JPH07117143B2 - Hydraulic control device for vehicle belt type continuously variable transmission - Google Patents

Hydraulic control device for vehicle belt type continuously variable transmission

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Publication number
JPH07117143B2
JPH07117143B2 JP29128890A JP29128890A JPH07117143B2 JP H07117143 B2 JPH07117143 B2 JP H07117143B2 JP 29128890 A JP29128890 A JP 29128890A JP 29128890 A JP29128890 A JP 29128890A JP H07117143 B2 JPH07117143 B2 JP H07117143B2
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JP
Japan
Prior art keywords
pressure
valve
line
gear ratio
hydraulic pressure
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Expired - Fee Related
Application number
JP29128890A
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Japanese (ja)
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JPH04191555A (en
Inventor
信幸 加藤
勇仁 服部
孝士 林
克己 河野
Original Assignee
トヨタ自動車株式会社
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Priority to JP2-23579 priority Critical
Priority to JP2357990 priority
Priority to JP13448990 priority
Priority to JP2-134489 priority
Priority to JP2-241764 priority
Priority to JP24176490 priority
Application filed by トヨタ自動車株式会社 filed Critical トヨタ自動車株式会社
Priority claimed from EP91300676A external-priority patent/EP0440422B1/en
Publication of JPH04191555A publication Critical patent/JPH04191555A/en
Publication of JPH07117143B2 publication Critical patent/JPH07117143B2/en
Anticipated expiration legal-status Critical
Expired - Fee Related legal-status Critical Current

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Description

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a hydraulic control device for a belt type continuously variable transmission for a vehicle.

2. Description of the Related Art A pair of variable pulleys respectively provided on a primary-side rotary shaft and a secondary-side rotary shaft, a transmission belt wound around the pair of variable pulleys to transmit power, and effective diameters of the pair of variable pulleys. There is known a belt type continuously variable transmission for a vehicle that includes a pair of hydraulic actuators that change the. As a hydraulic control device provided in such a belt type continuously variable transmission, there is, for example, one disclosed in Japanese Patent Laid-Open No. 64-49755. In this hydraulic control device, the tension control pressure (line hydraulic pressure) for controlling the tension of the transmission belt is used.
Is changed with the gear ratio, and the amount of change is adjusted by a pressure regulating valve having a bending characteristic that changes at a midway point based on the required output pressure and the gear ratio pressure. According to this, there is a feature that the tension control pressure generated by the pressure regulating valve is relatively approximate to the target change tendency of the optimum control pressure with a relatively inexpensive and highly reliable configuration.

Problems to be Solved by the Invention By the way, since the optimum control pressure for making the tension of the transmission belt a necessary and sufficient value is a curve, the polygonal tension control pressure obtained by the pressure regulating valve does not fall below the optimum control pressure. In such a state in which the tension control pressure is unnecessarily increased with respect to the optimum control pressure in the state in which the linear control is set as described above, there is a disadvantage that the power loss of the oil pump driven by the engine cannot be sufficiently reduced. .

The present invention has been made in the background of the above circumstances,
It is an object of the present invention to provide a hydraulic control device capable of approximating the tension control pressure related to the tension of the transmission belt to the optimum control pressure as much as possible.

Means for Solving the Problems The gist of the present invention for achieving such an object is to provide a pair of variable pulleys respectively provided on a primary-side rotating shaft and a secondary-side rotating shaft, and a pair of the variable pulleys. A hydraulic control device for a belt-type continuously variable transmission for a vehicle, comprising: a transmission belt wound between; and a pair of hydraulic actuators that change the effective diameters of a pair of variable pulleys,
(A) A valve element that operates based on a gear ratio that represents a gear ratio of the belt type continuously variable transmission, a required output pressure that represents a required output of the vehicle, and a signal pressure is provided. When the signal pressure is generated, the optimum control pressure of the tension of the transmission belt is generated, and when the signal pressure is not generated, the tension control pressure larger than the optimum control pressure is increased as the gear ratio increases. And a pressure regulating valve that is generated so as to increase as the required output of the vehicle increases,
(B) a signal pressure generating means for generating the signal pressure according to an electric signal; (c) a gear ratio determining means for determining a gear ratio of the belt type continuously variable transmission; and (d) a required output value of the vehicle. From the required output value sensor to be detected and (e) the relationship stored in advance, the signal pressure generating means is controlled so that the optimum control pressure is generated from the pressure regulating valve based on the actual gear ratio and the required output value. And an electronic control unit that operates.

According to this configuration, the electronic control unit causes the electronic control unit to generate the optimum control pressure from the pressure regulating valve based on the relationship stored in advance so as to generate the optimum control pressure based on the actual gear ratio and the required output value. The pressure generating means is controlled. Therefore, there is no portion where the tension control pressure is unnecessarily increased, the power loss is greatly improved, and the belt durability is also improved.

Embodiment Hereinafter, one embodiment of the present invention will be described in detail with reference to the drawings.

In FIG. 2, the power of the engine 10 includes a fluid coupling 12 with a lockup clutch, a belt type continuously variable transmission (hereinafter referred to as CVT).
14, a forward / reverse switching device 16, an intermediate gear device 18, and a differential gear device 20, and a drive wheel 24 connected to a drive shaft 22.
Is being transmitted to.

The fluid coupling 12 includes a pump impeller 28 connected to the crankshaft 26 of the engine 10, a turbine impeller 32 fixed to the input shaft 30 of the CVT 14 and rotated by oil from the pump impeller 28, and a damper 34. A lock-up clutch 36 fixed to the input shaft 30 via an engagement side oil chamber 33 connected to an engagement side oil passage 322 described below and a release side oil chamber 35 connected to a release side oil passage 324 described below. It has and. The fluid coupling 12 is constantly filled with hydraulic oil. For example, when the vehicle speed, the engine rotation speed, or the rotation speed of the turbine impeller 32 exceeds a predetermined value, the hydraulic oil is supplied to the engagement side oil chamber 33 and released. When the hydraulic oil flows out from the side oil chamber 35, the lockup clutch 36 is engaged and the crankshaft 26 and the input shaft 30 are directly connected. On the other hand, when the vehicle speed or the like becomes a predetermined value or less, hydraulic oil is supplied to the disengagement side oil chamber 35 and hydraulic oil flows out from the engagement side oil chamber 33, so that the lockup clutch 36 is released. .

The CVT 14 includes variable pulleys 40 and 42 of the same diameter provided on the input shaft 30 and the output shaft 38 thereof, respectively, and a transmission belt 44 wound around the variable pulleys 40 and 42. Variable pulleys 40 and 42 are for input shaft 30 and output shaft
Fixed rotating bodies 46 and 48 fixed to 38, respectively, and movable rotating bodies 50 and 52 provided to the input shaft 30 and the output shaft 38 so as to be movable in the axial direction and incapable of relative rotation about the axis.
The movable rotary bodies 50 and 52 are moved by the primary side hydraulic cylinder 54 and the secondary side hydraulic cylinder 56 that function as hydraulic actuators, so that the V groove width, that is, the hanging diameter (effective diameter) of the transmission belt 44 is changed. Thus, the gear ratio γ of the CVT 14 (= rotational speed N in of the input shaft 30 / rotational speed N out of the output shaft 38) is changed.
Since the variable pulleys 40 and 42 have the same diameter, the hydraulic cylinders 54 and 56 have the same pressure receiving area. Normal,
The pressure of the driven side of the hydraulic cylinders 54 and 56 is related to the tension of the transmission belt 44.

The forward / reverse switching device 16 is a well-known double pinion type planetary gear mechanism, and includes a pair of planetary gears 62 and 64 that are rotatably supported by a carrier 60 fixed to an output shaft 58 of the planetary gear mechanism 62 and 64 and mesh with each other. The planetary gear 6 fixed to the input shaft (output shaft of the CVT 14) 38 of the progression switching device 16 and on the inner peripheral side
2, a sun gear 66 that meshes with 2, a ring gear 68 that meshes with the planetary gear 64 on the outer peripheral side, a reverse brake 70 for stopping the rotation of the ring gear 68, the carrier 60 and the input shaft 38 of the forward / reverse switching device 16 are connected. And a forward clutch 72 for driving. The reverse brake 70 and the forward clutch 72 are frictional engagement devices of the type that are hydraulically actuated, and when they are not engaged together, the forward / reverse switching device 16
Is made neutral and power transmission is cut off. However, when the forward clutch 72 is engaged, the output shaft 38 of the CVT 14
And the output shaft 58 of the forward / reverse switching device 16 are directly connected to each other to transmit power in the forward direction of the vehicle. Further, when the reverse brake 70 is engaged, the output shaft 38 of the CVT 14 and the forward / reverse switching device 16
Since the rotation direction is reversed between the output shaft 58 and the output shaft 58, the power in the backward direction of the vehicle is transmitted.

FIG. 1 shows a hydraulic control circuit for controlling the vehicle power transmission device shown in FIG. The oil pump 74 serves as a hydraulic source of the hydraulic control circuit, and
By being integrally connected with the twelve pump impellers 28, the crankshaft 26 is constantly driven to rotate. The oil pump 74 sucks the hydraulic oil that has flowed back into an oil tank (not shown) through the strainer 76, and also sucks the hydraulic oil that is returned through the suction oil passage 78 to suck the first hydraulic oil.
Pump to line oil passage 80. In the present embodiment, the hydraulic oil in the first line oil passage 80 is leaked to the intake oil passage 78 and the lockup clutch pressure oil passage 92 by the overflow (relief) type first pressure regulating valve 100, so that the first line The first line hydraulic pressure Pl 1 in the oil passage 80 is adjusted. Further, the second line oil pressure Pl 2 in the second line oil passage 82 is adjusted by reducing the first line oil pressure Pl 1 by the second pressure adjusting valve 102 of the pressure reducing valve type.

First, the configuration of the second pressure regulating valve 102 will be described. As shown in FIG. 3, the second pressure regulating valve 102 includes a spool valve element 110, a spring seat 112, a return spring 114, and a plunger 11 that open and close between the first line oil passage 80 and the second line oil passage 82.
Equipped with 6. Further, at the shaft end of the spool valve element 110, the first land 118, the second land 120, the third land 120
Lands 122 are sequentially formed. Second land 120 and third
A chamber 126 into which the second line hydraulic pressure Pl 2 is introduced as a feedback pressure through the throttle 124 is provided between the land 122 and the spool valve element 110 and is biased in the valve closing direction by the second line hydraulic pressure Pl 2. It has become so. Further, on the end surface side of the first land 118 of the spool valve element 110, there is provided a chamber 130 into which a gear ratio specific pressure P r, which will be described later, is introduced via a throttle 128,
The spool valve element 110 is biased in the valve closing direction by the gear ratio specific pressure P r . In the second pressure regulating valve 102, the biasing force of the return spring 114 in the valve opening direction is applied to the spool valve element 110 via the spring seat 112. Further, a throttle pressure P th, which will be described later, is set on the end face side of the plunger 116.
A chamber 132 for actuating the valve is provided, and the spool valve element 110 is biased in the valve opening direction by the throttle pressure P th . Therefore, the pressure receiving area of the first land 118 is A 1 , the cross-sectional area of the second land 120 is A 2 ,
If the cross-sectional area of the land 122 is A 3 , the pressure receiving area of the plunger 116 is A 4 , and the urging force of the return spring 114 is W,
The spool valve element 110 is balanced at a position where the following expression (1) is established. That is, the spool valve element 110 is moved according to the equation (1), so that the port
A state in which the hydraulic oil in the first line oil passage 80, which is guided to 134a, is made to flow into the second line oil passage 82 via the port 134b, and the operation in the second line oil passage 82, which is guided to the port 134b. The state in which the oil is made to flow to the drain port 134c communicating with the drain is repeated, and the second line hydraulic pressure Pl 2 is generated. Since the second line oil passage 82 is a relatively closed system, the second pressure regulating valve 102 reduces the first line oil pressure Pl 1 which is a relatively high oil pressure as described above, and thus the second line oil pressure 82. Pl 2 is generated as shown in FIG.

Pl 2 = (A 4 · P th + W−A 1 · P r ) / (A 3 −A 2 ) ... (1) Between the first land 118 and the second land 120 of the spool valve 110. In the second line oil pressure reduction control valve 380, which will be described later.
A chamber 136 into which the signal pressure P soL4 is introduced is provided, and when the spool valve element 110 is biased in the valve closing direction by the signal pressure P soL4 , the second line hydraulic pressure Pl is set according to the magnitude thereof.
2 is corrected. The second line hydraulic pressure characteristic in this case will be described in detail later.

The first pressure regulating valve 100, as shown in FIG.
40, spring seat 142, return spring 144, first plunger 146, and second of the first plunger 146
A second plunger 148 having the same diameter as the land 155 is provided. The spool valve element 140 is connected to the port 1 that communicates with the first line oil passage 80.
A chamber 153 for opening and closing between the drain port 150a and the drain port 150b or 150c is provided on the end face of the first land 152 thereof for allowing the first line hydraulic pressure Pl 1 as a feedback pressure to act through the throttle 151. Therefore, the spool valve element 140 is biased in the valve opening direction by the first line hydraulic pressure Pl 1 . A chamber 156 for guiding the throttle pressure P th is provided between the first land 154 and the second land 155 of the first plunger 146 provided coaxially with the spool valve element 140, and the second land is also provided. Between the 155 and the second plunger 148, the oil pressure P in in the primary side hydraulic cylinder 54 is branched to the oil passage 305.
A chamber 157 for guiding the second line hydraulic pressure Pl 2 is provided on the end surface of the second plunger 148. Return spring 1
Since the urging force of 44 is applied to the spool valve element 140 in the valve closing direction via the spring seat 142, the pressure receiving area of the first land 152 of the spool valve element 140 is A 5 , and the first plunger is
Assuming that the cross-sectional area of the first land 154 of 146 is A 6 , the cross-sectional area of the second land 155 and the second plunger 148 is A 7 , and the urging force of the return spring 144 is W, the spool valve element 140 has the following formula (2). Is balanced at a position where is satisfied, and the first line hydraulic pressure Pl 1 is regulated.

