JPH04277363A - Control device of belt type continuously variable transmission for vehicle - Google Patents

Abstract

PURPOSE:To provide the control device of a belt type continuously variable transmission for a vehicle in which slip of a transmitting belt is not generated even at suddenly regaining grip with the road surface after slip of driving wheels. CONSTITUTION:When variation from the grip condition of driving wheels 24 to the slip condition is detected by a driving wheel frictional condition variation detecting means 320 corresponding to an electronic control device 110, belt pinching force for a transmitting belt is temporarily increased for a prescribed time, by a belt pinching force increasing means 322 corresponding to a solenoid valve 108 and a changeover valve 310. Hereby, even when grip with the road surface is suddenly regained on the driving wheels 24 by returning operation of an accelerator pedal after slip of the driving wheels 24, rotating inertia force generated by sudden drop of rotation hitherto is added to output torque of an engine 10, and transmitting torque of the transmitting belt 70 is suddenly increased, slip of the transmitting belt 70 is appropriately prevented.

Description

[Detailed description of the invention]

0001

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

0002

[Prior Art] A belt type vehicle is equipped with a pair of variable pulleys around which a transmission belt is wound, and by changing the effective diameter of the variable pulleys, the speed of the engine is continuously variable and transmitted to the drive wheels. 2. Description of the Related Art In continuously variable transmissions, there is known a control device for a belt-type continuously variable transmission for a vehicle that controls the squeezing force of a power transmission belt according to a throttle valve opening corresponding to an input torque and a gear ratio. For example, the one described in Japanese Patent Application Laid-Open No. 64-49753 is one such example.

0003

[Problems to be Solved by the Invention] However, in the conventional control device for a belt-type continuously variable transmission for vehicles as described above, the clamping force is adjusted so that the clamping force is small within a range where no slipping of the transmission belt occurs. The line oil pressure that achieves this is optimally controlled according to the throttle valve opening and gear ratio, reducing power loss and increasing the durability of the transmission belt. However, under the conditions where the clamping force of the transmission belt is controlled as described above, for example, when driving on a low-friction road such as a snowy or frozen road or when driving on a steep slope, the drive wheel slips and the grip with the road surface is reduced. If it was suddenly recovered, there was a risk that the transmission belt would slip. In other words, when the drive wheels slip, the rotational speed of the engine, transmission joints such as clutches and torque converters, primary and secondary pulleys, axles, etc. increases rapidly.
After that, the driver returns the accelerator pedal in an attempt to grip the driving wheels, so the driving wheels become gripped. When the drive wheels become gripped in this way, the previous rotation rapidly decreases, and the rotational inertia generated by this sudden decrease in rotation is added to the engine's output torque, causing a large torque to be applied to the transmission belt. will be added transiently.

The present invention was made against the background of the above-mentioned circumstances, and its object is to prevent the transmission belt from slipping even when the drive wheels suddenly regain grip with the road surface after slipping. An object of the present invention is to provide a control device for a belt-type continuously variable transmission for a vehicle.

[0005]

[Means for Solving the Problems] The gist of the present invention to achieve the above object is to provide a pair of variable pulleys around which a transmission belt is wound, and to change the effective diameter of the variable pulleys. A vehicle belt that controls the clamping force of the transmission belt according to the input torque and gear ratio in a vehicle belt-type continuously variable transmission that changes engine rotation steplessly and transmits the engine rotation to the drive wheels. A control device for a continuously variable transmission, comprising: (a) drive wheel friction state change detection means for detecting a change in the drive wheel from a grip state to a slip state; and (b) drive wheel friction state change detection means. The belt clamping force increasing means temporarily increases the belt clamping force for a predetermined period when it is detected that the drive wheel changes from a grip state to a slip state.

[0006]

[Operation] With this structure, when the drive wheel friction state change detection means detects a change in the driving wheel from the grip state to the slip state, the belt clamping force increasing means increases the belt clamping force of the transmission belt for a predetermined period of time. temporarily increased.

[0007]

[Effect of the invention] Therefore, when the drive wheel suddenly regains grip with the road surface by returning the accelerator pedal after the drive wheel has slipped, the rotation that occurs due to the sudden drop in the previous rotation. Even if inertia force is added to the output torque of the engine and the transmission torque of the transmission belt increases rapidly, the belt clamping force of the transmission belt is temporarily increased for a predetermined period by the belt clamping force increasing means, so that the transmission belt Slippage is suitably prevented.

[0008]

DESCRIPTION OF THE PREFERRED EMBODIMENTS An embodiment of the present invention will be described below in detail with reference to the drawings. FIG. 1 is a schematic diagram showing a vehicle power transmission device including a belt type continuously variable transmission to which an embodiment of the present invention is applied. In the figure, the power of an engine 10 is transmitted through a torque converter 12 with a lock-up clutch, a forward/reverse switching device 14, and a belt-type continuously variable transmission (hereinafter referred to as CVT) 1.
6, a reduction gear device 18, and a differential gear device 20 to be transmitted to drive wheels 24 connected to an axle 22.

