and A ₄ of the lands 212a and 212b of the sub spool 212 are
It is determined to satisfy the following four conditions.

First, if the minimum load hydraulic pressure at which the V-belt does not slip during engine braking is 0.6 kg / cm ² , the following equation holds.

0.6 × A ₁ = F ₁ (1) Secondly, the signal hydraulic pressure P ₂ when the sub spool 212 starts to operate is
When 0.43kg / cm ² , the following formula is established.

0.43 × time _{(A 3 -A 4) + F} 2 = F 1 ...... (2) oil pressure P ₇ at third coastdown is 2.9 kg / cm ^2,
When trying to raise the load hydraulic pressure P ₈ by 27 kg / cm ² , the following equation holds.

2.9 × A ₄ = 2.7 × A ₁ (3) Fourthly, when the load thrust force valve 210 reaches the position shown in the lower half of Fig. 2, the following equation holds.

P ₈ × A ₁ = P ₂ × (A ₃ −A ₄ ) + F ₂ (4) To satisfy the above four conditions, the main spool 211,
Sub spool 212, first spring 213, second spring 21
8 is required, and the minimum load hydraulic pressure, the slope of the proportional straight line, and the base point X can be set to arbitrary values by these configurations.
For example, if the second spring 218 is not provided, even if the minimum load hydraulic pressure and the slope of the proportional straight line are made to coincide, the base point X shifts to X ′ as shown by the broken line in FIG. 3, and the load thrust tends to become insufficient, On the other hand, even if the minimum load hydraulic pressure and the base point X are made to coincide with each other, the slope of the proportional straight line becomes large as shown by the alternate long and short dash line in FIG. Therefore, in order to completely match the characteristics shown by the solid line in FIG. 3, in addition to the main spool 211, the sub spool 212, and the first spring 213,
The second spring 218 is essential.

The line pressure P _L is directly input to the input port 214 of the load thrust control valve 210 without passing through the manual valve 110.
The reason is that the hydraulic pressure OFF next to both pulleys 12, 14 when placed in the N range during high-speed traveling, by the V-belt 15 is so loosened by centrifugal force, to always enter the line pressure P _L, the pulley 12 This is to apply the minimum pressure to the V and belt 14 so that the V belt 15 does not sag.

FIG. 4 is a block diagram of the controller 300. 310 is a sensor that detects the engine speed N _in (the rotation speed of the input shaft 4), 311 is a sensor that detects the vehicle speed V (the rotation speed of the output shaft 32), 312 Is a sensor that detects the rotation speed N _out of the driven shaft 13 (the input rotation speed of the starting clutch 20 or the rotation speed of the driven pulley 14), and 313 is the P, R, N, D, and L shift positions of the manual valve 110. 314 is a sensor for detecting the throttle opening. The signals from the sensors 310 to 313 are input to the input interface 315, and the signal from the sensor 314 is the A / D converter 3
Converted to digital signal at 16. 317 is a central processing unit (CPU), 318 is a read only memory (ROM) in which programs and data for controlling the solenoid valves 301, 302, 303 are stored, and 319 is a signal or parameter temporarily sent from each sensor. Random access memory to store (RA
M), 320 are output interfaces, and these CPU31
7, ROM 318, RAM 319, output interface 320, input interface 315 and A / D converter 316 are interconnected by a bus 321. The output of the output interface 320 is output as a control signal to each solenoid valve 301, 302, 303 via the output driver 322, and the solenoid valves 301 to 303 control the output hydraulic pressures P ₁ , P ₃ , P ₅ by these control signals to control the shift control. , Start control and direct connection control are being executed.

Of the solenoid valves 301 to 303, the load thrust control solenoid valve 302,
The direct coupling control solenoid valve 303 is a normally closed type, whereas the gear ratio control solenoid 301 is a normally open type. Therefore, the solenoid valve 30
2, 303 drains when the electric signal is OFF and generates the maximum signal hydraulic pressure when ON, whereas the solenoid valve 301 generates the maximum signal hydraulic pressure when OFF and is drained when ON. In this way, the gear ratio control solenoid valve 301 is configured to be a normally open type.
Even if the solenoid valve 301 should fail, the signal oil pressure P ₁ is generated, the gear ratio control valve 190 is activated to increase the gear ratio control oil pressure P ₆ , and the gear is constantly shifted to the high speed ratio side, which is accompanied by a sudden downshift. This is to avoid shock.