Pl 1 = [(P in orPl 2 ) ・ A 7 + P th (A 6 −A 7 ) + W] / A 5
(2) In the first pressure regulating valve 100, the primary side hydraulic cylinder 54
When the internal oil pressure P in is higher than the second line oil pressure Pl 2 (Pl 2 = internal oil pressure P out of the secondary side hydraulic cylinder 56 in the steady state),
Although the first plunger 146 and the second plunger 148 are separated from each other, the thrust due to the oil pressure P in in the primary hydraulic cylinder 54 acts in the valve closing direction of the spool valve element 140, but the hydraulic pressure P in the primary hydraulic cylinder 54 is increased. When in is lower than the second line hydraulic pressure Pl 2 , the first plunger 146 and the second plunger 148 contact each other, so that the second line hydraulic pressure Pl 2 acting on the end surface of the second plunger 148 causes Thrust is spool valve 140
Acts in the valve closing direction of. That is, the primary side hydraulic cylinder
The second plunger 148, which receives the internal oil pressure P in and the second line oil pressure Pl 2 , applies the acting force based on the higher one of these oil pressures in the valve closing direction of the spool valve 140. In addition, between the first land 152 and the second land 159 of the spool valve element 140, a first line hydraulic pressure reduction control valve 4 described later is provided.
A chamber 160 to which the second line hydraulic pressure Pl 2 is supplied from 40 via an oil passage 161 is provided. The second line oil pressure Pl 2 acting in the chamber 160 acts in the direction of decreasing the first line oil pressure Pl 1 , and the first line oil pressure Pl 2 is in the N (neutral) and P (parking) ranges. When the reduction control valve 440 is operated and the second line hydraulic pressure Pl 2 is supplied to the chamber 160, the first line hydraulic pressure Pl 1 is reduced. First in this case
The line hydraulic pressure characteristic will be described in detail later.

Returning to FIG. 1, the throttle pressure P th represents the actual throttle valve opening θ th in the engine 10, and is generated by the throttle valve opening detection valve 180. Further, the gear ratio pressure P r represents the actual gear ratio of the CVT 14, and is generated by the gear ratio detection valve 182. That is, the throttle valve opening detection valve 180 includes a cam 184 that is rotated together with a throttle valve (not shown), and the cam 184.
Of the plunger 186 that is engaged with the cam surface of the cam 184 and is driven in the axial direction in relation to the rotation angle of the cam 184, and the thrust from the plunger 186 applied via the spring 188 and the first line hydraulic pressure Pl 1 . It is provided with a spool valve element 190 that reduces the first line hydraulic pressure Pl 1 by being positioned at a position where the thrust and the thrust are in equilibrium, and generates the throttle pressure P th corresponding to the actual throttle valve opening θ th . Figure 5 is shows the relationship between the throttle pressure P th and the throttle valve opening theta th, throttle pressure P th is the first pressure regulating valve 100 through the oil passage 84, second pressure regulating valve 102, and the third tone Each is supplied to the pressure valve 220.

Further, the gear ratio detection valve 182 is provided on the input side movable rotating body 50 of the CVT 14.
A detection rod 192 that is slidably contacted with and moved in the axial direction by a displacement equal to the displacement in the axial direction, a spring 194 that transmits an urging force corresponding to the position of the detection rod 192, and a spring 194 from this spring 194. A spool valve element 198 that changes the discharge flow rate to the drain by receiving the second line hydraulic pressure Pl 2 and being positioned at a position where the thrust forces of the two are balanced while receiving the urging force. Therefore, for example, the gear ratio γ becomes smaller and the fixed rotating body 4 on the input side of the CVT 14
When the movable rotating body 50 approaches 6 (reduces the V groove width),
The detection rod 192 is pushed in. Therefore, the flow rate of the working oil supplied from the second line oil passage 82 through the orifice 196 and discharged to the drain by the spool valve 198 is reduced, so that the working hydraulic pressure on the downstream side of the orifice 196 is increased. This operating oil pressure is the gear ratio pressure (signal) Pr
As shown in FIG. 6, the gear ratio γ is increased as the gear ratio γ decreases (changes to the speed increasing side). The gear ratio P r thus generated is supplied to the second pressure regulating valve 102 and the third pressure regulating valve 220 as oil pressure signals through the oil passage 86.

Here, the gear ratio detection valve 182 is provided with a second line oil pressure Pl 2 supplied from the second line oil passage 82 through the orifice 196.
Speed ratio pressure P r by changing the amount relief of the hydraulic fluid
Therefore, the gear ratio specific pressure P r is limited to be a value equal to or higher than the second line oil pressure Pl 2 while the second pressure regulating valve 102 operating according to the above equation (1) is used. The second line hydraulic pressure Pl 2 is decreased as P r is increased.
For this reason, when the gear ratio P r increases to a predetermined value and becomes equal to the second line oil pressure P l 2 , both thereafter become saturated and constant. FIG. 7 shows the second line hydraulic pressure Pl which is regulated in the second pressure regulating valve 102 in relation to the above-mentioned gear ratio specific pressure P r.
2 shows the basic output characteristics. This basic output characteristic is
It is obtained by the operation when the fourth signal pressure P SOL4 supplied to the chamber 136 of the second pressure regulating valve 102 is zero or constant, and is initially constant when the gear ratio γ increases from the minimum value. Although it is a value, it is a polygonal line characteristic that increases linearly after it matches the transmission specific pressure P r . FIG. 8 shows an ideal curve showing an optimum control pressure necessary and sufficient for controlling the tension of the transmission belt 44, and the basic output characteristic of the second pressure regulating valve 102 is approximated to the ideal curve. There is. It should be noted that, even though it is basically approximated as described above, as shown by the difference between the broken line and the solid line in FIG. 21 described later between the ideal curve and the polygonal basic output characteristic. There is a region where the second line hydraulic pressure Pl 2 is unnecessarily increased with respect to the ideal curve. In order to eliminate such a difference between the broken line and the solid line and match the second line hydraulic pressure Pl 2 with the ideal curve, the fourth line in the forward traveling state in which the lockup clutch 36 is engaged, as described later. The signal pressure P SOL4 is adjusted by being duty-driven by the electronic control unit 460.

The third pressure regulating valve 220 is for generating an optimum third line hydraulic pressure Pl 3 for operating the reverse brake 70 and the forward clutch 72 of the forward / reverse switching device 16. That is, the third pressure regulating valve 220 includes a spool valve element 222 that opens and closes between the first line oil passage 80 and the third line oil passage 88, a spring seat 224, a return spring 226, and a plunger 228. Between the first land 230 and the second land 232 of the spool valve element 222, there is provided a chamber 236 into which the third line hydraulic pressure Pl 3 is introduced as a feedback pressure through the throttle 234, and the spool valve element 222 is the third valve. It is designed to be urged in the valve closing direction by the line hydraulic pressure Pl 3 . Further, on the first land 230 side of the spool valve element 222, there is provided a chamber 240 into which the gear ratio specific pressure P r is guided via the throttle 238.
22 is biased in the valve closing direction by the gear ratio specific pressure P r . In the third pressure regulating valve 220, the valve opening direction urging force of the return spring 226 is applied to the spool valve element 222 via the spring seat 224. Also, plunger 2
A chamber 242 for exerting a throttle pressure P th is provided on the end face of 28, and the spool valve element 222 is provided with the throttle pressure P th.
The valve is biased in the valve opening direction by th . Further, between the first land 244 of the plunger 228 and the second land 246 having a smaller diameter than the first land 244, the third line hydraulic pressure P is set only during reverse travel.
A chamber 248 for guiding l 3 is provided. Therefore, the third line hydraulic pressure Pl 3 is adjusted to an optimum value based on the gear ratio specific pressure P r and the throttle pressure P th from the same formula as the formula (1). This optimum value is a value necessary and sufficient for surely transmitting torque without slippage in the forward clutch 72 or the reverse brake 70. Also, when moving backwards,
Since the third line hydraulic pressure Pl 3 is introduced into 248, the force for urging the spool valve element 222 in the valve opening direction is increased and the third line hydraulic pressure Pl 3 is increased. As a result, in the forward clutch 72 and the reverse brake 70, suitable torque capacities can be obtained during forward travel and during reverse travel, respectively.

The third line hydraulic pressure Pl 3 adjusted as described above is supplied to the forward clutch 72 or the reverse brake 70 by the manual valve 250. That is,
The manual valve 250 includes a spool valve element 254 that is moved in association with the operation of the shift lever 252 of the vehicle, and does not output the third line hydraulic pressure Pl 3 when the shift lever 252 is operated in the N range. But L (low),
When operated to the S (second) or D (drive) range, the third line hydraulic pressure Pl 3 is exclusively output from the output port 258 to the forward clutch 72 and the reverse inhibit valve 420.
In the state where the oil is drained from the reverse brake 70 at the same time as it is supplied to the chamber 432 of the vehicle and is operated to the R (reverse) range, the third line hydraulic pressure Pl 3 is output from the output port 256 to the third pressure regulating valve 220,
Lockup control valve 320, first line hydraulic pressure reduction control valve 440
Chamber 452, and reverse inhibit valve 420 port 42
Supply to 2a and its reverse inhibit valve 42
The oil is supplied from 0 to the reverse brake 70, and at the same time oil is drained from the forward clutch 72. When the P range is operated, both the forward clutch 72 and the reverse brake 70 are drained. The accumulators 342 and 340 are for gradually raising the hydraulic pressure to smoothly advance the frictional engagement, and are connected to the forward clutch 72 and the reverse brake 70, respectively. Further, the shift timing valve 210 adjusts the transient inflow flow rate by closing the throttle 212 according to the increase in the hydraulic pressure in the hydraulic cylinder of the forward clutch 72.

First line hydraulic pressure Pl 1 regulated by the first pressure regulating valve 100
And the second line hydraulic pressure P regulated by the second pressure regulating valve 102.
L 2 is supplied to one or the other of the primary side hydraulic cylinder 54 and the secondary side hydraulic cylinder 56 by the shift control valve device 260 in order to adjust the speed ratio γ of the CVT 14. The shift control valve device 260 includes a shift direction switching valve 262 and a flow rate control valve.
It is composed of 264. In addition, those shift direction switching valves 2
The fourth line hydraulic pressure Pl 4 for driving the 62 and the flow rate control valve 264 is generated by the fourth pressure regulating valve 170 based on the first line hydraulic pressure Pl 1 and guided by the fourth line oil passage 370. There is.

The shift direction switching valve 262 is a spool valve controlled by the first solenoid valve 266, and the flow control valve 264 is a spool valve controlled by the second solenoid valve 268. For example, when the first solenoid valve 266 is on and the second solenoid valve 268 is off, the working oil in the first line oil passage 80 is the shift direction switching valve 262, the flow control valve 264 and the two-way control valve 264. The hydraulic oil in the primary hydraulic cylinder 54 is discharged to the drain through the primary oil passage 300, the flow rate control valve 264, and the speed change direction switching valve 262 while being made to flow into the secondary hydraulic cylinder 56 through the secondary oil passage 302. Thus, the gear ratio γ of the CVT 14 is quickly changed in the deceleration direction. Conversely, the first solenoid valve 266 is off and the second solenoid valve 266 is off.
When the solenoid valve 268 is on, the first line oil passage 8
The hydraulic oil in 0 is made to flow into the primary hydraulic cylinder 54 through the speed change-over valve 262, the flow control valve 264 and the primary oil passage 300, while the hydraulic oil in the secondary hydraulic cylinder 56 is The oil is discharged to the second line oil passage 82 through the oil passage 302, the flow rate control valve 264, and the shift direction switching valve 262, and the CVT14
The gear ratio γ of is rapidly changed in the speed increasing direction.

FIG. 9 shows the relationship between the driving states of the first electromagnetic valve 266 and the second electromagnetic valve 268, the speed change direction of the CVT 14, and the changing speed of the speed change ratio γ. The first solenoid valve 266 and the second solenoid valve
In the speed change mode (C) in which both solenoid valves 268 are in the ON state, the hydraulic oil in the second line oil passage 82 is the bypass oil passage 29.
5, the throttle 296 and the check valve 29 provided in parallel
While being supplied to the secondary side hydraulic cylinder 56 through 8, the operating oil in the primary side hydraulic cylinder 54 is gradually discharged from a small gap that is positively or inevitably formed in the sliding portion of the piston. It is supposed to be done.

As described above, the secondary hydraulic cylinder 56 and the second line oil passage
Since the bypass oil passage 295 is provided between the valve 82 and 82, the pulsation that occurs in the hydraulic pressure P out in the secondary hydraulic cylinder 56 is preferably suppressed in synchronization with the duty driving of the flow rate control valve 264.
This is because the spike-like upper peak of the hydraulic pressure P out in the secondary side hydraulic cylinder 56 is released by the throttle 296, and the lower peak of P out is compensated through the check valve 298.