The torque converter 12 is connected to the engine 1.
The converter output shaft 32 is fixed to a pump impeller 28 connected to the crankshaft 26 of No. 0, and a converter output shaft 32 provided concentrically between the crankshaft 26 and the central shaft 54 of the rear-stage forward/reverse switching device 14. A turbine impeller 34 receives oil from the pump impeller 28 and is rotated; a stator impeller 38 is fixed to a non-rotating member via a one-way clutch 36; and a converter output shaft 32 via a damper 40.
A lock-up clutch 42 is fixed to the lock-up clutch 42, and when the lock-up clutch 42 is not engaged, torque is transmitted at an amplification factor corresponding to the input/output rotational speed ratio. The lock-up clutch 42 is configured such that, for example, the vehicle speed
When the engine rotational speed or the rotational speed of the turbine impeller 34 exceeds a predetermined value, the crankshaft 26 is activated.
This directly connects the converter output shaft 32 and the converter output shaft 32.

The forward/reverse switching device 14 is a double pinion type planetary gear device that is selectively switched to a forward gear or a reverse gear depending on the operating position of a shift lever 126, which will be described later. It is provided concentrically with the input shaft 44 of. This planetary gear device includes a sun gear 4 fixed to a converter output shaft 32 that functions as an input shaft of a forward/reverse switching device 14.
6, a ring gear 48 provided concentrically with the sun gear 46, a pair of planet gears 50 and 52 that mesh with one and the other of the ring gear 48 and the sun gear 46, and mesh with each other, and a ring gear 48 provided concentrically with the sun gear 46 and the ring gear 48. a central shaft 54, a flange portion 56 extending from the central shaft 54 toward the outer circumference, and a flange portion 56;
and a carrier pin 58 that is erected in a direction parallel to the axis of the central shaft 54 and rotatably supports the pair of planetary gears 50 and 52. Further, this planetary gear device includes a forward clutch 62 for coupling between the converter output shaft 32 and the carrier 60.
and a reverse brake 64 for connecting the ring gear 48 and the fixed housing 63. Therefore, when forward clutch 62 is engaged, converter output shaft 32 and carrier 60 are connected, and
Since converter output shaft 32 and center shaft 54 rotate integrally, CVT 16 and below are rotated in the forward direction. On the contrary, the reverse brake 6 is used instead of the forward clutch 62.
4 is engaged, the housing 63 and the ring gear 4 are engaged.
Since the ring gear 48 is in a non-rotating state, the central shaft 54 is rotated in the opposite direction to the converter output shaft 32, and the CVT 16 and below are rotated in the reverse direction.

The CVT 16 has variable pulleys 66 and 68 provided on its input shaft 44 and output shaft 45, respectively.
and a transmission belt 70 wound around the variable pulleys 66 and 68. This transmission belt 70 is disclosed in, for example, Japanese Patent Application Laid-open No. 61-116146,
As described in Japanese Patent No. 62445, the belt block is composed of a large number of belt blocks which are pressed by the inner wall surfaces of the V grooves of the variable pulleys 66 and 68 and are arranged in series along an endless annular hoop or chain. There is. The variable pulleys 66 and 68 are connected to fixed rotating bodies 72 and 74 fixed to the input shaft 44 and output shaft 45, respectively, and to the input shaft 4.
Hydraulic cylinders 80 and 82 consist of movable rotary bodies 76 and 78 provided on the output shaft 4 and the output shaft 45, respectively, so as to be movable in the axial direction but not relatively rotatable around the axis, and the movable rotary bodies 76 and 78 function as hydraulic actuators.
By moving the CVT1, the V groove width, that is, the hanging diameter (effective diameter) of the transmission belt 70 is changed.
6 (=rotational speed Nin of input shaft 44/rotational speed Nout of output shaft 45) is changed. The hydraulic cylinder 80 is operated exclusively to change the gear ratio γ, and the hydraulic cylinder 82 is operated exclusively to change the transmission belt 7.
It is operated so that the minimum clamping force is obtained within a range where no slippage occurs. The hydraulic pump 84 constitutes a hydraulic power source of a CVT hydraulic control device (not shown), and is constantly driven to rotate by a pump impeller 28 that rotates together with the engine 10.

[0012] The output shaft 45 of the CVT 16 is provided with a first gear 86 that functions as an output gear. Further, a second gear 90 that meshes with the first gear 86 and a third gear 92 having a smaller diameter are fixed to a rotating shaft 88 that is rotatably provided around an axis parallel to the axis of the first gear 86. The third gear 92 is meshed with the large diameter gear 94 of the differential gear 20 . The first gear 86 and the second gear 90
, and the third gear 92 constitute the reduction gear device 18. The differential gear device 20 includes a pair of differential small gears 96 that are rotatably supported around an axis perpendicular to the rotational axis of the axle 22 and rotate integrally with the large diameter gear 94, and the differential small gears 96. and a pair of differential large gears 98 that mesh with the axle shaft 22 and are connected to the axle shaft 22. Therefore, after the power transmitted from the reduction gear device 18 is equally distributed to the left and right axles 22 in the differential gear device 20,
The signal is transmitted to the left and right drive wheels 24.