In the above embodiment, the auxiliary spool 2 of the load thrust control valve 210 is
Although the hydraulic pressure P ₇ was introduced from the coast down control valve 200 to 12 and the load hydraulic pressure P ₈ was increased during the coast down,
This configuration is not mandatory. Therefore, the equation (3) is not necessary in the present invention.

Further, in the above embodiment, the load hydraulic pressure and the starting clutch hydraulic pressure are controlled by the single solenoid valve 302, but it goes without saying that they may be controlled by separate solenoid valves.

Further, the V-belt type continuously variable transmission of the present invention is not limited to the one in which the gear ratio control oil chamber is provided in the drive side pulley and the load thrust control oil chamber is provided in the driven side pulley. Good.

〔The invention's effect〕

As is clear from the above description, according to the present invention, the load thrust control valve is provided with the main spool, the sub spool, the first spring and the second spring.
Determined by the spring load of the spring, the slope of the proportional straight line between the load hydraulic pressure and the engine generated torque can be determined by the pressure receiving area of the main and auxiliary spools, and the load hydraulic pressure starts to rise in proportion to the engine generated torque from the minimum load hydraulic pressure. Is the second
It can be determined by the spring load of the spring.

Therefore, the load thrust characteristic can be freely set to a desired value, and the optimum load thrust characteristic corresponding to the change in the engine generated torque can be obtained.

[Brief description of drawings]

FIG. 1 is a schematic configuration diagram of an example of a V-belt type continuously variable transmission according to the present invention, FIG. 2 is a circuit diagram of a hydraulic control device, FIG. 2A is an enlarged view of a load thrust control valve, and FIG. 3 is a load. Fig. 4 is a characteristic diagram of hydraulic pressure and signal hydraulic pressure, and Fig. 4 is a block diagram of the controller. 1 ... Engine, 5 ... Input shaft, 6 ... Direct connection clutch,
10 ... continuously variable transmission, 12 ... drive side pulley, 14 ... driven side pulley, 16 ... gear ratio control oil chamber, 17 ... load thrust control oil chamber, 20 ... start clutch, 32 ... … Output shaft, 160…
... Clutch control valve, 210 ... Load thrust control valve (control valve), 211 ... Main spool, 212 ... Sub spool, 213 ...
… First spring, 214 …… input port, 215 …… output port, 216 …… drain port, 217 …… right end chamber (output hydraulic pressure return port), 218 …… second spring, 220 …… signal hydraulic pressure input port, 302 …… Solenoid valve for controlling load thrust (solenoid valve).

Claims (1)

[Claims] 1. A gear thrust control oil chamber for controlling a load thrust required for torque transmission is provided on one of a drive side pulley and a driven side pulley, and a gear ratio control for controlling a gear ratio on the other side. A V-belt type continuously variable transmission provided with an oil chamber includes a control valve for controlling the hydraulic pressure of the load thrust control oil chamber and a solenoid valve for generating a signal hydraulic pressure for controlling the damage control valve. The control valve includes an input port to which an operating pressure is input from a hydraulic pressure source, a drain port, and an output port that is formed between the two ports and outputs hydraulic pressure to the load thrust control oil chamber.
A main spool that selectively communicates the input port and the drain port with the output port, and an output hydraulic pressure feedback port to which the output hydraulic pressure is fed back to urge the main spool in the direction in which the output port and the drain port communicate with each other, From the sub spool arranged on the drain port side of the main spool, the first spring interposed between both spools and biasing the spools in opposite directions, and the first spring biasing the sub spool to the main spool side. A V-belt type continuously variable transmission characterized in that a second spring having a small spring force and a signal hydraulic pressure input port through which the signal hydraulic pressure of the solenoid valve is guided so as to urge the auxiliary spool toward the main spool side are provided. Machine hydraulic control device.