Here, the first line hydraulic pressure Pl 1 in the CVT 14 is as shown in FIG. 10 during positive drive travel (when the drive torque T is positive), and during engine brake travel (when the drive torque T is negative). At the time), the hydraulic pressure value as shown in Fig. 11 is desired. 10th
FIG. 11 and FIG. 11 both show the hydraulic pressure value required when the gear ratio γ is changed within the entire range in a state where the input shaft 30 is rotated with a constant shaft torque. .
In the present embodiment, since the primary side hydraulic cylinder 54 and the secondary side hydraulic cylinder 56 have the same pressure receiving area, the hydraulic pressure P in in the primary side hydraulic cylinder 54> the secondary side hydraulic cylinder 56 during forward drive traveling in FIG. The hydraulic pressure P out of FIG. 11 is P out > P in when the engine brake is running in FIG. 11, and both are hydraulic pressure inside the driving side hydraulic cylinder> hydraulic pressure inside the driven side hydraulic cylinder. Since the above-mentioned P in at the time of forward drive is to generate the thrust of the hydraulic cylinder on the driving side, the thrust for generating the target speed ratio γ can be generated in the hydraulic cylinder, and the power loss is reduced. to reduce, first line pressure Pl 1 is the pressure regulated to a value obtained by adding and sufficient margin hydraulic α required for the P in is desired. However, it is impossible to regulate the first line hydraulic pressure Pl 1 shown in FIG. 10 and FIG. 11 based on the hydraulic pressure in one hydraulic cylinder. Therefore, in the present embodiment, the first pressure regulating valve is Second plunger 148 on 100
Is provided, and the biasing force based on the higher hydraulic pressure of P in and the second line hydraulic pressure Pl 2 is transmitted to the spool valve element 140 of the first pressure regulating valve 100. This allows
For example, as shown in FIG. 12, during no-load traveling where the curve indicating P in and the curve indicating P out intersect, the first line hydraulic pressure Pl 1 is either P in or the second line hydraulic pressure Pl 2 . It is controlled to a value obtained by adding a margin value α to a high hydraulic pressure value. As a result, the first line hydraulic pressure Pl 1 is controlled to a necessary and sufficient value, and the power loss is minimized. By the way, the 12th
The first line hydraulic pressure Pl 1 ′ shown by the broken line in the figure is the second plunger 1
This is a case where 48 is not provided, and a large margin hydraulic pressure is unnecessarily generated in a range where the gear ratio γ is small.

The margin value α is a necessary and sufficient value for changing the gear ratio γ at a desired speed within the entire gear ratio change range of the CVT 14 to obtain the desired gear ratio γ, and is clear from the equation (2). As described above, the first line hydraulic pressure Pl 1 is increased in relation to the throttle pressure P th . The pressure receiving area of each part of the first pressure regulating valve 100 and the biasing force of the return spring 144 are set as such. At this time, the first pressure regulating valve 10
As shown in FIG. 13, the first line oil pressure Pl 1 adjusted by 0 increases according to P in or P out and the throttle pressure P th , but is saturated at the maximum value corresponding to the throttle pressure P th. It is supposed to be done. This allows
Even if the hydraulic pressure P in in the primary hydraulic cylinder 54 increases while the reduction ratio of the V groove width of the primary variable pulley 40 is mechanically prevented when the gear ratio γ reaches its maximum value, it is always more than that. Over-pressurization of the first line hydraulic pressure Pl 1 controlled to be increased by the margin value α is prevented.

In the first pressure regulating valve 100, the hydraulic oil flown out from the port 150b is guided to the lockup clutch pressure oil passage 92, and the lockup clutch pressure regulating valve 310 causes the fluid coupling 1 to operate.
The lock-up clutch hydraulic pressure P CL having a pressure suitable for operating the second lock-up clutch 36 is adjusted. That is, the lock-up clutch pressure regulating valve 310 receives the lock-up clutch hydraulic pressure P CL as the feedback pressure and urges the spool valve element 312 in the valve opening direction, and the spool valve element 312 in the valve closing direction. Spring 314 and a chamber 316 to which the clutch hydraulic pressure P CL is supplied through a lock-up sudden release valve 400 which will be described later at the time of sudden release
And a plunger 317 that biases the spool valve element 312 in the valve closing direction by receiving the hydraulic pressure of the chamber 316, and the spool valve element 312 balances the thrust force based on the feedback pressure and the thrust force of the spring 314. Thus, the hydraulic oil in the lock-up clutch pressure oil passage 92 is caused to flow out to generate a constant lock-up clutch hydraulic pressure P CL . Further, when the clutch oil pressure P CL is supplied to the chamber 316 at the time of sudden release, the clutch oil pressure P CL is increased in order to release the lockup clutch 36 more quickly. The hydraulic oil flown out of the lockup clutch pressure regulating valve 310 is sent out for lubrication of each part of the transmission through the throttle 318 and the lubricating oil passage 94, and is returned to the intake oil passage 78 of the oil pump 74.

Lockup clutch hydraulic pressure regulated as described above
P CL is selectively supplied to the engagement side oil passage 322 and the release side oil passage 324 of the fluid coupling 12 by the lockup control valve 320 so that the lockup clutch 36 is brought into the engaged state or the released state. Has become. That is, the lockup control valve 32
0 indicates that the lock-up clutch pressure oil passage 92 is the engagement side oil passage 3
A spool valve element 326 that is selectively connected to the 22 and the release side oil passage 324, and a spring 328 that biases the spool valve element 326 toward the release side are provided. The third line hydraulic pressure Pl 3 is introduced from the output port 256 of the manual valve 250 through the oil passage 257 to the upper end surface side (spring 328 side) of the spool valve element 326 only when the R range is selected. In the other ranges, the chamber 334 to be drained is provided, while on the lower end surface side (the exhaust spring 328 side) of the spool valve element 326, the signal pressure P SOL3 when the third solenoid valve 330 of the normally open type is ON. A chamber 332 into which is introduced is provided. When the third solenoid valve 330 is in the on state (closed state), a signal pressure P SOL3 equal to the clutch hydraulic pressure P CL is generated downstream of the throttle 331, but the third solenoid valve 330 is in the off state (open state). When, the downstream side of the throttle 331 is drained and the signal pressure P SOL3 is eliminated. Such aperture 331
And the solenoid valve 330 constitutes a means for generating the signal pressure P SOL3 , and the signal pressure P SOL3 is, in addition to the lockup control valve 320, a second line hydraulic pressure reduction control valve 380, a lockup sudden release valve 400, Each is supplied to the reverse inhibit valve 420.

Therefore, in the shift range other than the R range, when the third solenoid valve 330 is in the ON state, the signal pressure P is supplied to the chamber 322.
Although SOL3 is introduced, since the chamber 334 is set to the atmospheric pressure, the spool valve element 326 is positioned on the spring 328 side, so that the hydraulic oil in the lockup clutch pressure oil passage 92 engages with the engagement side oil passage 322. Is supplied to the lockup clutch 36 and the lockup clutch 36 is brought into an engaged state. On the contrary, when the third solenoid valve 330 is in the OFF state, the chamber 332 is set to the atmospheric pressure, so that the spool valve element 326 is positioned downward in FIG. 1 according to the biasing force of the spring 328. Lockup clutch pressure oil passage
The hydraulic oil in 92 is supplied to the release side oil passage 324, and the lockup clutch 36 is released. In addition, when the shift position is changed to the R range, the third position is set in the chamber 334.
Since the line oil pressure Pl 3 is supplied, the urging force of the third line oil pressure Pl 3 and the spring 328 is larger than the urging force of the spool valve element 326 based on the signal pressure P SOL3 .
Regardless of the open / closed state of the solenoid valve 330, the spool valve element 326 is preferentially positioned on the lower side of FIG. 1 and the lockup clutch 36 is released.

It should be noted that the hydraulic oil that flows out from the throttle 336 when engaged and the hydraulic oil that flows out from the lockup control valve 320 by being returned from the lockup clutch 36 via the engagement side oil passage 322 when disengaged are: After the pressure is adjusted to a certain value or less by a cooler hydraulic control valve 338, the oil is returned to an oil tank (not shown) via an oil cooler 339.

The back pressure control of the accumulators 342 and 340 provided in the forward clutch 72 and the reverse brake 70, respectively, will be described. Lockup clutch Pressure oil passage 92 throttling 344
The hydraulic oil that leaked through the
Controlled by the solenoid valve 346, the hydraulic pressure is changed with respect to the duty ratio D s4 as shown in FIG. That is, the throttle 344 and the fourth solenoid valve 346 function as a signal pressure generating unit that generates the signal pressure P SOL4 . In this way, the signal pressure P SOL4 regulated by the drive duty ratio D s4 of the fourth solenoid valve 346 is applied to the solenoid pressure switching valve via the oil passage 348.
Guided to 350. When the shift position is in the P, R, and N ranges, the hydraulic cylinder of the forward clutch 72 is drained by the manual valve 250, so that the solenoid pressure switching valve 350 outputs the signal pressure P SOL4 through the oil passage 354 to the first position. Four
The hydraulic pressure in the oil passage 356 is drained while allowing the pressure regulator valve 170 to be applied. However, when shifting from the N range to the D, S, and L ranges, the forward clutch 72
When the hydraulic pressure in the hydraulic cylinder is initial, the accumulator
Due to the relaxing action of 342, it rises with the passage of time according to a predetermined function and rises to the third line hydraulic pressure Pl 3 simultaneously with the engagement. From this, before engagement of the forward clutch 72, the signal pressure P SOL4 in the oil passage 348 is equal to the solenoid pressure switching valve 350.
When the forward clutch 72 is engaged, the solenoid pressure switching valve 350 drains the oil passage 354 and the signal pressure P in the oil passage 348.
SOL4 is allowed to be led to the second line hydraulic pressure reduction control valve 380 and the lockup sudden release valve 400 via the oil passage 356.

Here, the back pressure control of the accumulators 340 and 342 is N → D
This is performed in order to reduce the shift shock (engagement shock) at the time of shift and N → R shift. When the clutch is engaged, the increase in the hydraulic pressure in the hydraulic cylinder is suppressed for a minute time to reduce the shock. Therefore, the fourth line hydraulic pressure Pl 4 controlled by the fourth pressure regulating valve 170 is changed to the back pressure port 366 of the accumulator 342 for the forward clutch 72 and the back pressure port 368 of the accumulator 340 for the reverse brake 70 by changing the fourth line hydraulic pressure Pl 4 . The oil is supplied through the 4-line oil passage 370, and the accumulator 342,
Controls the action of the hydraulic pressure change mitigation by 0

The fourth pressure regulating valve 170 includes a spool valve element 171 that opens and closes between the first line oil passage 80 and the fourth line oil passage 370, and a spring 172 that biases the spool valve element 171 in the valve opening direction. I have it. The first land 173 and the second land of the spool valve 171
A chamber 176 is provided between the land 174 and the land 174 to introduce the fourth line hydraulic pressure Pl 4 through the throttle hole 175 to act as a feedback pressure. On the other hand, on the end surface of the spool valve element 171 on the spring 172 side, Signal pressure P acting in the valve opening direction
A chamber 177 for introducing SOL4 is provided, and the end surface of the spool valve element 171 on the non-spring 172 side is open to the atmosphere. In the fourth pressure regulating valve 170 thus configured, the spool valve element 171
However, the urging force in the valve closing direction based on the feedback pressure corresponding to the fourth line hydraulic pressure Pl 4 and the urging force in the valve opening direction by the spring 172 and the valve opening direction based on the signal pressure P SOL4 are balanced. As a result, the fourth line hydraulic pressure Pl 4 is adjusted to a pressure corresponding to the signal pressure P SOL4 . That is, during N → D shift and N → R shift, the signal pressure P SOL4 is passed through the solenoid pressure switching valve 350 and the fourth pressure regulating valve
While being supplied to 170, as shown in FIG.
Since the line hydraulic pressure Pl 4 is controlled to a value corresponding to the duty ratio D s4 of the fourth solenoid valve 346, the fourth solenoid valve is generated so as to generate a back pressure suitable for reducing shift shock (engagement shock). 346 is duty driven. Further, the hydraulic pressure in the forward clutch 72 rises to the third line hydraulic pressure Pl 3 , so that the signal pressure P supplied to the fourth pressure regulating valve 170 is increased.
When SOL4 there is in the chamber 177 is blocked by the solenoid pressure switching valve 350 is opened to the atmosphere, the fourth line pressure Pl 4 is relatively low 4 kg / c corresponding to the urging force of the valve opening direction of the spring 172
It is controlled to a constant pressure of about m 2 . The fourth line hydraulic pressure Pl 4 adjusted to this constant pressure is exclusively used for the speed change direction switching valve 262.
And used as a driving oil pressure for the flow control valve 264. The accumulator 372 provided in the oil passage 354 is connected to the signal pressure P SOL4 related to the duty driving frequency of the fourth solenoid valve 346.
It is for absorbing the pulsation of.

Returning to FIG. 1, the second line hydraulic pressure reduction control valve 380 controls the influence of the centrifugal hydraulic pressure generated in the secondary hydraulic cylinder 56 in order to approximate the second line hydraulic pressure Pl 2 to the optimum control pressure P opt. The signal pressure P generated by the fourth solenoid valve 346 when the second line oil pressure Pl 2 is reduced to prevent it
SOL4 acts on the chamber 136 of the second pressure regulating valve 102. The second line hydraulic pressure reduction control valve 380 has a port 382a communicating with the oil passage 356,
Between the port 382b and the drain port 382c communicating with the hydraulic chamber 136 of the second pressure regulating valve 102 via the oil passage 384, and between the first position which is the upper end of the movement stroke and the second position which is the lower end of the movement stroke. A spool valve element 386 that is slidably disposed and a spring 388 that biases the spool valve element 386 toward the second position are provided. Therefore, when the third solenoid valve 330 is in the off state (open state), the pressure inside the chamber 390 is exhausted, the spool valve 386 is positioned at the second position, and the ports 382b and 382c communicate with each other so that the second pressure regulating valve is opened. Since the inside of the hydraulic chamber 136 of 102 is drained, the second line hydraulic pressure Pl 2 is controlled according to the equation (1). However, when the third solenoid valve 330 is in the ON state (closed state), the signal pressure P SOL3 (clutch pressure P CL ) is introduced into the chamber 390 on the lower end side of the spool valve 386, and the spool valve 386 is in the first position. And the ports 382a and 382b are communicated with each other. At this time, when the forward clutch 72 is in the engaged state, the signal pressure P SOL4 generated corresponding to the drive duty ratio of the fourth solenoid valve 346 causes the oil passages 348, 356, the ports 382a, 382b, and the oil passage 384. Is supplied into the hydraulic chamber 136 of the second pressure regulating valve 102 via. This clutch pressure P CL is the second pressure regulating valve
Since the spool valve element 110 of 102 is biased in the valve closing direction, the second line hydraulic pressure Pl 2 is regulated according to the following equation (3), and as shown by the alternate long and short dash line in FIG. It is made lower than the second line hydraulic pressure Pl 2 . The third solenoid valve 3
Even if 30 is in the on state, if the fourth solenoid valve 346 is in the off state, the second line hydraulic pressure Pl 2 is controlled as usual according to the above equation (1).