[0013] The electronic control unit 110 for CVT includes a CPU
112, a ROM 114, a RAM 116, an interface (not shown), and the like. an input shaft rotation sensor 120 and an output shaft rotation sensor 122 that detect the respective values, a throttle valve opening sensor 124 that detects the throttle valve opening provided in the intake pipe of the engine 10, and a shift operation lever 1.
A signal representing the vehicle speed SPD, a signal representing the input shaft rotation speed Nin, and an output shaft rotation are output from the operation position sensor 128 for detecting any of the 26 operation positions, that is, L, S, D, N, R, and P ranges. A signal representing the speed Nout, a signal representing the throttle valve opening θth, and a shift operation lever 126
A signal representing the operating position Ps of each is supplied. CPU 112 of the electronic control device 110
The RO is stored in advance using the temporary storage function of the RAM 116.
The input signal is processed according to the program stored in M114, and the first solenoid valve 102 and the second solenoid valve in the hydraulic control circuit 100 are
The solenoid valve 104, the linear solenoid valve 106, and the third solenoid valve 108 are controlled respectively.

FIG. 2 shows the main parts of the hydraulic control circuit 100, and the hydraulic pump 84 is connected to the hydraulic control circuit 10.
0 hydraulic pressure source. The hydraulic pump 84 sucks the hydraulic oil that has returned to an oil tank (not shown) through the strainer 140, and also sucks the hydraulic oil that has been returned through the return oil passage 142 and pumps it to the first line oil passage 144. In this embodiment, the hydraulic oil in the first line oil passage 144 is leaked to the return oil passage 142 and the lock-up clutch pressure oil passage 148 by the overflow (relief) type first pressure regulating valve 146, so that the hydraulic oil in the first line oil passage 144 is The first line oil pressure Pl1 in the oil passage 144 is regulated. In addition, as the first line oil pressure Pl1 is reduced by the second pressure regulating valve 150 of the pressure reducing valve type, the second line oil passage 15
The second line oil pressure Pl2 in the second line is regulated. Note that the first line oil passage 144 is provided with a relief valve 154 for preventing an abnormal pressure increase above a predetermined pressure.

The second pressure regulating valve 150 includes a spool valve 160 that opens and closes between the first line oil passage 144 and the second line oil passage 152, a spring seat 162, a return spring 164, and a plunger 166. Also,
At the shaft end of the spool valve element 160, there are a first land 168, a second land 170, and a third land 172 whose diameters increase in order.
are formed sequentially. An oil chamber 176 is provided between the second land 170 and the third land 172, into which the second line hydraulic pressure Pl2 is introduced through the throttle 174 as feedback pressure, and the spool valve element 160 is closed by the second line hydraulic pressure Pl2. The valve is biased toward the valve. Further, an oil chamber 180 is provided on the end surface side of the first land 168 of the spool valve element 160, through which a gear ratio pressure Pr (described later) is guided through a throttle 178, and the spool valve element 160 is closed by the gear ratio pressure Pr. The valve is biased toward the valve. In the second pressure regulating valve 150, the biasing force of the return spring 164 in the valve opening direction is applied to the spring seat 16.
2 to the spool valve element 160. Further, an oil chamber 182 is provided between the first land 168 and the second land 170 of the spool valve 160, into which the signal pressure PsolL output from the linear solenoid valve 106 is introduced. Spool bento 1
60 is biased in the valve closing direction. Furthermore, the end surface side of the plunger 166 is provided with a throttle pressure P, which will be described later.
An oil chamber 184 for applying th is provided,
The spool valve element 160 is biased in the valve opening direction by this throttle pressure Pth. Therefore, the pressure receiving area (cross-sectional area) of the first land 168 is A1, the cross-sectional area of the second land 170 is A2, the cross-sectional area of the third land 172 is A3, the pressure receiving area of the small diameter land 188 of the plunger 166 is A4, and the return spring When the urging force of 164 is W, the spool valve 160 is balanced at a position where Equation 1 is satisfied. That is, spool valve 1
60 is moved according to formula 1, so that
The hydraulic oil in the first line oil passage 144 is transferred to the second line oil passage 15.
2 and the state in which the hydraulic oil in the second line oil passage 152 flows out to the drain port are repeated, and the second line oil pressure Pl2 is regulated. Since the second line oil passage 152 is a relatively closed system, the second pressure regulating valve 150 reduces the first line oil pressure Pl1, which is a relatively high oil pressure, as described above.
Line oil pressure Pl2 is generated as shown in FIG. Note that FIG. 3 shows the characteristics when the throttle pressure Pr is constant, and the line shows the pressure regulation value (basic output pressure Pmec) when the signal pressure PsolL is zero, and the curve shows the ideal pressure depending on the signal pressure PsolL. It shows the pressure regulation value when the pressure is regulated to Popt. That is, the signal pressure Ps
olL is the basic output pressure Pm to obtain the ideal pressure Popt.
It is supplied to control the drop value from ec. Furthermore, an oil chamber 190 is provided between the large-diameter land 186 and the small-diameter land 188 of the plunger 166 to which a signal pressure PsolL for temporarily increasing the second line oil pressure Pl2 is supplied.

