JPH04285361A - Control device for belt type continuously variable transmission for vehicle - Google Patents

Abstract

PURPOSE:To provide a control device for a belt type continuously variable transmission which will not cause slippage of a transmission belt even in the case where a driving wheel runs while slip and grip are repeated. CONSTITUTION:When the change width DELTAG of a rotary acceleration does not exceed DELTAGO, ordinary tension control pressure control is executed because an increase value DELTAPG is zero. When it exceeds DELTAGO, however, an increase value DELTAPG is determined according to the change width DELTAG and a speed change ratio gamma, so that the ideal pressure POPT is increased 4 the increase value DELTAPG and the second line oil pressure Pl2 regulated to agree with the ideal pressure POPT is also increased by the above increase. Accordingly, even it the instantaneous slip and grip of a driving wheel 24 are repeatedly caused according to the unevenness of the road surface in running on a bad road, and comparatively large transmission torque is repeatedly applied to a transmission belt 70 in relation thereto, the slippage of the transmission belt 70 can be 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 conditions where the clamping force of the power transmission belt is controlled as described above, if the drive wheels repeatedly slip and grip momentarily on a rough, uneven road, for example, the drive wheels may contact the road surface. There was a risk that the transmission belt might slip when the grip was suddenly regained. In other words, while the drive wheels are slipping, the rotational speeds of the engine, transmission joints such as clutches and torque converters, primary and secondary pulleys, axles, etc. increase rapidly, but the drive wheels are unable to grip the road surface. When this occurs, the drive wheels become in a state of static friction and the previous rotation rapidly decreases, and the rotational inertia generated by this sudden decrease in rotation becomes added to the engine's output torque, causing the transmission belt to A relatively large transmission torque will be applied repeatedly.

The present invention was made against the background of the above-mentioned circumstances, and its object is to provide a vehicle belt type vehicle in which the transmission belt does not slip even when the driving wheels repeatedly slip and grip. An object of the present invention is to provide a control device for a continuously variable transmission.

[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 rotational acceleration detection means for continuously determining rotational acceleration of the drive wheel;
(b) a change width determining means for determining a change width of the rotational acceleration of the drive wheel determined by the drive wheel rotational acceleration detection means; and (c) a rotational acceleration of the drive wheel determined by the change width determination means. and a belt clamping force increasing means for increasing the belt clamping force when the range of change exceeds a predetermined judgment reference value.

[0006]

[Operation] By doing this, the range of change in the rotational acceleration of the drive wheel determined by the drive wheel rotational acceleration detection means is determined by the range of change determination means, and the range of change in the rotational acceleration of the drive wheel is determined in advance. If the criterion value is exceeded,
The belt clamping force is increased by the belt clamping force increasing means.

[0007]

[Effects of the Invention] Therefore, even in a running state where a relatively large transmission torque is repeatedly applied to the transmission belt due to repeated instantaneous slips and grips of the drive wheels, the belt clamping force of the transmission belt is reduced by the belt clamping force increasing means. is increased, so slippage of the transmission belt 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 and the linear solenoid valve 106 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 oil 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 Pl1 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 decreasing value from ec.

[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 side hydraulic cylinder 82 internal hydraulic pressure Pout) If it is higher than the first plunger 2
06 and the second plunger 208 are separated, 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, but the hydraulic pressure Pin in the primary side hydraulic cylinder 80 is on the second line. When the oil pressure is lower than Pl2, the first plunger 206 and the second plunger 20
8 are in contact with each other, 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 second plunger 208, which receives the hydraulic pressure Pin in the primary side hydraulic cylinder 80 and the second line hydraulic pressure Pl2, applies an operating force based on the higher of these hydraulic pressures to the spool valve element 2.
00 in the valve closing direction. 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 control valve 254 includes a spool valve 256 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 control valve 254, the second connection oil passage 266, and the speed change 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 second pressure regulating valve 1 has the characteristic shown in FIG.
50 oil chambers 182. The urging force of the spring 300 is WL1, and the pressure receiving area of the spool valve 302 is AL.
1. If the thrust of the solenoid 306 is FL1, the above signal pressure PsolL is controlled according to Equation 3.