Pl 2 = (A 4 · P th + W−A 1 · P r − (A 2 −A 1 ) · P CL ] / (A
3- A 2 ) (3) Next, the lock-up sudden release valve 400 provided to improve the release response of the lock-up clutch 36 is the port 402a communicating with the clutch pressure oil passage 92 and the lock-up. Port 402b communicating with hydraulic chamber 316 at the end surface of plunger 317 of clutch pressure regulating valve 310 via oil passage 404, drain port 402
c, and a port 402d communicating with the engagement side oil passage 322 to the lockup clutch 36, and a slidable arrangement between a first position which is the upper end and a second position which is the lower end of the moving stroke. A spool valve element 406 and a spring 408 that biases the spool valve element 406 toward the second position are provided.
In the chamber 410 on the lower end side of the spool valve element 406, when the forward clutch 72 is engaged, the clutch pressure P CL is introduced when the fourth solenoid valve 346 is in the on state, and is discharged when it is in the off state. It Further, the chamber 412 on the upper end side (spring 408 side) of the spool valve element 406 is introduced with the signal pressure P SOL3 (clutch pressure P CL ) when the third solenoid valve 330 is in the on state, and is discharged when it is in the off state. Is pressed. The lockup sudden release valve 400 is the third solenoid valve 330 and the fourth solenoid valve described above.
Although controlled by 346, the spool valve element 406 is placed in the first position and the clutch pressure P CL is set to the port 402a only when the third solenoid valve 330 is off and the fourth solenoid valve 346 is on. , The port 402b, and the oil passage 404 to be guided to the hydraulic pressure chamber 316 of the clutch pressure regulating valve 310 to increase the clutch pressure P CL, and at the same time, through the engagement-side oil passage 322, the fluid coupling 1
The hydraulic oil discharged from the second engagement side oil chamber 33 is drained from the upstream side of the cooler 339 via the ports 402d and 402c, so that the lockup clutch 36 is rapidly released. In the other states of the third solenoid valve 330 and the fourth solenoid valve 346, the spool valve element 406 is located at the second position. At this time, the lockup sudden release valve 400 not only reduces the flow resistance of the working oil discharged from the engagement side oil chamber 33 of the fluid coupling 12, but also releases the fluid coupling 12 by the lockup clutch pressure regulating valve 310. Since the clutch pressure P CL supplied to the side oil chamber 35 is increased, a high release response of the lockup clutch 36 is obtained.

The reverse inhibit valve 420, which is provided to prohibit reverse during forward traveling, is
Is in the R range, the port 422a to which the third line hydraulic pressure Pl 3 is supplied from the output port 256, the reverse brake 7
Port 422 communicating with 0 hydraulic cylinder via oil passage 423
b, the drain port 422c, the spool valve element 424 slidably disposed between the first position, which is the upper end of the movement stroke, and the second position, which is the lower end, and the spool valve element 424.
And a spring 426 for urging the valve toward the first position. The chamber 428 on the upper end side of the spool valve element 424 has a third
The signal pressure P SOL3 (clutch pressure P CL ) is introduced through the oil passage 430 when the solenoid valve 330 is in the on state, and is exhausted when the solenoid valve 330 is in the off state. In the chamber 432 on the other end side (spring 426 side) of the spool valve element 424, the manual valve 250 is
When in the S, L range, the third line hydraulic pressure Pl 3 is introduced from its output port 258. In the reverse inhibit valve 420 thus configured, the third line hydraulic pressure Pl 3 in the chamber 432 is discharged and the signal pressure P 3 is released to the chamber 428.
When the spool valve element 424 is positioned at the second position (lower end) by introducing SOL3 (clutch pressure P CL ), the communication between the ports 422a and 422b is cut off, and the reverse brake 70 is actuated. Oil supply cut off and port
By connecting the 422c and the port 422b to each other, the hydraulic oil in the hydraulic cylinder of the reverse brake 70 is drained, so that the forward / reverse switching device 16 is prohibited from switching to the reverse direction. Therefore, the shift lever is
When the 252 is mistakenly operated from the D range to the N range to the R range, the electronic control unit 460, which will be described later, turns on the third solenoid valve 330 to bring the forward / reverse switching device 16 into the neutral state. It

When the shift position is in the N or P range, the first line hydraulic pressure reduction control valve 440 provided to reduce the first line hydraulic pressure Pl 1 by a predetermined pressure to suppress the belt noise is the drain port 442a and the first adjustment valve. Port 422b communicating with chamber 160 between first land 152 and second land 159 of pressure valve 100 via oil passage 161, and port 442c communicating with second line oil passage 82.
A plunger 444, a spool valve 446 that opens and closes between the second line oil passage 82 and the chamber 160 of the first pressure regulating valve 100, and a spring 448 that biases the spool valve 446 in the valve opening direction. There is. The chamber 450 on the lower end surface of the plunger 444 is communicated with the output port 258 of the manual valve 250 that outputs the third line hydraulic pressure Pl 3 in the forward drive range, and the chamber between the plunger 444 and the spool valve 446. 452 is in communication with the output port 256 of the manual valve 250 that outputs the third line hydraulic pressure Pl 3 in the R range. Therefore, in the D, S, L, and R ranges, the spool valve 446 is positioned at the upper end, the inside of the chamber 160 of the first pressure regulating valve 100 is set to the atmospheric pressure through the drain port 442a, and the first line hydraulic pressure Pl 1 is set to the above ( The pressure is adjusted to a normal value according to the equation (2). But,
In the N and P ranges, the spool valve 446 is located at the lower end, and the second line hydraulic pressure Pl 2 is supplied into the chamber 160 of the first pressure regulating valve 100. Therefore, the spool valve 1 of the first pressure regulating valve 100
Since 40 is urged in the valve opening direction based on the second line hydraulic pressure Pl 2 acting in the chamber 160, the first line hydraulic pressure Pl 1 is reduced. As a result, the clamping force on the transmission belt 44 is reduced as much as possible within the range where slippage does not occur, the noise level of the belt is reduced, and the durability of the transmission belt 44 is enhanced.

In FIG. 2, the electronic control unit 460 functions as the control means of the present embodiment, and is the first solenoid valve 266, the second solenoid valve 268, and the third solenoid valve in the hydraulic control circuit of FIG.
By driving 330 and the fourth solenoid valve 346, the gear ratio γ of the CVT 14 and the lockup clutch 36 of the fluid coupling 12 are controlled. The electronic control unit 460 includes a so-called microcomputer including a CPU, RAM, ROM, etc., in which a vehicle speed sensor 462, a CVT for detecting the rotation speed of the drive wheels 24, a CVT.
An input shaft rotation sensor 464 and an output shaft rotation sensor 466 that detect the rotation speeds of the 14 input shafts 30 and the output shafts 38, respectively.
A throttle valve opening sensor 468 for detecting the opening of a throttle valve provided in the intake pipe of the engine 10, the shift lever 2
Operation position sensor 470 for detecting the operation position of 52, brake switch for detecting the operation of the brake pedal
472, from the engine rotation sensor 474 for detecting the rotation speed of the engine 10, a signal representing the vehicle speed V, a signal representing the input shaft rotation speed N in , a signal representing the output shaft rotation speed N out, and a throttle valve opening θ th. Is supplied, a signal representing the operation position P s of the shift lever 252, a signal representing the brake operation, and a signal representing the engine rotation speed N e , respectively. The CPU in the electronic control unit 460 uses the temporary storage function of RAM while ROM
The input signal is processed according to a program stored in advance in the first solenoid valve 266, the second solenoid valve 268, and the third solenoid valve 33.
0, outputs a signal for driving the fourth solenoid valve 346.

In the electronic control unit 460, initialization is executed when the power is turned on, and then a main routine (not shown) is executed to read the input signal from each sensor, and input based on the read signal. Axis 30
Rotation speed N in , rotation speed N out of the output shaft 38, gear ratio γ of the CVT 14, vehicle speed V, etc. are calculated, and lockup control of the lockup clutch 36, belt clamping pressure optimization control, The shift control of the CVT 14 and the like are executed sequentially or selectively.

In the lockup control, for example, when the vehicle speed V exceeds a predetermined engagement determination reference value, the third solenoid valve 330 is turned on and the lockup clutch 36 is engaged, but the vehicle speed V is predetermined. If the release reference value falls below or other conditions such as a braking operation are satisfied, the third solenoid valve 330 is turned off and the lockup clutch 36 is turned off.
Is released.

In the shift control of the CVT 14, for example, the engine 1
The target input shaft rotation speed N in * is determined based on the actual throttle valve opening θ th and the vehicle speed V from a relationship obtained in advance so that the optimum fuel consumption rate and drivability are optimally set to 0. Any one of the shift modes shown in FIG. 9 is determined so that the shaft rotation speed N in * and the actual input shaft rotation speed N in coincide with each other, and the first solenoid valve 26 corresponding to that mode is selected.
6 and the second solenoid valve 268 are driven. Further, in this shift control, the control mode of FIG. 17 is selected according to a predetermined condition, and the third solenoid valve 330 and the fourth solenoid valve 330 corresponding to the selected mode are selected.
The solenoid valve 346 is driven.

In the belt clamping pressure control, the belt clamping pressure optimization routine shown in FIG. 18 is executed. First, in step S1 of the figure, the output torque T e of the engine 10 is calculated based on the actual engine rotation speed N e and the throttle valve opening θ th from the prestored relationship shown in FIG. 19, for example. In the following step S2, the theoretical value P theory of the second line hydraulic pressure for obtaining an ideal pinching force which is as small as possible and can transmit torque without the transmission belt 44 slipping is stored in advance in the following equation (4). From the relationship, it is calculated based on the actual gear ratio γ (= N in / N out ) calculated in the step (not shown) and the output torque T e calculated in step S1. This relationship is the same as the theoretical formula described in, for example, JP-A-60-53258, and for example, CVT14
The input torque is T in , the friction coefficient of the transmission belt 44 is μ, the wedge angles of the variable pulleys 40 and 42 are α, the hanging diameter of the variable pulley 42 is D out , and the margin ratio is K ′, the slip of the transmission belt 44 is The thrust force W out of the secondary side hydraulic cylinder 56 that does not generate is given by the following equation (4) ′. The above equation (4) approximates (4) ′ to obtain the thrust force W out and It is obtained by dividing by the pressure receiving area of the cylinder 56.

P theory = K 1 (1 + γ) ・ T e …… (4) However, K 1 is a constant.

In the following step S3, the centrifugal hydraulic pressure P cfg generated in the secondary hydraulic cylinder 56 is calculated based on the actual output shaft rotation speed N out from the following equation (5) stored in advance. 20th
In the figure, the solid line shows the basic output pressure P mec of the second pressure regulating valve 102 when the speed ratio γ is the minimum value γ min , and the broken line is necessary for the clamping pressure of the transmission belt which is reduced by the generation of the centrifugal hydraulic pressure P cfg. Pressure P opt , and the difference between the broken line and the solid line is a surplus (excessive) hydraulic pressure and corresponds to the magnitude of the centrifugal hydraulic pressure P cfg .

In step S4, in order to reduce the above-mentioned surplus (excess) hydraulic pressure as much as possible, the following formula (6) stored in advance is used:
Theoretical value P of the second line hydraulic pressure Pl 2 calculated in step S2
The optimum control pressure P opt is calculated based on the theory and the centrifugal hydraulic pressure P cfg calculated in step S3.

P cfg = K 2 · N out 2 (5) However, K 2 is a constant.

P opt = P theory −P cfg (6) In step S5, the basic output pressure P of the second pressure regulating valve 102 is set.
mec , that is, a mechanical set pressure mechanically determined from the configuration of the chamber 136 of the second pressure regulating valve 102 in a state where the signal pressure P SOL4 is not supplied is stored in advance from the actual gear ratio γ and the throttle valve opening degree. It is calculated based on θ th . This relationship is shown by the solid line in FIG. 21, for example, and is stored in the form of a data map as shown in the following expression (7).

P mec = map (γ, θ th ) ... (7) FIG. 22 is an example showing the contents of the above step S5,
9 shows a mechanical set pressure estimation routine for accurately calculating the basic output pressure P mec of the second pressure regulating valve 102 with a small program capacity. In the figure, in step S5-1 corresponding to the gear ratio determining means, the gear ratio γ of the CVT 14 is calculated based on the actual input shaft rotation speed N in and output shaft rotation speed N out , and in step S5-2, Logarithm of ratio log
γ is calculated from an approximate expression (8) stored in advance. 23rd
The figure shows the value calculated by this approximation formula (8) and the logarithmic value logγ
Shows a high correlation with.

In step S5-3, the gear ratio specific pressure P r, which is the break point of the solid line in FIG. 21, is calculated from the following pre-stored equation (9) representing the straight line indicated by the chain double-dashed line in FIG. This equation (9) also shows the characteristic of FIG.