[0016]

[Math 1]

The first pressure regulating valve 146 has a spool valve 200
, spring seat 202, return spring 204
, the first plunger 206, and the first plunger 2
The second plunger 208 has the same diameter as the second land 215 of 06
It is equipped with The spool valve 200 opens and closes between the first line oil passage 144, the return oil passage 142, and the lock-up clutch pressure oil passage 148, and has a first line as feedback pressure on the end surface of the first land 212. An oil chamber 213 is provided for applying hydraulic pressure Pl1 through a throttle 211, and the spool valve element 200 is biased in the valve opening direction by this first line hydraulic pressure Pl1. An oil chamber 216 for guiding the throttle pressure Pth is provided between the first land 214 of the first plunger 206, which is provided coaxially with the spool valve element 200, and the second land 215, which has a smaller diameter. Further, an oil chamber 217 is provided between the second land 215 and the second plunger 208 for guiding the hydraulic pressure Pin in the primary side hydraulic cylinder 80 via a branch oil passage 220, and a second
An oil chamber 218 for guiding the second line oil pressure Pl2 is provided on the end face of the plunger 208. Since the biasing force of the return spring 204 is applied to the spool valve element 200 in the valve closing direction via the spring seat 202, the pressure receiving area of the first land 212 of the spool valve element 200 is A5, and the pressure receiving area of the first plunger 206 is 1st land 214
The cross-sectional area of the second land 215 and the second plunger 208 is A7, and the return spring 20 has a cross-sectional area of A6.
If the biasing force of 4 is W, the spool valve 200 is calculated by formula 2.
The first line oil pressure Pl1 is regulated at a position where the line oil pressure Pl1 is balanced.

[0018]

[Math 2]

In the first pressure regulating valve 146, the hydraulic pressure Pin in the primary side hydraulic cylinder 80 is equal to the second line hydraulic pressure Pl2.
(In a steady state, Pl2 = secondary hydraulic cylinder 82 internal hydraulic pressure Pout), the first plunger 20
6 and the second plunger 208 are spaced apart, and the thrust by the hydraulic pressure Pin in the primary side hydraulic cylinder 80 acts in the valve closing direction of the spool valve element 200. When the oil pressure is lower than Pl2, the first plunger 206 and the second plunger 208
Because of this, the thrust force due to the second line oil pressure Pl2 acting on the end surface of the second plunger 208 acts on the spool valve element 200 in the valve closing direction. That is, the primary hydraulic cylinder 80 internal hydraulic pressure Pin and the second line hydraulic pressure P
The second plunger 208 receives the oil pressure from the spool valve 20 and applies an acting force based on the higher of these oil pressures to the spool valve 20.
It acts in the valve closing direction of 0. As a result, the optimum first line oil pressure is generated even during positive torque driving and negative torque driving.

The throttle pressure Pth is the required output, that is, the actual throttle valve opening θ in the engine 10.
It corresponds to th, and the throttle valve opening detection valve 2
28. Further, the gear ratio pressure Pr represents the actual gear ratio of the CVT 14, and is generated by the gear ratio detection valve 232. The throttle valve opening detection valve 228 and the gear ratio detection valve 232 are configured similarly to that described in, for example, Japanese Patent Application Laid-Open No. 64-49749.
is a plunger 230 that engages with a cam surface of a cam (not shown) that is rotated together with the throttle valve of the engine 10.
, and generates a throttle pressure Pth as shown in FIG. 4 in response to a thrust corresponding to the displacement of the plunger 230. The gear ratio detection valve 232 also includes a detection rod 234 that slides on the input side movable rotating body 76 of the CVT 16 and is moved in the axial direction by an amount of displacement equal to the amount of displacement in the axial direction. By changing the relief amount on the downstream side of the throttle 236 in accordance with this, a gear ratio pressure Pr as shown in FIG. 5 is generated.

Here, the gear ratio detection valve 232 generates the gear ratio pressure Pr by changing the release amount of the hydraulic oil of the second line hydraulic pressure Pl2 supplied from the second line oil passage 152 through the throttle 236. Because it is a thing,
While the gear ratio pressure Pr is limited to a value equal to or higher than the second line oil pressure Pl2, the second pressure regulating valve 150, which operates according to Equation 1, increases the second line oil pressure Pl2 as the gear ratio pressure Pr increases. reduce Therefore, when the gear ratio pressure Pr increases to a predetermined value and becomes equal to the second line oil pressure Pl2, both are saturated and constant from then on, and as shown in FIG. 3, the change characteristics of the basic output pressure Pmec are , in the process where the gear ratio γ changes from the maximum value to the minimum value, it initially decreases linearly in response to the increase in the gear ratio pressure Pr, but after reaching the gear ratio pressure Pr, it becomes a constant value. Becomes a characteristic.

The third pressure regulating valve 240 is connected to the forward/reverse switching device 14.
The hydraulic pressure and first electromagnetic valve 1 actuate the reverse brake 64 and the forward clutch 62 that are frictionally engaged in the
02, which generates a third line hydraulic pressure Pl3 that functions as a base pressure lot hydraulic pressure for the second solenoid valve 104 and the linear solenoid valve 106, for example, in Japanese Patent Application Laid-Open No. 64-4974.
The structure is similar to that described in Publication No. 9. The third line oil pressure Pl3 is adjusted to an optimal value so that the forward clutch 62 or the reverse brake 64 has a torque capacity that can reliably transmit torque without slippage.