[0029]

[Math 3]

The electronic control device 110 repeatedly executes speed change control, tension control pressure control, etc. of the CVT 16 as appropriate. 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 stored in the ROM 11.
The actual throttle valve opening θ is determined from the shift curve stored in 4.
Target input shaft rotation speed N based on th and vehicle speed SPD
in° is determined, and one of the speed change modes is selected so that the actual input shaft rotational speed Nin matches the target input shaft rotational speed Nin°, and the first solenoid valve 102 and the second solenoid valve
By driving the electromagnetic 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 structure of the functions of this embodiment. In the figure, the drive wheel 24
When the rotational acceleration of the driving wheel 24 is detected by the driving wheel rotational acceleration detecting means 320, the change width of the rotational acceleration of the driving wheel 24 is determined by the change width determining means 322, and the rotational acceleration of the driving wheel 24 is determined by the change width determining means 322.
When the range of change in the rotational acceleration exceeds a predetermined criterion value, the belt clamping force increasing means 324 temporarily increases the belt clamping force. As a result, slippage of the transmission belt 70 is suitably prevented when driving on a rough road where a relatively large transmission torque is repeatedly applied to the transmission belt 70 due to repeated instantaneous slips and grips of the drive wheels 24.

FIG. 8 is a flowchart illustrating the tension control pressure control operation of the electronic control device 110. In step SA1 in the figure, the latest engine rotational speed Ne,
The throttle valve opening θth, the input shaft rotational speed Nin, and the output shaft rotational speed Nout are read, and the gear ratio γ (=Nin/Nout) is calculated. In the following step SA2, 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. And step SA3
For example, an increase value ΔPG for eliminating slippage of the transmission belt 70 caused by driving on a rough road is determined by the routine for determining the increase value ΔPG in FIG. 9, for example.

Step SS1 in FIG. 9 corresponds to the drive wheel rotational acceleration detection means 320, which detects the rotational acceleration of the drive wheel 24 (for example, the rotation per unit time of several milliseconds to several tens of milliseconds). For example, G is the difference between the vehicle speed SPD-1 in the previous control cycle and the vehicle speed SPD in the current control cycle (SPD-S
PD-1). In subsequent steps SS2 and SS3, the maximum value Gmax and minimum value Gmin of the rotational acceleration G within a certain period of about several seconds, for example, from several hundred previous control cycles to the current control cycle, are calculated, respectively. Next, in step SS4 corresponding to the change width determining means 322, the change width ΔG (=
Gmax-Gmin) is calculated. Then, in step SS5, an increase value ΔPG is calculated based on the actual rotational acceleration change range ΔG from the pre-stored relationship shown in FIG.
is determined. The relationship shown in FIG. 10 can be obtained by experimentally determining in advance the necessary and sufficient increase value ΔPG to eliminate the instantaneous slip of the drive wheels 24 and the slip of the transmission belt 70 caused by grip when driving on a rough road. It is a function of rotational acceleration change width ΔG and gear ratio γ.

Drive wheel vehicle speed SP when the vehicle is running on a rough road
D, the driving wheel rotational acceleration G, and the transmission torque (tension) of the transmission belt 70 are as shown in FIG.
The rotational acceleration change width ΔG is determined sequentially from the difference between the maximum peak Gmax and the minimum peak Gmin of the driving wheel rotational acceleration G pulsations within a certain period of time. It is required. Rotational acceleration change width Δ obtained sequentially in this way
G changes as shown in FIG. 12, for example.

When the increase value ΔPG is determined as described above, in step SA4 of FIG. 8, the actual gear ratio γ, engine 10
The ideal pressure Popt is calculated based on the output torque Te (input torque of the CVT 16), the output shaft rotational speed Nout, and the increase value ΔPG. This ideal pressure Po
pt is the minimum theoretical optimum pressure for transmitting the input torque without slipping of the transmission belt 70. Note that the second term on the right side of Equation 4 is a correction term for centrifugal oil pressure. Further, k1 and k2 in Equation 4 are constants.