P r = a 1 log γ + b 1 (9) Then, in steps S5-4 and S5-5, the pressure P1 corresponding to the actual running state is determined from the equation (10) showing the straight line rising to the right in FIG. To be done. That is, first, step S5-
In Fig. 4, the y-axis (pressure axis) intercept value b 2 of the equation (10) corresponding to the actual traveling state is the actual throttle valve opening degree from the relationship of the following equation (11) stored in advance in Fig. 25, for example. is determined based on the theta th, in step S5-5, the pressure P1 corresponding to the actual running state on the basis of the logarithmic values logγ and intercept value b 2 of the actual gear ratio γ from a pre-stored relationship (10) It is calculated. This pressure P1 is a value in the upward-sloping portion of the basic output characteristic of the second pressure regulating valve 102.

P1 = a 2 log γ + b 2 (10) b 2 = map2 (θ th ) (11) In step S5-6, the pressure value and the throttle valve opening in the solid line parallel to the horizontal axis in FIG. The pressure P2 corresponding to the actual traveling state is determined based on the actual throttle valve opening θ th from the relationship with the degree θ th , for example, the relationship (12) stored in advance shown in FIG. This pressure P2 is the second pressure regulating valve 102.
It is a value in a portion of the basic output characteristics of the above which is parallel to the horizontal axis.

P2 = map1 (θ th ) (12) Then, in step S5-7, the pressure P1 obtained in step S5-5 is equal to or higher than the gear ratio specific pressure P r obtained in step S5-3. It is judged whether it is a value or not, and the pressure
If it is determined that P1 is equal to or greater than the gear ratio specific pressure P r , the basic output pressure P mec of the second pressure regulating valve 102 is determined in step S5-8.
Is the above-mentioned pressure P1, but if it is determined that the pressure P1 is smaller than the gear ratio specific pressure P r , the basic output pressure P mec is set to the pressure P2 in step 5-9.

Returning to FIG. 18, in the subsequent step S6, the second pressure regulating valve 10
The difference between the second basic output pressure P mec and the optimum control pressure P opt , that is, the reduced hydraulic pressure value P down that decreases from the basic output pressure P mec is calculated from the following equation (13). That is, in this step S6, the difference between the basic output pressure P mec of the second pressure regulating valve 102 shown by the solid line in FIG. 21 and the ideal curve P opt shown by the broken line is calculated as the reduced hydraulic pressure value P down . is there.

Next, at step S7, the drive signal (duty ratio) I SOL4 of the fourth solenoid valve 346 for canceling the lowered hydraulic pressure value P down is calculated from the prestored relationship (14) shown in FIG. 26, for example. Determined based on P down .

P down = P mec -P opt (13) I SOL = g (P down ) (14) Then, in step S8, the drive signal I SOL4 is output.

By repeatedly performing the above operation in the forward traveling state in which the lockup clutch 36 of the vehicle is engaged, the second line hydraulic pressure Pl 2 is accurately approximated to the ideal curve shown by the broken line in FIG. Oil pump 74
The power loss for driving the motor is reduced as much as possible.

As described above, according to the present embodiment, the basic output pressure P mec and the optimum control pressure P determined by the mechanical structure of the second pressure regulating valve 102 are set.
In order to eliminate the difference P down from opt , in other words, the difference
Optimal control pressure P opt that has decreased from the basic output pressure P mec by P down
So that is output from the second pressure regulating valve 102.
The fourth solenoid valve 346 is driven by the drive signal I SOL4 by 60, so that the second pressure regulating valve 102 sets the second line hydraulic pressure Pl 2 based on the signal pressure P SOL4 generated by the fourth solenoid valve 346. By adjusting the pressure, the second line pressure Pl 2 is approximated to the ideal curve with high accuracy. As a result, the clamping pressure on the second line hydraulic pressure Pl 2 and the transmission belt 44 based on the hydraulic pressure is set to a necessary and sufficient value, so that the power loss for driving the oil pump 74 is preferably improved. In the present embodiment, the second line hydraulic pressure Pl 2 corresponds to the tension control pressure of the transmission belt 44.

Further, in the electronic control unit 460, the logarithmic value logγ of the gear ratio γ is obtained, and the logarithmic value logγ of the gear ratio γ, the slot valve opening (required output value) θ th, and the basic output pressure P of the second pressure regulating valve 102. Pre-stored relationship with mec [map (γ,
θ th )] based on the logarithmic value log γ of the actual gear ratio γ and the throttle valve opening (requested output value) θ th
Since the basic output pressure P mec is calculated by 102, there is an advantage that the program capacity is significantly reduced as compared with the case where two-dimensional map interpolation is used to improve control accuracy. That is, when the gear ratio axis of the output characteristic of the pressure regulating valve is a uniform scale axis, the line showing the tension control pressure of the pressure regulator valve is a straight line until it reaches the gear ratio specific pressure P r , but after reaching it, it becomes a curve, If a curve is to be estimated by two-dimensional map interpolation, a huge number of grid points will be required. Incidentally, FIG. 27 shows the basic output pressure P mec of the second pressure regulating valve 102 on the axis representing the gear ratio γ on a uniform scale. As is clear from the figure, in a region where the basic output pressure P mec is higher than the gear ratio specific pressure P r , a logarithmic curve is obtained, and if this is attempted to be approximated by ordinary two-dimensional map interpolation, a large number of grid points (data points) It is necessary to represent the change curve of the basic output pressure P mec by using, and the storage capacity of the data points is large and the calculation becomes complicated. On the other hand, in the present embodiment, as shown in FIG. 21, the axis representing the logarithmic value logγ of the gear ratio γ is adopted, so that it is represented by a bending characteristic composed of two straight lines and a small number of data points. Since the basic output pressure P mec represented by is stored in the ROM in advance, the storage capacity is significantly small and the calculation is simple.

Moreover, according to the present embodiment, since the fourth solenoid valve 346 is configured in the normally open type, even if the fourth solenoid valve 346 is used.
Even if the disconnection failure of 346 occurs, the second line hydraulic pressure Pl 2 increases by the decreased hydraulic pressure value P down and the basic output pressure of the second pressure regulating valve 102 increases.
Since it only returns to P mec , there is an advantage that the transmission belt 44 does not slip and the vehicle can run without problems.

Next, FIG. 28 shows a detailed flow chart of the step S7 for determining the drive signal I SOL4 of the fourth solenoid valve 346.
In step S7-1 of the figure, it is determined whether or not the optimum control pressure P opt is equal to or higher than the gear ratio specific pressure P r calculated in step S5-3, and in step S7-2, the second control is performed. It is determined whether or not the basic output pressure P mec of the pressure valve 102 is less than or equal to the gear ratio P r . If it is determined in step S7-1 that the optimum control pressure P opt is greater than or equal to the transmission specific pressure P r , in step S7-3 the signal pressure is calculated from the following equation (15).
When P SOL4 is determined, it is determined in step S7-1 that the optimum control pressure P opt is smaller than the gear ratio P r and the basic output pressure P mec is determined in step S7-2.
When it is determined that it is equal to or lower than P r , the signal pressure I SOL4 is determined from the following equation (16) in step S7-4, and the optimum control pressure P opt is determined from the gear ratio specific pressure P r in step S7-1. If it is determined to be small and it is determined in step S7-2 that the basic output pressure P mec is higher than the gear ratio specific pressure P r , in step S7-5 the signal pressure is calculated from the following equation (17). P SOL4 is decided.

P SOL4 = (A 3 ′ / A 2 ′) P down …… (15) P SOL4 = [(A 1 + / A 3 ′) / A 2 ′]) P down …… (16) P SOL4 = (A 3 ′ / A 2 ′) P down + (A 1 / A 2 ′) (P r-
P opt ) (17) However, A 3 ′ = A 3 −A 2 and A 2 ′ = A 2 −A 1 .

Here, considering the signal pressure P SOL4 applied to the chamber 136, the second pressure regulating valve 102 operates so that the following equation (19) is established. Therefore, the signal pressure P SOL4 applied to the chamber 136 to obtain the target control pressure by lowering it from the basic output pressure P mec
And the reduced pressure P actually reduced by this signal pressure P SOL4
The relationship with down is derived from equation (19), and
As shown in A of Fig. 21, the optimum control pressure P opt is the gear ratio specific pressure.
Although the above equation (15) If it is P r above, FIG. 21 of the basic output pressure P mec as shown in B is smaller than the speed ratio pressure P r when the above equation (16), and 21 As shown in C of the figure, the optimum control pressure P opt is smaller than the transmission specific pressure P r and the basic output pressure P
When mec is a larger than the speed ratio pressure P r, by adding a portion and a lower portion above the speed ratio pressure P r, the above equation (17)
Becomes

Pl 2 = (A 4 · P th + WA 1 · P r ) / (A 3 -A 2 ) ... (19) In the following step S7-6, the drive signal I of the fourth solenoid valve 346 is generated.
From the pre-stored relationship of FIG. 29 between SOL4 and the signal pressure P SOL4 generated thereby, the above steps S7-3 to
The duty ratio of the drive signal I SOL4 for obtaining the signal pressure P SOL4 determined in any of S7-5 is determined. Then, in step S8, the drive signal I SOL4 is output.

In the present embodiment, as described above, the fourth solenoid valve 346 is driven so that the reduced pressure P down from the basic output pressure P mec is obtained, and the second line hydraulic pressure Pl 2 becomes the optimum control pressure P opt . since the approximated, as shown in FIG. 28, according to the area to reduce the basic output pressure P mec (15) equation (16), the signal pressure P based on the relationship that is selected from (17) Since SOL4 is determined, there is an advantage that the tension of the transmission belt 44 can be controlled more accurately.

Next, another embodiment of the present invention will be described. In the following embodiments, the same parts as those described above will be designated by the same reference numerals and the description thereof will be omitted.

Instead of step S2 in FIG. 18, the routine shown in FIG. 30 may be executed. According to this, the theoretical target value P theory of the second line hydraulic pressure Pl 2 is obtained even in reverse (non) drive traveling, and the second line hydraulic pressure Pl 2 is suitably controlled based on this. That is, in step S2-1, it is determined whether or not the engine output torque T e is negative. If it is determined in step S2-1 that the output torque T e of the engine is not negative, the vehicle is in the positive drive traveling state, so step S2-2 similar to step S2 described above is executed to set the theoretical target value. P theory is calculated. However, if it is determined in step S2-1 that the output torque T e of the engine is negative, it means that the vehicle is in the reverse drive traveling (engine braking traveling) state. Therefore, from the relationship stored in advance in step S2-3. Based on the actual gear ratio γ, the primary hydraulic cylinder 54 and the secondary hydraulic cylinder during reverse drive travel
A thrust ratio γ of 56 is calculated. FIG. 31 shows the thrust W in required for the primary hydraulic cylinder 54 and the secondary hydraulic cylinder 56 in order to prevent the transmission belt 44 from slipping when a predetermined input torque T in is input to the CVT 14. Thrust W
thrust ratio between the out I the (= W out / W in) , the time of forward drive traveling, at no load, shows for each of the reverse drive traveling,
The previously stored relationship is the speed change ratio γ during reverse driving.
And the thrust ratio γ.

According to the present embodiment, the transmission belt is used even during reverse driving.
Since the second line hydraulic pressure Pl 2 is accurately controlled so that the tension of 44 becomes a necessary and sufficient value, not only the power loss is suppressed but also the transmission belt 44 when the engine brake is running.
Noise is suppressed and the durability of the transmission belt 44 is enhanced.

Further, as shown in FIG. 32, in the vehicle in which the third solenoid valve 330 is exclusively used for the engagement control of the lockup clutch 36, the optimum control pressure P is set between the steps S4 and S5 of FIG. opt, that is, the first line pressure Pl 2 is increased by a predetermined pressure to surely prevent the transmission belt 44 from slipping when the vehicle starts.
The routine shown in FIG. 33 may be inserted. Step S9 in the figure
In -1, it is determined whether or not the accelerator pedal (not shown) is returned, based on whether or not the idle switch LL (not shown) is in the ON state, and step S9
Whether the vehicle speed V is not greater than the predetermined criterion value V o is determined in -2. The criterion value V o, which corresponds to the period of increasing the second line pressure Pl 2 to prevent slippage of the driving belt 44 at the start of the vehicle, for example, a value of 2 to 5km / h is employed It When it is determined that the accelerator pedal is returned in step SS9-1, or when the vehicle speed V is determined to be greater than the determination reference value V o even if no accelerator pedal is returned, step S9 -3 or S9-5, the correction value ΔP is set to zero. However, the above steps S9-1 and
In S9-2, the accelerator pedal is depressed and the vehicle speed V
When it is determined that is equal to or smaller than the determination reference value Vo, the content of the correction value ΔP is set to a predetermined pressure increase correction value α in step S9-4. Then, in step S9-6, the correction value ΔP is added to the content of the optimum control pressure P opt .

According to this embodiment, since the second line hydraulic pressure Pl 2 is increased when the vehicle starts, the transmission belt 44 when starting the vehicle
Slippage is eliminated. That is, the optimum control pressure P opt basically determined by the equation (4) is effective in the state of continuously rotating, but the variable pulleys 40 and 42 are stationary, so that the friction of the transmission belt 44 is reduced. Although it is not sufficient when the coefficient μ is apparently small, the second line hydraulic pressure Pl 2 is increased by a predetermined pressure α between the stationary state and the rotation of the variable pulleys 40 and 42. The slippage of the transmission belt 44 when the vehicle starts is suitably eliminated. FIG. 34 is a time chart showing changes in the throttle valve opening θ th , the second line hydraulic pressure Pl 2 , and the vehicle speed V when the vehicle starts, and the second line hydraulic pressure Pl 2 is only a predetermined pressure α when the vehicle starts. Has been elevated. In the figure, the broken line is the target optimum control pressure P opt .