The speed change control valve unit 250 includes a CVT 14
In order to adjust the gear ratio γ, one and the other of the primary hydraulic cylinder 80 and the secondary hydraulic cylinder 82 are controlled using the first line hydraulic pressure Pl1 and the second line hydraulic pressure Pl2 as source pressures. This speed change control valve unit 250 is composed of a speed change direction switching valve 252 and a flow rate control valve 254. Note that the third line oil pressure Pl3 is used as a pilot pressure for driving the speed change direction switching valve 252 and the flow rate control valve 254.

The speed change direction switching valve 252 is the first solenoid valve 102.
The flow rate control valve 254 includes a spool valve 258 whose position is controlled by the second electromagnetic valve 104. For example, when the first solenoid valve 102 is in the on state and the second solenoid valve 104 is in the on state, the spool valves 256 and 25
8 are both located on the upper side, so the first line oil passage 1
The hydraulic oil in 44 is supplied to the speed change direction switching valve 252, the first connection oil passage 260, the flow rate control valve 254, and the secondary oil passage 262.
The hydraulic oil in the primary hydraulic cylinder 80 flows through the primary oil passage 264, the flow rate control valve 254, the second connecting oil passage 266, and the shift direction switching valve. 332 and is discharged to the drain, and the gear ratio γ of the CVT 16 is quickly changed in the deceleration direction.

In this state, when the second solenoid valve 104 is turned off, the spool valve element 258 is positioned at the lower side, so that the first solenoid valve 258 passes through the bypass oil passage 272, which is provided with the one-way valve 268 and the throttle 270 in parallel. The hydraulic oil in the 2-line oil passage 152 is made to flow into the secondary hydraulic cylinder 82, while the hydraulic oil in the primary hydraulic cylinder 80 is discharged to the drain through the gap between the sliding parts, and the gear ratio γ of the CVT 16 is changed.
is gradually changed in the direction of deceleration.

On the other hand, when the first solenoid valve 102 is off and the second solenoid valve 104 is on, the hydraulic oil in the first line oil passage 144 flows through the relatively large throttle 278.
, the shift direction switching valve 252, the second connection oil passage 266, the flow rate control valve 254, and the primary oil passage 264 to flow into the primary hydraulic cylinder 80, while the hydraulic oil in the secondary hydraulic cylinder 82 flows into the secondary hydraulic cylinder 80. Next side oil passage 262, flow rate control valve 254, first connection oil passage 260, and speed change direction switching valve 2
52 and is discharged to the second line oil passage 152, and C
The gear ratio γ of the VT16 is quickly changed in the direction of speed increase.

Furthermore, when the second solenoid valve 104 is turned off in this state, the hydraulic oil in the first line oil passage 144 is
The hydraulic oil in the secondary hydraulic cylinder 82 is caused to flow into the primary hydraulic cylinder 80 through the third connecting oil passage 276 equipped with the speed change direction switching valve 252 and the throttle 274.
The oil is discharged to the second line oil passage 152 through the bypass oil passage 272, and the gear ratio γ of the CVT 16 is gradually changed in the speed increasing direction. Shift control is performed by selectively switching between these four shift modes. Note that the ON and OFF marks given to the first solenoid valve 102 and the second solenoid valve 104 in FIG. Compatible. Further, in the first solenoid valve 102 and the second solenoid valve 104, the ball-shaped valve closes the drain port and communicates the input port with the output port in the energized state.
It is a three-way switching valve that outputs line oil pressure Pl3, and in a non-excited state, a ball-shaped valve closes an input port and communicates an output port with a drain port.

The linear solenoid valve 106 reduces the signal pressure Psol by reducing the third line oil pressure Pl3.
A spool valve element 302 that generates L and is biased in the valve closing direction by a spring 300, an oil chamber 304 that receives a signal pressure PsolL in order to apply feedback oil pressure to the spool valve element 302, and a continuous A solenoid 308 having a core 306 that applies a thrust in the valve opening direction that changes to the spool valve element 302,
The signal pressure Ps has the characteristic shown in FIG.
olL is supplied to the oil chamber 182 of the second pressure regulating valve 150. The urging force of the spring 300 is WL1, the spool valve 3
The pressure receiving area of 02 is AL1, and the thrust of solenoid 306 is F.
When L1 is assumed, the above signal pressure PsolL is controlled according to Equation 3.

[0029]

[Math 3]

[0030] The switching valve 310 is connected to the signal pressure Psol.
L is supplied to the oil chamber 182 and oil chamber 19 of the second pressure regulating valve 150.
0, a spring 314 that biases the spool valve 312 to the normal mode side, and an output pressure from the third solenoid valve 108 (third line oil pressure Pl3). The oil chamber 316 receives the oil. Therefore, when the third solenoid valve 108 is in the OFF state, the spool valve element 312 is positioned on the normal mode side according to the urging force of the spring 314, and the signal pressure Pso
Since 1L is applied to the oil chamber 182, the second line oil pressure Pl2 can be optimally controlled. However, when the third solenoid valve 108 is turned on, the spool valve element 312 is moved to the tension control pressure up mode side against the urging force of the spring 314, and the signal pressure PsolL is not applied to the oil chamber 190. Therefore, in order to increase the transmission capacity of the power transmission belt 70 when the drive wheel 24 changes from a slip state to a grip state, the second line oil pressure Pl2 is increased according to Equation 4. Note that the large diameter land 186 of the plunger 166
The cross-sectional area of is A4'.

[0031]

[Math 4]

The electronic control device 110 controls the speed change of the CVT 16, the tension control pressure control, the control to increase the clamping force when the drive wheels 24 change from a sliding friction state (slip state) to a static friction state (grip state), etc. Repeat as necessary. The above-mentioned speed change control is determined in advance so as to obtain the fuel efficiency and drivability of the engine 10, and is
The target input shaft rotational speed Nin° is determined from the shift curve stored in 114 based on the actual throttle valve opening θth and the vehicle speed SPD, and the actual input shaft rotational speed Nin matches the target input shaft rotational speed Nin°. By selecting one of the speed change modes and driving the first solenoid valve 102 and the second solenoid valve 104, the speed ratio γ is adjusted by the speed change control valve unit 250. The above-mentioned speed change diagram is, for example, one selected in advance from a plurality of types of speed change diagrams corresponding to the driving range selected by the shift lever 126, similar to the one described in Japanese Patent Application Laid-Open No. 60-205067. used.

FIG. 7 is a functional block diagram illustrating the main components of the functions of this embodiment. In the figure, the drive wheel 24
The drive wheel friction state change detection means 3 corresponding to the electronic control device 110 detects the change from the grip state to the slip state.
20, the belt clamping force of the transmission belt 70 is temporarily increased for a predetermined period by the belt clamping force increasing means 322 corresponding to the third electromagnetic valve 108 and the switching valve 310. For this reason, when the drive wheel 24 suddenly regains grip with the road surface by releasing the accelerator after slipping, the rotational inertia generated by the sudden drop in rotation is applied to the output torque of the engine 10. Even if the transmission torque of the transmission belt 70 increases rapidly due to the addition, slippage of the transmission belt 70 is suitably prevented.

FIG. 8 is a flowchart illustrating the operation of the clamping force increase control of the electronic control device 110. In step SA1 in the figure, the rotational acceleration (amount of change in rotational speed per unit time) G of the driving wheels 24 is, for example, the difference (SPD-SPD-) between the vehicle speed SPD-1 in the previous control cycle and the vehicle speed SPD in the current control cycle. Calculated from 1). In the following step SA2, it is determined whether the content of the flag F1 is "1". This flag F1
When the content is "1", it indicates that the drive wheel 24 is in a slipping state, that is, in a sliding friction state. Initially, the content of flag F1 is "0", so
If the judgment in step SA2 is negative, step SA3
is executed. In step SA3, it is determined whether the rotational acceleration G of the drive wheels 24 is greater than a pre-stored determination reference value G1. This determination reference value G1 is a predetermined value for detecting slip of the drive wheels 24 on the road surface. If the determination at step SA3 is negative, the slip state of the drive wheels 24 has not occurred, and therefore the normal control mode for the second line oil pressure Pl2 is selected at step SA12. In this normal control mode, a necessary and sufficient tension control pressure is generated by the tension control pressure control routine shown in FIG.

FIG. 9 is a flowchart illustrating the operation of the electronic control unit 110 in tension control pressure control. In step SS1 of the figure, the latest engine rotational speed Ne is
, throttle valve opening θth, input shaft rotational speed Nin, and output shaft rotational speed Nout are read, and the gear ratio γ (=Nin/Nout) is calculated. In the following step SS2, the actual output torque Te of the engine 10 is calculated based on the engine rotational speed Ne and the throttle valve opening θth from a well-known relationship stored in advance in the ROM 114.

When the output torque Te of the engine 10 is calculated as described above, in step SS3, the actual gear ratio γ, the output torque of the engine 10 (the input torque of the CVT 16) )T
The ideal pressure Popt is calculated based on e and the output shaft rotational speed Nout. This ideal pressure Popt is the minimum theoretical optimum pressure for transmitting the input torque without causing the transmission belt 70 to slip. In addition, formula 5
The second term on the right side is a correction term for centrifugal oil pressure, and the third term on the right side is a margin value. Further, k1 and k2 in Equation 5 are constants.

[0037]

[Math 5]

In the following step SS4, the second pressure regulating valve 15
The basic output pressure Pmec of 0 is calculated based on the actual gear ratio γ and throttle valve opening θth from a pre-stored relationship shown in FIG. 3, for example. Then, step S
In S5, the difference between the basic output pressure Pmec of the second pressure regulating valve 150 and the ideal pressure Popt, that is, the basic output pressure Pm
The oil pressure reduction value Pdown to be lowered from ec is calculated from Equation 6. This oil pressure reduction value Pdown is a value for making the second line oil pressure Pl2 match the ideal pressure Popt. Then, in step SS6, a drive current value IsolL for causing the linear solenoid valve 106 to generate a control pressure PsolL equal to the oil pressure reduction value Pdown.
is determined, for example, from the pre-stored relationship shown in FIG. 6, and in step SS7, the drive current value IsolL
is output. As a result, the second line oil pressure Pl2 is made equal to the ideal pressure Popt.