Incidentally, instead of step S9-2 in FIG. 33, as shown in FIG. 35, step S9-7 for determining a predetermined period from the start of the vehicle based on the gear ratio γ of the CVT 14 is provided. Good. In step S9-7, it is determined whether or not the gear ratio γ is equal to or greater than the prestored determination reference value γ o . If the gear ratio γ is not equal to or greater than the determination reference value γ o , step S9-
5 is executed, but if it is above, it means that the vehicle is starting, so step S9-4 is executed. The gear ratio γ of the CVT 14 is controlled based on the throttle valve opening θ th and the vehicle speed V from the relationship stored in advance as described above, and is reduced from the maximum value γ max when the vehicle starts from a stopped state. Therefore, the judgment reference value γ o is set to the second line hydraulic pressure.
A value is selected so that slip does not occur in the transmission belt 44 even when Pl 2 becomes the optimum control pressure P opt .

FIG. 36 shows a main part of a hydraulic control circuit according to another embodiment of the present invention. In the hydraulic control circuit of this embodiment, the above-mentioned first
Compared with the illustrated embodiment, the second line oil pressure reduction control valve 38
0 is the first relay valve 380 controlled by the fourth solenoid valve 346
In addition, the first line oil pressure reduction control valve 440 is changed to the second relay valve 440 controlled by the third solenoid valve 330, and
The first pressure regulating valve 100, the second pressure regulating valve 102, the shift direction switching valve 262,
The configurations of the lock-up control valve 320, the lock-up sudden release valve 400 and the reverse inhibit valve 420 are partially changed, and further, the solenoid pressure switching valve 350 and the accumulator 372 are removed, and the linear valve 500 is newly provided. ing. Therefore, the electronic control unit 460 causes the third
One of the control modes shown in the figure can be selectively implemented.

As shown in detail in FIG. 38, the second pressure regulating valve 102 of the present embodiment.
The plunger 116 has a land 117 having a cross-sectional area A 4 on which the throttle pressure P th acts and a cross-sectional area larger than that.
A land 119 provided with A 4 ′ is formed, and an output signal pressure P SOLL for increasing the second line hydraulic pressure Pl 2 by a predetermined pressure is provided in the chamber 133 between the lands.
It comes to be supplied from. In addition, the second pressure regulating valve 10
The output signal pressure P SOLL is supplied to the second chamber 136 from the port 382c of the first relay valve 380. For this reason,
When the second line hydraulic pressure Pl 2 of this embodiment is reduced by the output signal pressure P SOLL of the linear valve 500, the following equation (20)
Determined according to.

Pl 2 = [A 4 · P th + WA 1 · P r- (A 2 -A 1 ) ・ P SOLL ] / (A 3 -A 2 ) ……
(20) The linear valve 500 controls the output signal pressure P in relation to the voltage value or current value of the drive signal supplied from the electronic control unit 460.
The SOLL is configured to be continuously changed.
That is, as shown in detail in FIG. 39, the linear valve 500
Has a pressure reducing valve type valve mechanism, and
A spool valve 502 operated to regulate the output signal pressure P SOLL from Pl 4, a linear solenoid 504 excited by a drive signal supplied from an electronic control unit 460, and an excitation state of this linear solenoid 504. The core 506 that urges the spool valve 502 toward the boost side, and the spool valve 50
The spring 508 that urges 2 to the buck side and the spool valve 50
A feedback oil chamber 510 to which the output signal pressure P SOLL is guided in order to bias 2 toward the step-down side. The spool valve 502 is operated so as to move to a position where the biasing force applied from the core 506 to the pressure increasing side and the biasing force applied from the spring 508 and the feedback oil chamber 510 to the pressure reducing side are balanced. As a result, according to the output characteristics shown in FIG. 40, the drive signal supplied from the electronic control unit 460,
For example the drive current output signal pressure P SOLL in response to changes in I SOLL
Is continuously changed. Thus, the linear valve 500
The output signal pressure P SOLL obtained by reducing the fourth line hydraulic pressure Pl 4 supplied to the input port 512 of the first relay valve 380 is output from the output port 514 of the linear valve 500.
Since the fourth solenoid valve 346 is in the ON state, that is, the first relay valve 380 is in the ON state, it is supplied to the chamber 133 of the second pressure regulating valve 102 and the second solenoid valve 346
The line hydraulic pressure Pl 2 is increased by a predetermined pressure corresponding to the output signal pressure P SOLL , while the fourth solenoid valve 346 is in the on state, that is, when the first relay valve 380 is in the off state, the chamber 136 of the second pressure regulating valve 102. The second line hydraulic pressure Pl 2
Is reduced by a predetermined pressure corresponding to the output signal pressure P SOLL . That is, in the present embodiment, the back pressure control of the accumulators 340 and 342, the second line hydraulic pressure Pl 2 that is lowered by a predetermined amount during high vehicle speed traveling or N range operation.
Line oil pressure drop control, 2nd line oil pressure Pl during sudden deceleration shift
When the second line hydraulic pressure up control for increasing 2 by a predetermined amount is executed, the fourth line hydraulic pressure Pl 4 or the second line hydraulic pressure Pl 2 is controlled by the output signal pressure P SOLL of the linear valve 500. The ON state of the linear valve 500 in FIG. 37 means the state where the output signal pressure P SOLL is being output.

Next, the operation of the electronic control unit 460 during traveling in the C mode of FIG. 37 in the present embodiment will be described with reference to the flowchart of FIG.

First, in step SS1, the input shaft rotation speed N in , the output shaft rotation speed N out , the engine rotation speed N e , and the throttle valve opening θ th are read, and then in step SS2 the actual gear ratio γ is calculated. Subsequently, in step SS3, the output torque of the engine 10, in other words, the input torque T of the CVT 14 is calculated based on the actual engine speed N e and the throttle valve opening θ th from the prestored relationship shown in FIG. After the in is calculated, in step SS4, the CVT is calculated based on the actual input torque T in and the vehicle speed V so that the engine 10 operates along the optimum curve that obtains the twisting cost and drivability.
The gear ratio control for controlling the gear ratio γ of 14 and whether or not a predetermined lockup condition is satisfied is determined based on the vehicle speed V and the like, and when the lockup condition is satisfied,
Lock-up control for selecting C mode in Fig. 37 is executed.

Then, step SS5 and subsequent steps for controlling the second line oil chamber Pl 2 are executed when the fourth solenoid valve 346 is off. First, in step SS5, the effective diameter D in of the primary side variable pulley 40, that is, the hanging diameter D in of the transmission belt 44, is calculated based on the actual gear ratio γ from the pre-stored relationship shown in FIG.
Is calculated. In the following step SS6, the actual input torque T in , the actual hanging diameter D in of the transmission belt 44 and the output shaft rotation speed N are calculated from the relationship stored in the following equation (21).
The ideal pressure P opt is calculated based on out . The following equation (2
The second term on the right side of 1) is a correction term for centrifugal hydraulic pressure, and the third term on the right side is
The term is a margin value. Further, C 1 and C 2 in the following equation (21) are constants.

P opt = C 1 · T in / D in -C 2 · N out2 + ΔP (21) However, C 1 and C 2 are constants.

Next, in step SS7, the gear ratio specific pressure P r is calculated from the following equation (22) based on the hanging diameter D in of the transmission belt 44.

Pr = C 3 · D in -C 4 (22) However, C 3 and C 4 are constants.

Then, in step SS8, it is determined whether or not the ideal pressure P opt calculated in step SS6 is equal to or higher than the gear ratio specific pressure P r calculated in step SS7.
That is, it is determined whether or not the ideal pressure P opt is located in a region above the line indicating the gear ratio specific pressure P r in FIGS. 44 to 46.

If the determination in step SS8 is affirmative, the contents of the speed ratio pressure P r 'below equation (24) in step SS9 is replaced with the speed ratio pressure P r calculated in step SS7. However, if the determination in step SS8 is negative, the content of the gear ratio specific pressure P r ′ in equation (24) described later in step SS10 is replaced with the ideal pressure P opt calculated in step SS6. .

In the following step SS11, the actual throttle valve opening θ th is calculated from the previously stored data map representing the relationship of the following equation (23).
The throttle pressure P th is calculated based on the above, and in step SS12, P mec ′ is calculated based on the gear ratio specific pressure P r ′ and the throttle pressure P th from the stored relationship shown in the following equation (24). To be done. Then, in step SS13, the above value P mec ′ is calculated from the prestored relation shown in the following equation (25).
And the ideal pressure P opt calculated in step SS6 above
After the output signal pressure P SOLL of linear valve 500 is determined based on, at step SS14, the drive current value P SOLL for obtaining its output signal pressure P SOLL a predetermined stored relationship shown in FIG. 40 is determined The driving current value I SOLL is output in step SS15.

P th = M apth ) …… (23) P mec ′ = C 5 + C 6・ P th -C 7・ P r ′ …… (24) P SOLL = C 8 (P mec ′ −P opt ) (25) However, C 5 , C 6 , C 7 , and C 8 are constants.

Here, in step SS8, it is determined whether or not the ideal pressure P opt calculated in step SS6 is located in a region above the diagonal line showing the gear ratio specific pressure P r in FIGS. 44, 45 and 46. It is for judging. If the ideal pressure P opt is located in the area above the diagonal line indicating the gear ratio specific pressure P r , (24)
The value P mec ′ shown in FIG. 44 is obtained by setting the content of P r ′ on the right side of the equation to the actual transmission specific pressure P r . This value
P mec ′ is the second when the output signal pressure P SOLL is not applied.
Value Pl 2 (= P mec ) at which pressure regulating valve 102 regulates pressure according to equation (20)
Is the same value as. For this reason, the second line hydraulic pressure Pl 2 is set to the above P
Output signal pressure P for reducing from mec ′ to ideal pressure P opt
The ideal pressure I opt is obtained by calculating SOLL , that is, the output signal pressure P SOLL for obtaining the step-down value (P mec ′ −P opt ).

On the contrary, when the ideal pressure P opt is located in the region below the diagonal line indicating the gear ratio P r , the content of P r ′ on the right side of the equation (24) is replaced with the ideal pressure P opt. As a result, the value P mec ′ shown in FIGS. 45 and 46 is obtained. This value P mec ′ is not located on a line parallel to the vertical axis corresponding to the value of the gear ratio γ at that time like the ideal pressure P opt , but is indicated by the diagonal line indicating the gear ratio specific pressure P r and the ideal value. It is located on a line parallel to the vertical axis that passes through the intersection with a line parallel to the horizontal axis indicating the pressure P opt , and on this line the output signal pressure P for reducing the value P mec ′ to the ideal pressure P opt . SOLL is required.

The output signal pressure P SOLL thus obtained is also a value for reducing the value P mec to the ideal pressure P opt on a line parallel to the vertical axis corresponding to the value of the gear ratio γ at that time. That is, in the present embodiment, by calculating the step-down value (P mec ′ −P opt ) in the region above the diagonal line showing the gear ratio specific pressure P r , in each of the three cases as in the first embodiment, The complicated calculation to calculate is eliminated. Moreover, since the calculation is simplified in this way, there is an advantage that the calculation error by the computer is reduced and the pressure adjustment accuracy of the second line hydraulic pressure Pl 2 is improved.

Further, according to the present embodiment, the output signal pressure P SOLL having a magnitude corresponding to the drive signal which is an analog signal is output from the linear valve 500, so that the second line hydraulic pressure Pl 2 and the fourth line hydraulic pressure Pl 2
Since the line oil pressure Pl 4 is controlled, not only the durability of the valve is improved, but also the fuel consumption, noise, and the shift feeling of the CVT 14 are improved, and the pulsation corresponding to the duty frequency is removed from the signal pressure. There is an advantage that the accumulator 372 is unnecessary. Also, for example, the fourth line hydraulic pressure used as back pressure for the accumulators 340 and 342.
The control of Pl 4 has an advantage that pressure regulation accuracy and responsiveness are obtained as compared with the case of duty driving.

In addition, according to the present embodiment, even if the linear solenoid 504 of the linear valve 500 is broken, the second line hydraulic pressure Pl 2 remains at the value P mec.
Since there is no slippage of the transmission belt 44, there is an advantage that the vehicle can continue running.

Further, another embodiment of the present invention will be described. In this embodiment, step SS13-1 shown in FIG. 47 and new SS13-2 are executed instead of step SS13 shown in FIG.
In FIG. 47, in step SS13-2, the output signal pressure P SOLL is calculated by using the following equation (26).
This equation (26) is obtained by multiplying the right side of the above equation (25) by the negative torque correction coefficient K 1 .

P SOLL = K 1 · C 8 (P mec ′ −P opt ) ... (26) Step SS executed before step SS13-2 above
In 13-1, the negative torque correction coefficient K 1 is set to step SS in FIG. 47.
It is calculated based on the actual input torque T in from the pre-stored relationship shown in 13-1. In this relationship, when the input torque T in is negative, the negative torque correction coefficient K 1 is “0”, and when the input torque T in is larger than the predetermined positive torque value T a , the negative torque correction coefficient K 1 is set. Is set to "1" and the input torque
When T in is smaller than the predetermined positive torque value T a and larger than “0”, the input torque T in is increased from “0” to “1”. As a result, when the input torque T in is larger than the predetermined positive torque value T a , the output signal pressure P SOLL having the same value as in the embodiment of FIG. 41 is used to regulate the second line hydraulic pressure Pl 2. However, if the input torque T in is negative,
The output signal pressure P SOLL becomes zero and the second line hydraulic pressure Pl 2 becomes the 21st.
When the input torque T in is smaller than a predetermined positive torque value T a and larger than “0”, the output signal pressure P SOLL changes from zero to zero as the input torque T in increases. To be increased.