[0039]

[Math 6]

Returning to FIG. 8, if the determination at step SA3 is affirmative, it means that the drive wheels 24 are in a slip state, so the content of the flag F1 is changed to "1" in step SA4.
1'', the up mode of the second line oil pressure Pl2 is selected in step SA5. This second
In the line oil pressure Pl2 up mode, the third solenoid valve 108, which has been in the OFF state, is switched to the ON state, and the drive current value IsolL is set to the maximum value. As a result, the third line pressure Pl3, which is the maximum signal pressure PsolL, is supplied to the oil chamber 190, and the second line oil pressure Pl2, which is related to the transmission capacity or tension of the transmission belt 70, is made higher than the basic pressure Pmec. That is, in this up mode, the line indicating the second line oil pressure Pl2 in FIG. 3 is shifted upward.

As described above, when the drive wheel 24 is in a slip state and the flag F1 is set, the determination at step SA2 in the next cycle is affirmative, so the content of the flag F2 is set to "1" at step SA6. It is determined whether or not there is. The content of this flag F2 is “1”.
'' indicates that the drive wheels 24 are about to grip the road surface, that is, the friction state of the drive wheels 24 is about to change from a sliding friction state to a static friction state. Initially, the content of the flag F2 is not "1" and the determination in step SA6 is negative, so it is determined in step SA7 whether the rotational acceleration G of the drive wheel 24 is smaller than a pre-stored determination reference value G2. be done. This judgment reference value G2 is a negative value smaller than the judgment reference value G1, and is a value for detecting the start of grip of the drive wheels 24 by the road surface.

As long as the slip state of the drive wheels 24 continues, the judgment in step SA7 is denied, so step SA5 is continuously executed. However, when the accelerator pedal is released, the drive wheels 24 are When the rotational speed NFR is about to be gripped by
At A8, the contents of the flag F2 are set to "1", and at step SA9, the timer T starts counting, and then step SA5 is continuously executed. Therefore, in the next cycle, since the determinations in steps SA2 and SA6 are affirmative, it is determined in step SA10 whether or not the time count by timer T is smaller than a pre-stored determination reference value To. This judgment reference value To corresponds to the duration of time from when the grip of the driving wheels 24 in the up mode of the second line oil pressure Pl2 is determined, and is sufficiently longer than the period until the grip of the driving wheels 24 ends. It is set to a large value, for example several seconds.

As long as the time measurement content of timer T is smaller than the judgment reference value To, the judgment in step SA10 is affirmed, so the execution of step SA5 is continued. However, if the time measurement content of timer T is smaller than the judgment reference value To. When the step SA1 is reached, the determination at step SA10 is denied, so the step SA1
After the contents of flags F1 and F2 are each reset to "0" in step SA12, the second line oil pressure normal mode is selected again in step SA12.

When driving on a road surface with low frictional resistance such as a snow-covered road or frozen road, or when driving on a slope, when the accelerator pedal is depressed as shown at time A in the time chart of FIG. A related slippage of the drive wheels 24 occurs. When the start of slipping of the drive wheels 24 is detected as shown at time B in FIG. Since the switching valve 310 is switched by turning on the third solenoid valve 108, the second line oil pressure Pl2 is made higher than the basic pressure Pmec, and the torque transmission of the transmission belt 70 is attempted to be increased. Normally, the driver returns the accelerator pedal in an attempt to eliminate the slip condition, and in connection with this, the drive wheels 24 begin to grip the road surface, resulting in a decrease in the rotational speed NFR and a rotational acceleration that is the rate of change thereof. A decrease in G occurs. Such an up mode of the second line oil pressure Pl2 is activated when a predetermined time To has elapsed since the start of grip of the drive wheels 24 was detected as shown at point C in FIG. It is temporarily increased until point E.

Therefore, according to this embodiment, the drive wheels 2
4 suddenly regains grip with the road surface after slipping, the rotational inertia generated by the sudden drop in rotation is added to the output torque of the engine 10, causing the transmission torque of the transmission belt 70 to decrease. Even if the sudden increase occurs, the belt squeezing force of the transmission belt 70 is temporarily increased by the second line oil pressure Pl2 for a predetermined period that is sufficient to cover the period until the grip of the drive wheel 24 is completed. This effectively prevents slipping.

Next, another embodiment of the present invention will be described. In the following description, parts common to those in the above-described embodiments are designated by the same reference numerals, and the description thereof will be omitted. In this example,
A rear wheel rotation sensor (not shown) is provided to detect the rotation speed of the rear wheel of the vehicle, which is a non-driving wheel. FIG. 11 is a flowchart showing another operation of the electronic control device 110, and the parts different from the flowchart of FIG. 8 will be mainly explained. In step SB1 of the figure, the driving wheel 24
rotational speed NFR of and rotational speed NR of non-driving wheels (not shown)
The difference with R (=NFR-NRR), that is, the rotational speed difference Δ
N is calculated. In step SB3, it is determined whether or not the rotational speed difference ΔN is larger than a predetermined determination reference value ΔN1 stored in advance. Time B in FIG. 10 corresponds to this determination. This judgment reference value ΔN1 is the driving wheel 2
This is for detecting a change from the grip state No. 4 to the slip state. Also, in step SB7,
It is determined whether or not the rotational speed difference ΔN is smaller than a predetermined determination reference value ΔN2 stored in advance. Time D in FIG. 10 corresponds to this judgment. This judgment standard value ΔN2
is a value smaller than the determination reference value ΔN1, and is used to determine that the slip state of the drive wheels 24 has become smaller due to completion of recovery of the grip of the drive wheels 24 on the road surface.

In this embodiment as well, when it is detected that the drive wheel 24 has changed from the grip state to the slip state as shown at time B in FIG. Since the second line oil pressure Pl2 is temporarily increased for a predetermined period of time until time E in FIG. Even if the amount increases rapidly, slippage of the transmission belt 70 is suitably prevented.

Although one embodiment of the present invention has been described above with reference to the drawings, the present invention can also be applied to other aspects. For example, in the embodiment shown in FIG.
Although the start of the grip state of the drive wheel 24 in the slip state was detected based on the rotational acceleration G of 4, a member rotating together with the drive wheel 24, for example, the output shaft
, the rotational acceleration of the input shaft 44 or the like may be detected.

[0049] Furthermore, the rotational speed difference ΔN (=NFR−NRR
The start of grip of the drive wheels 24 may be detected based on whether the rate of change in ) exceeds a predetermined criterion value.

Furthermore, in the embodiment shown in FIG. 11 described above,
Although the second line oil pressure Pl2 is temporarily increased until a predetermined period T1 has elapsed after completion of the grip of the drive wheels 24 is determined in step SB7, the period T1 may be set to zero. That is, step SB3
Step SB starts from the judgment of the slip start of the drive wheels 24 based on
The second line oil pressure Pl2 may be temporarily increased until it is determined that the grip recovery of the drive wheels 24 is completed.

Further, in the embodiment shown in FIG. 8, in order to prevent the transmission belt 70 from slipping due to the grip of the drive wheel 24 in the slipping state, the second line hydraulic pressure Pl is
2 was temporarily raised to a pressure higher than the basic pressure Pmec of the second pressure regulating valve 150, but the basic pressure Pmec
It may even be raised to . In this case,
In order to temporarily raise the second line oil pressure Pl2 to the basic pressure Pmec, it is only necessary to set the signal pressure PsolL output from the linear solenoid valve 106 to the minimum value, and
There is an advantage that the solenoid valve 108 and the switching valve 310 are not required.

Furthermore, in the above embodiment, the second line hydraulic pressure Pl2 regulated by the second pressure regulating valve 150 was applied to the CVT 16 through the speed change control valve unit 250, but the secondary hydraulic cylinder of the CVT 16 A hydraulic circuit of the type that acts directly on 82 may also be used.

Further, in the above-mentioned embodiment, the second line oil pressure Pl2 was regulated by decreasing the basic pressure Pmec by Pdown, but the electronic control unit 11
A pressure regulating valve that directly generates the second line oil pressure Pl2 according to a signal from 0 may be used.

The above-mentioned embodiment is merely one embodiment of the present invention, and various modifications may be made to the present invention without departing from the spirit thereof.

[Brief explanation of the drawing]

FIG. 1 is a diagram illustrating the configuration of a vehicle power transmission device including an embodiment of the present invention.

FIG. 2 is a diagram showing a main part configuration of the hydraulic control circuit of the embodiment shown in FIG. 1;

FIG. 3 is a diagram showing the output characteristics of the second pressure regulating valve in FIG. 2;

FIG. 4 is a diagram showing the output characteristics of the throttle valve opening detection valve of FIG. 2;

FIG. 5 is a diagram showing the output characteristics of the gear ratio detection valve of FIG. 2;

6 is a diagram showing the output characteristics of the linear solenoid valve of FIG. 2. FIG.

FIG. 7 is a diagram illustrating main parts of the functional configuration of the embodiment in FIG. 1;

FIG. 8 is a flowchart illustrating the main part of the operation of the electronic control device of the embodiment of FIG. 1;

9 is a flowchart illustrating the operation of a tension control pressure control routine executed in the second line oil pressure normal mode of FIG. 8;

10 is a time chart illustrating the state of the vehicle obtained by the control illustrated in the flowchart of FIG. 8; FIG.

FIG. 11 is a diagram corresponding to FIG. 8 in another embodiment of the present invention.

[Explanation of symbols]

10 Engine 16 Belt type continuously variable transmission 66, 68 Variable pulley 70 Transmission belt 320 Drive wheel friction state change detection means 322 Belt clamping force increasing means

Claims (1)

[Claims] [Claim 1] A vehicle belt type vehicle comprising a pair of variable pulleys around which a transmission belt is wound, and by changing the effective diameter of the variable pulleys, the rotation of the engine is transmitted to the drive wheels in a stepless manner. A control device for a belt-type continuously variable transmission for a vehicle, which controls the clamping force of the transmission belt according to input torque and gear ratio in the continuously variable transmission, wherein the drive wheel changes from a grip state to a slip state. drive wheel friction state change detection means for detecting a change in the state of the belt; and when the drive wheel friction state change detection means detects that the drive wheel has changed from a grip state to a slip state, the belt clamping force is adjusted to a predetermined value. 1. A control device for a belt-type continuously variable transmission for a vehicle, comprising: means for increasing belt clamping force temporarily for a period of time.