As a result, in this embodiment, when the input torque T in is negative as during engine braking, the second line hydraulic pressure Pl 2 is increased from the ideal pressure P opt to P mec , so during negative torque traveling. The slippage of the transmission belt 44, which easily occurs, is preferably eliminated. That is, in general, the transmission belt 44
The thrust force W (T in ) for preventing the slip of the vehicle is constant regardless of whether the input torque T in is positive or negative, while the thrust force ratio (= W out / W in ) during negative torque traveling is higher than that during positive torque traveling. 1
Since the gear ratio γ cannot be maintained unless it is increased above, the thrust force W in of the primary side hydraulic cylinder 54 during negative torque traveling becomes larger than the thrust force W out of the secondary side hydraulic cylinder 56, and the thrust force W out is the above thrust force. It tends to be smaller than W (T in ). Therefore, to prevent slippage of the transmission belt 44,
As shown by the solid line in Fig. 48, in the region where the input torque T in is negative, it is necessary to increase the second line hydraulic pressure Pl 2 with respect to the input torque | T in | Since the torque | T in | itself is unstable as shown by the broken line, the second line hydraulic pressure Pl 2 (= P mec ) is eventually used to prevent slippage even in the presence of such variations.
Is set to a value as shown by the one-dot chain line. The predetermined positive torque value T a is a torque value corresponding to the minimum of the above-mentioned one-dot chain line. Then, when the input torque T in falls below a predetermined positive torque value T a at which the second line oil pressure Pl 2 starts to increase with the increase of the negative torque, the second line oil pressure Pl 2 is changed from the ideal pressure P opt to P mec. The negative torque correction coefficient K 1 is changed so as to approach.

In the embodiment shown in FIG. 49, instead of step SS6 in FIG. 41, steps SS6-1, SS6-2, SS6 shown in FIG.
-3, SS6-4 is executed. Referring to FIG. 49, in step SS6-1, the input torque T is calculated based on the actual throttle valve opening θ th from the prestored relationship shown in FIG.
The corrected torque value T c when in is zero, that is, the vertical axis intercept torque value T c is determined, and in step SS6-2, the actual engine speed N is determined from the prestored relationship shown in FIG. Based on e , the positive torque value T A corresponding to the minimum value of the corrected input torque T ins is determined. The positive torque value T A corresponding to the minimum value of the corrected input torque T ins may be a constant value, but by performing the above,
The modified input torque T ins corresponding to the 10 characteristics can be obtained.
In the following step SS6-3, based on the vertical axis intercept torque T C and the positive torque value T A , step SS6-3 in FIG. 49 is executed.
The relationship shown in is determined and the corrected input torque T ins is calculated based on the actual input torque T in . And step SS
In 6-4, the ideal pressure P opt is calculated based on the corrected input torque T ins obtained in step SS6-3 from the following equation (27) similar to step SS6 in FIG. 41.

P opt = C 1 · T ins / D in -C 2 / N out2 + ΔP (27) The ideal pressure P opt obtained in this way is such that the input torque T in approaches zero when the engine is running. As the pressure is set closer to the set pressure P mec of the second pressure regulating valve 102, the slippage of the transmission belt 44 during the negative torque traveling is preferably prevented as in the embodiment of FIG. 47.

Further, according to the present embodiment, even when traveling with negative torque,
There is an advantage that the centrifugal oil pressure can be corrected. Incidentally, in the embodiment of FIG. 47, when the negative torque correction coefficient K 1 is set to “0” during the negative torque running, the output signal pressure P SOLL calculated by the equation (26) becomes zero. However, when the engine brake travels at high speed, the centrifugal hydraulic pressure cannot be corrected although it needs to be corrected.

Further, instead of step SS13 in FIG. 41, steps SS13-11 and SS13-12 in FIG. 52 may be executed. In step SS13-11, after the coefficient K 2 is determined based on the duty ratio of the actual second solenoid valve 268 from a pre-stored relationship shown in its frame, the step S13-1
In 2, after the coefficient K 2 is determined based on the following equation (28),
In step SS13-12, the linear valve is calculated based on the following equation (28).
The output signal pressure P SOLL of 500 is determined. Step SS13 above
The relationship shown in the frame of -11 is the second solenoid valve 2 related to the opening degree of the flow rate control valve 264, in other words, the shift speed of the CVT 14.
It represents the relationship between the duty ratio of 68 and the coefficient K 2. In the state where the CVT 14 is in a relatively gentle shift state, that is, in the slow shift state, the content of the coefficient K 2 is set to "1", and the slow shift state. It is set in advance that the content of the coefficient K 2 changes toward “0” in accordance with the sudden shift state from.

P SOLL = K 2 · C 8 · (P mec ′ −P opt ) ... (28) where C 8 is a constant.

According to this embodiment, in the above step SS13-11, CV
As T14 changes from the slow shift state to the rapid shift state, the coefficient K
2 is determined to be small and the output signal pressure P SOLL (= step-down amount from P mec ) of the linear valve 500 is determined to be low according to the above equation (28), so that the basic output pressure of the second line hydraulic pressure Pl 2 is reduced. The width of the descent from P mec is reduced. For example, when the content of the coefficient K 2 becomes “0”, there is no decrease in the second line hydraulic pressure Pl 2 from the basic output pressure P mec , and the second
The line oil pressure Pl 2 is made equal to P mec and the optimum control is completely eliminated. That is, during the gear shift of the CVT 14, the thrust equilibrium state of the primary side hydraulic cylinder 54 and the secondary side hydraulic cylinder 56 is disturbed and the premise of the clamping pressure control for the transmission belt 44 is broken, so the second line hydraulic pressure Pl 2 since optimal control of seems to not be established, the output signal pressure P SOLL
By raising the second line pressure Pl 2 with a lower, margin pressure in the second line pressure Pl 2 is larger prior to the actual shifting of the CVT 14, than is slippage of the transmission belt 44 is eliminated . Here, the reason why the second line oil pressure Pl 2 is increased prior to the actual shift of the CVT 14 as described above is that the duty of the second solenoid valve 268, which is a shift speed command of the CVT 14, is set as shown in FIG. 53. When the ratio is calculated and the opening degree of the flow rate control valve 264 is changed based on the calculated ratio, the actual shift of the CVT 14 is performed.
The coefficient K 2 is determined from the duty ratio of the solenoid valve 268, the output signal pressure P SOLL of the linear valve 500 is determined, and the second line hydraulic pressure is adjusted by the pressure regulating operation of the second pressure regulating valve 102 based on the output signal pressure P SOLL. This is because Pl 2 can be changed. As a result, as shown in FIG. 54, even if the input torque T in suddenly increases, the second line hydraulic pressure Pl 2 increases in a state of exceeding the optimum pressure corresponding to the input torque T in . There is an advantage that slippage is suitably eliminated. Incidentally, in the above-described embodiment, as shown in FIG. 55, the input torque T in and the target control pressure P opt are based on the actual change of the engine rotation speed N e caused by the actual shift of the CVT 14. together but sequentially calculated, the since the target control pressure the second line pressure Pl 2 by the second pressure regulating valve 102 so as to match the P opt is changed, the second line pressure Pl 2 in response to the input torque change change There is an unavoidable occurrence of the delay time D until it is caused, and there is a risk that the transmission belt 44 slips due to a change in the input torque due to the inertia torque. Therefore, as shown in FIG. 56, a relatively large margin pressure must be provided for the second line hydraulic pressure Pl 2, and the power loss for generating the hydraulic pressure could not be sufficiently reduced.

Further, steps SS6-11, SS6-12, SS6-13, SS6-14 in FIG. 57 may be executed instead of step SS6 in FIG. 41. In the figure, in step SS6-11, the actual throttle valve opening θ
A torque T c including a margin for inertial torque is calculated based on th , and in step SS6-12, a coefficient K is calculated based on the actual duty ratio of the second solenoid valve 268 from the relationship stored in advance within the frame. After 2 is determined, in step SS6-13, the correction torque T ins for calculating the ideal pressure P opt is calculated based on the following equation (29). Relationship shown in the frame in step SS6-12 as in step SS13-11 of Examples, to represent the relationship between the duty ratio and the coefficient K 2 of the second solenoid valve 268 associated with the shift speed of the CVT14 , and the state where CVT14 is relatively gradual shifting state, i.e. a "1" the contents of the coefficient K 2 is a gentle shifting state, the contents of the coefficient K 2 with the an abrupt shifting state from the slow speed state It is predetermined so as to change toward “0”.

T ins = T c -K 2 · (T c -T in ) (29) where T c is the second line hydraulic pressure suitable for a sudden gear shift associated with a sudden change in the input torque T in. It is a temporary input torque value to generate Pl 2 , and the right side of the above equation (29) is
When the coefficient K 2 is “1” (during a slow shift), the original input torque T in is obtained , and when the coefficient K 2 is “0” (during a sudden shift), the temporary input torque value T is c . In step SS6-14, the ideal pressure P opt is calculated according to the following equation (30).

P opt = C 1 · T ins / D in -C 2 · N out2 + ΔP (30) In the present embodiment, the ideal pressure P opt is set according to the duty ratio of the second solenoid valve 268, which is the shift command of the CVT 14. Since the correction is made, the second line oil pressure Pl 2 is increased prior to the actual change of the input torque, and the same effect as the embodiment of FIG. 52 is obtained. Further, according to the present embodiment, in the right side of the above equation (30), only the first term is corrected corresponding to the change of the coefficient K 2 , and the second term for correcting the centrifugal hydraulic pressure is Since it does not change in relation to the coefficient K 2 , there is an advantage that the centrifugal hydraulic pressure is suitably corrected even in a sudden speed change state during high speed traveling.

Further, steps SS6-11, SS6-12, SS6-1 in FIG. 57 above.
Instead of 3, steps SS6-21 and SS6-22 in FIG. 58 may be executed. In the figure, step SS6-2
In the case of 1, the increased torque value T is based on the actual duty ratio of the second solenoid valve 268 from the relationship stored in advance in the frame.
After ina is determined, in step SS6-22, the following equation (31)
Correction torque T ins for calculating ideal pressure P opt based on
Is calculated. The relationship shown in the frame of step SS6-21 represents the relationship between the duty ratio of the second solenoid valve 268 and the increased torque value T ina related to the shift speed of the CVT 14.
In the state where the CVT 14 is in a relatively gentle shift state, that is, in the slow shift state, the content of the increased torque value T ina is set to “0”,
It is predetermined that the content of the increased torque value T ina increases toward a predetermined maximum value as the slow shift state shifts to the rapid shift state. The maximum value of this increasing torque value T ina is abbreviated as P mec when the ideal pressure P opt when T ins obtained by adding this maximum value in the following formula (31) is substituted into the formula (30). Determined to be equal. Next, the increased torque value T ina obtained as described above is substituted into the following equation (31) to calculate the correction torque T ins .

T ins = T in + T ina (31) Then, in step SS6-14, the ideal pressure P opt is calculated according to the equation (30).

In the present embodiment, the ideal pressure P opt is corrected according to the duty ratio of the second solenoid valve 268, which is the shift command of the CVT 14, and is corrected according to the change of the coefficient K 2 on the right side of the above equation (30). Since only the first term is applied, the same effect as the embodiment shown in FIG. 57 can be obtained.

Also, the above-mentioned steps SS13-11 and SS13-12 in FIG.
Alternatively, steps SS13-21 and SS13-22 of FIG. 59 may be executed. In step SS13-21, after the coefficient K 3 are determined based on the actual deceleration G from a pre-stored relationship shown in its frame, in step SS13-22, based on the following equation (32) Linear The output signal pressure P SOLL of the valve 500 is determined. The relationship shown in the frame of step SS13-21 above is
The relationship between the actual deceleration G of the vehicle and the coefficient K 3 is expressed. In a region where the deceleration G of the vehicle is relatively moderate, the content of the coefficient K 3 is set to “1” and the deceleration G is increased. It is predetermined that the content of the coefficient K 3 changes toward “0” as

P SOLL = K 3 · C 8 (P mec ′ −P opt ) (32) According to this embodiment, as the deceleration G of the vehicle increases, the content of the coefficient K 3 is determined to be a smaller value. At the same time, the output signal pressure P SOLL of the linear valve 500 (= the amount of pressure reduction from P mec ) is determined to be a low value, so that the fall width of the second line hydraulic pressure Pl 2 from the basic output pressure P mec is reduced. As a result, in the present embodiment, even when the input torque T in to the CVT 14 suddenly increases due to the inertia torque when the vehicle suddenly stops or the wheels are locked, the second line hydraulic pressure Pl 2 is increased and the transmission belt 44 has a high pressure. Sliding is preferably prevented. The steps SS13-21 and SS13-22 of this embodiment may be executed in addition to the steps SS13-11 and SS13-12 of FIG. In this case, the linear valve 50 according to the following equation (33)
The output signal pressure P SOLL of 0 is determined.

P SOLL = P 2 · K 3 · C 8 (P mec ′ −P opt ) (33) Further, instead of steps SS6-12 and SS6-13 in FIG. 57, step SS6− in FIG. 32, SS6-33 may be executed. In the figure, in step SS6-32,
A constant based on the actual deceleration G from the relationship stored in advance.
After K 3 is determined, in step SS6-33, the ideal pressure P opt
The correction torque T ins is calculated based on the following equation (34) to calculate
Is calculated. The relationship used in step SS6-32 is the same as the relationship shown in step SS13-21 of FIG.

T ins = T c -K 3 · (T c -T in ) ... (34) In the present embodiment, the correction torque T ins is determined according to the magnitude of the deceleration G of the vehicle, and step SS6-14 is performed. Since the calculated ideal pressure P opt is increased, the same effect as the embodiment of FIG. 59 can be obtained. In addition, step SS6 of the present embodiment
-32 and SS6-33 also follow steps SS6-12 and SS6-13 in Fig. 57.
May be executed in addition to. In this case, the following equation (35)
The correction torque T ins is calculated in accordance with.

T ins = T c -K 2 · K 3 · (T c -T in ) (35) Further, instead of steps SS6-21 and SS6-22 in FIG. 58, step SS6− in FIG. 41, SS6-42 may be executed. In the figure, in step SS6-41,
After the increasing torque value T inb is determined based on the actual deceleration G from the relationship stored in advance, in step SS6-42,
To calculate the ideal pressure P opt , the correction torque T ins is calculated based on the following equation (36). It is shown in the frame of the above step SS6-41. Relationship between the torque increase value T inb be applied to T in order to calculate the deceleration G of the vehicle corrected torque T ins, the correction torque T ins obtained by adding the maximum value of the torque increase value T inb Ideal pressure when substituting for (30)
It is determined that P opt is P mec .

T ins = T in + T inb (36) In this embodiment, since the ideal pressure P opt is corrected according to the deceleration G of the vehicle, the same effect as the embodiment of FIGS. 59 and 60 is obtained. Is obtained. In addition, steps SS6-41 and SS6 of the present embodiment.
-42 may also be executed in addition to steps SS6-21 and SS6-22 in FIG. In this case, the correction torque T ins is calculated according to the following equation (37).

T ins = T in + T ina + T inb ...... (37) Further, instead of the electromagnetic valve driving signal determination routine of Figure 28,
The routine shown in FIG. 62 may be used. In step S7-10 in FIG. 62, it is determined whether or not the brake operation has been performed. When it is determined that the brake is not operated, in step S7-11, the normal solenoid valve drive signal I SOL4 is determined according to the routine composed of steps S7-1 to S7-6 in FIG. After that, the solenoid valve drive signal I SOL4 is output in step S7-13. However, if it is determined in step S7-10 that the brake is being operated, the step
In S7-12, the content of solenoid valve drive signal I SOL4 is "0".
Is set to. As described above, according to the present embodiment, when the vehicle brake operation is performed, the solenoid valve drive signal I
By setting the content of SOL4 to "0", P SOL4 =
Since the second line oil pressure Pl 2 is increased to 0 and becomes equal to P mec , the second line oil pressure Pl 2 is increased and the slippage of the transmission belt 44 is prevented when the vehicle is suddenly controlled or the wheels are locked. It is.

The embodiment of the present invention has been described above with reference to the drawings.
The present invention also applies to other aspects.

For example, in step S5-1 of the above-described embodiment, the actual gear ratio γ of the CVT 14 was calculated based on the input shaft rotation speed N in and the output shaft rotation speed N out. A sensor for detecting the directional position may be provided and may be determined by a signal from this sensor.

Further, in the above-described embodiment, the throttle valve opening detection valve 180
Although the throttle pressure P th generated by the above is used, in a vehicle such as a vehicle equipped with a diesel engine that does not use a throttle valve, a hydraulic pressure corresponding to the accelerator pedal operation amount may be used. In such a case, for example, the cam 184 of the above-described embodiment may be mechanically associated with the accelerator pedal so as to rotate with the depression of the accelerator pedal.

Further, in the above-described embodiment, the throttle valve opening sensor 468 is used to control the throttle valve opening θ th corresponding to the required output value of the vehicle.
However, a sensor for detecting other amounts such as the accelerator pedal operation amount and the fuel injection amount is provided, and the throttle pedal opening θ
It may be used instead of th .

Further, in the above-described embodiment, the fourth solenoid valve 346 and the throttle 344
Was used as the signal pressure generating means for generating the signal pressure P SOL4 , but other means such as a linear solenoid valve for generating the signal pressure by operating in response to a command from the electronic control unit 460 may be used. .

In the above-described embodiment, the step of correcting the centrifugal oil pressure
Although S3 was provided, it may not be necessarily provided for vehicles that do not travel at high speeds.

Further, in step S4 of the above-mentioned embodiment, the optimum control pressure P
A predetermined margin pressure value may be added when opt is obtained.

In the shift control of the CVT 14 in the above-described embodiment, the target rotation speed N in * is controlled so that the actual input shaft rotation speed N in matches, but the gear ratio γ = N in / N out . Therefore, even if the gear ratio γ is controlled so that the actual gear ratio γ coincides with the target gear ratio γ * , it is substantially the same.

The above description is merely one embodiment of the present invention, and the present invention can be variously modified without departing from the spirit thereof.

[Brief description of drawings]

FIG. 1 is a detailed circuit diagram of a hydraulic control device for operating the device of FIG. FIG. 2 is a skeleton view of a vehicle power transmission device provided with a hydraulic control device according to an embodiment of the present invention. FIG. 3 is a diagram showing the second pressure regulating valve of FIG. 1 in detail. FIG. 4 is a detailed view of the first pressure regulating valve of FIG. FIG. 5 is a diagram showing the output characteristics of the throttle valve opening detection valve of FIG. FIG. 6 is a diagram showing the output characteristic of the gear ratio detection valve of FIG. FIG. 7 is a diagram showing the output characteristic of the second pressure regulating valve of FIG. FIG. 8 is a diagram showing ideal characteristics of the second line hydraulic pressure. FIG. 9 is a diagram for explaining the relationship between the operating states of the first solenoid valve and the second solenoid valve in the shift control valve device of FIG. 1 and the CVT shift state of FIG. FIG. 10, FIG. 11 and FIG. 12 are views for explaining the relationship between the gear ratio of the CVT of FIG. 2 and the hydraulic pressure value of each part.
FIG. 10 is a diagram showing a positive torque traveling state, FIG. 11 is a diagram showing an engine braking traveling state, and FIG. 12 is a diagram showing an unloaded traveling state. FIG. 13 is a diagram showing an output characteristic of the first pressure regulating valve of FIG. 4 with respect to the primary-side hydraulic cylinder hydraulic pressure or the second line hydraulic pressure. FIG. 14 is a diagram showing a change characteristic between the duty ratio of the fourth solenoid valve and the hydraulic pressure which is continuously changed in relation to the duty ratio in the hydraulic circuit of FIG. FIG. 15 is a diagram showing a change characteristic between the duty ratio of the fourth solenoid valve and the fourth line hydraulic pressure continuously changed in association with the duty ratio in the hydraulic circuit of FIG. FIG. 16 is a diagram for explaining the second line hydraulic pressure that changes in relation to the vehicle speed (centrifugal hydraulic pressure). FIG. 17 shows control modes (A), (B),
Third solenoid valve and fourth solenoid valve in (C), (D) and (E)
It is a figure which shows the operating state of a solenoid valve. FIG. 18 is a flow chart for explaining an essential part of the operation of the electronic control unit of FIG. FIG. 19 is a diagram showing the output torque characteristics of the engine stored in advance in the electronic control unit of FIG. FIG. 20 is a diagram showing the basic output pressure of the second pressure regulating valve of FIG. 1 and the optimum control pressure lowered by the centrifugal hydraulic pressure in relation to the vehicle speed when the gear ratio is the minimum value γ min . FIG. 21 is a diagram showing the basic output pressure and the optimum control pressure of the second pressure regulating valve in relation to the gear ratio and the throttle valve opening. FIG. 22 is a view showing the basic output pressure estimation routine of FIG. FIG. 23 is a diagram showing the correlation between the logarithmic value of the gear ratio and the value using an approximate expression. Figures 24 and 25 show the relationship map 1
th) and map 2 a (theta th) is a diagram showing respectively. 26th
The figure is a diagram showing the relationship between the drive signal I SOL4 of the fourth solenoid valve of FIG. 1 and the reduced hydraulic pressure value P down of the second line hydraulic pressure obtained by the signal pressure generated by the second solenoid valve.
FIG. 27 is a diagram showing a basic output pressure characteristic when the gear ratio axis is set to a uniform scale. FIG. 28 is a diagram showing a second solenoid valve drive signal determination routine of the embodiment of FIG. FIG. 29 is a diagram showing the relationship used in FIG. 28. FIG. 30 is a diagram showing a routine for obtaining a basic output pressure during reverse drive traveling in another embodiment of the present invention. FIG. 31 is a diagram showing the relationship used in FIG. FIG. 32 is a view showing a main part of a hydraulic circuit according to another embodiment of the present invention,
FIG. 33 is a diagram showing an optimum control pressure correction routine for a vehicle provided with the hydraulic circuit. FIG. 34 is a time chart explaining the operation obtained by FIG. 33. FIG. 35 is a view corresponding to FIG. 33 in another embodiment of the present invention. 36th
The figure shows a hydraulic control circuit according to another embodiment of the present invention. FIG. 37 is a chart showing the control modes of the embodiment of FIG. FIG. 38 is a diagram for explaining in detail the configuration of the second pressure regulating valve of the embodiment of FIG. 36. FIG. 39 is an enlarged view of the linear valve of FIG. 36. FIG. 40 is a diagram showing the output characteristic of the linear valve of FIG. 39. FIG. 41 is a flow chart for explaining the control operation of the embodiment shown in FIG. First
42 and 43 are diagrams showing the relationships used in the flowchart of FIG. 41, respectively. 44, 45
FIG. 46 and FIG. 46 are views for explaining the operation of FIG. 41, respectively, and FIG. 44 shows a state in which the ideal pressure is in the region above the slope line showing the gear ratio, FIG. 45 and FIG. The figure shows a state in which the ideal pressure is in a region below the slope line indicating the gear ratio specific pressure. FIG. 47 is a diagram showing another example of a part of the flowchart of FIG. 41. FIG. 48 is a view for explaining the operation of the embodiment shown in FIG. 47. FIG. 49 is a diagram showing another example of a part of the flowchart of FIG. 41. Figure 50 and Figure
FIG. 51 is a diagram showing the relationships used in the embodiment of FIG. 49. 52, 57, and 58,
FIG. 42 is a diagram showing another example of a part of the flowchart of FIG. 41. FIG. 53 is a time chart for explaining the operation sequence of the embodiment in FIG. 52, and FIG. 54 is a time chart showing changes in the input torque and the second line hydraulic pressure in the embodiment in FIG. 55 and 56 are views corresponding to FIGS. 53 and 54 in the embodiment of FIG. 41. First
FIG. 59, FIG. 60, FIG. 61 and FIG. 62 are flowcharts for explaining the main part of another embodiment of the present invention. 14: CVT (belt type continuously variable transmission) 40, 42: Variable pulley 44: Transmission belt 54: Primary hydraulic cylinder (hydraulic actuator) 56: Secondary hydraulic cylinder (hydraulic actuator) 102: Second pressure regulating valve 344: Throttle} (Signal pressure generating means) 346: Fourth solenoid valve} (Signal pressure generating means) 460: Electronic control unit 468: Throttle valve opening sensor (request output value sensor) 500: Linear valve (signal pressure generating means) Step S5-1, SS2: Gear ratio determination means Pl 2 : Second line hydraulic pressure (tension control pressure) P SOL4 , P SOLL : Signal pressure

 ─────────────────────────────────────────────────── ─── Continuation of the front page (72) Inventor Nobuyuki Kato 1 Toyota Town, Toyota City, Aichi Prefecture Toyota Automobile Co., Ltd. (56) References JP 63-38041 (JP, A) JP 63-34247 (JP, A) JP-A-59-19754 (JP, A) JP-A-64-49758 (JP, A)

Claims (1)

[Claims]
1. A pair of variable pulleys respectively provided on a primary side rotating shaft and a secondary side rotating shaft, a transmission belt wound between the pair of variable pulleys, and an effective diameter of the pair of variable pulleys. A hydraulic control device for a vehicle belt-type continuously variable transmission, comprising: a pair of hydraulic actuators to be changed, wherein a gear ratio specific pressure representing a gear ratio of the belt type continuously variable transmission and a required output representing a required output of the vehicle. A valve element that operates based on the pressure and the signal pressure is provided. According to the operation of the valve element, an optimal control pressure of the tension of the transmission belt is generated in the state where the signal pressure is generated, and the signal pressure is generated. In a state in which the tension control pressure is not set, the tension control pressure larger than the optimum control pressure is increased as the gear ratio is increased, and is increased as the required output of the vehicle is increased. Generated signal pressure generating means, gear ratio determining means for determining the gear ratio of the belt type continuously variable transmission, required output value sensor for detecting a required output value of the vehicle, and from the relationship stored in advance, the actual An electronic control unit that controls the signal pressure generating means so that the optimum control pressure is generated from the pressure regulating valve based on the gear ratio and the required output value. Machine hydraulic control device.
JP29128890A 1990-02-01 1990-10-29 Hydraulic control device for vehicle belt type continuously variable transmission Expired - Fee Related JPH07117143B2 (en)

Priority Applications (6)

Application Number Priority Date Filing Date Title
JP2-23579 1990-01-02
JP2357990 1990-02-01
JP13448990 1990-05-24
JP2-134489 1990-05-24
JP24176490 1990-09-12
JP2-241764 1990-09-12

Applications Claiming Priority (3)

Application Number Priority Date Filing Date Title
EP91300676A EP0440422B1 (en) 1990-02-01 1991-01-28 Hydraulic control apparatus for belt-and-pulley type continuously variable transmission, incorporating means for optimizing belt tensioning pressure
DE1991608754 DE69108754T2 (en) 1990-02-01 1991-01-28 Hydraulic control system of a continuously variable belt transmission with belt pressure optimization.
US07/647,424 US5157992A (en) 1990-01-02 1991-01-29 Hydraulic control apparatus for belt-and-pulley type continuously variable transmission, incorporating means for optimizing belt tensioning pressure

Publications (2)

Publication Number Publication Date
JPH04191555A JPH04191555A (en) 1992-07-09
JPH07117143B2 true JPH07117143B2 (en) 1995-12-18

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JP29128890A Expired - Fee Related JPH07117143B2 (en) 1990-02-01 1990-10-29 Hydraulic control device for vehicle belt type continuously variable transmission

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Country Link
JP (1) JPH07117143B2 (en)

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* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JP4649907B2 (en) * 2004-08-05 2011-03-16 株式会社豊田中央研究所 Control device for continuously variable transmission
JP4847567B2 (en) * 2009-08-26 2011-12-28 ジヤトコ株式会社 Continuously variable transmission and control method thereof

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