[0036]

[Math 4]

In the following step SA5, 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 A6, 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 5. 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 SA7, 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 SA8, the drive current value IsolL
is output. As a result, the second line oil pressure Pl2 is made equal to the ideal pressure Popt.

[0038]

[Math 5]

As described above, in this embodiment, when the variation width ΔG of rotational acceleration does not exceed the predetermined value shown, for example, ΔGo in FIG. executed. However, when the variation width ΔG of the rotational acceleration exceeds ΔGo, the increase value ΔPG determined according to the variation width ΔG and the gear ratio γ is determined as shown in FIG. Since the pressure is increased by the value ΔPG and the pressure is regulated to match the ideal pressure Popt, the second line oil pressure Pl2 is also increased by the increased value ΔPG. In step SA4, the increase value ΔPG is added to the ideal pressure Popt to increase the second line oil pressure Pl2, thereby increasing the clamping force of the transmission belt 70.
A4 corresponds to the belt clamping force increasing means 324.

Therefore, in this embodiment, when driving on a rough road, repeated instantaneous slips and grips occur in the drive wheels 24 due to the unevenness of the road surface, and in connection therewith, a relatively large transmission torque is applied to the transmission belt 70. Even if the pressure is applied repeatedly, the second line oil pressure Pl2 is increased as described above, so that the transmission belt 70 is suitably prevented from slipping.

Furthermore, according to this embodiment, since the increase value ΔPG is determined according to the relationship shown in FIG. The advantage is that the second line oil pressure Pl2 is increased by a necessary and sufficient value to prevent the transmission belt 70 from slipping due to repetition of the above, and the power loss is minimized and the durability of the transmission belt 70 is increased. There is.

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 FIG. 13, in steps SB1 and SB2, the step SA
Similarly to SA1 and SA2, the input signal is read and the output torque Te of the engine 10 is calculated. In the following step SB3, the actual gear ratio γ and the output torque Te of the engine 10 are determined from the pre-stored relationship shown in Equation 6.
and the ideal pressure Po based on the output shaft rotational speed Nout.
pt is calculated. Note that the third term on the right side of Equation 6 is the margin pressure.

[0043]

[Math 6]

In steps SB4 to SB6, similarly to steps SS1 to SS4 described above, the drive wheel rotational acceleration G is
Then, in step SB7, it is determined whether or not the variation width ΔG is larger than a pre-stored predetermined criterion value ΔGo. If the determination in step SB7 is negative, the running state is such that fluctuations in rotational acceleration G due to minute slips and grips of the drive wheels 24 are small, so steps SB9 to SB12 similar to steps SA5 to SA8 are carried out.
is executed directly. However, if the determination in step SB7 is affirmative, in step SB8, a pre-stored constant value ΔPk is added to the target pressure Popt determined in step SB3, thereby updating the target pressure Popt. . This constant value ΔPk was obtained experimentally in advance in order to eliminate the momentary slip of the drive wheels 24 and the slip of the transmission belt 70 due to grip when driving on a rough road.

In this embodiment as well, when driving on a rough road, repeated instantaneous slips and grips occur in the drive wheels 24 due to unevenness of the road surface, and in connection therewith, a relatively large transmission torque is repeatedly applied to the transmission belt 70. Even when the pressure is applied, the second line oil pressure Pl2 is increased as described above, so that slipping 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 rotational acceleration G of 4 is calculated based on the vehicle speed signal SPD, it may be calculated based on the output shaft rotational speed Nout.

Furthermore, in the above embodiment, the second line oil pressure Pl2 was temporarily increased only while the variation range ΔG of the rotational acceleration G of the drive wheels 24 exceeded the predetermined value ΔGo. It may be increased by a preset period T after ΔG exceeds a predetermined value ΔGo.

Furthermore, in the embodiment shown in FIG. 13,
In step SB8, the ideal pressure Popt is changed by adding a constant increase value ΔPk to the second line oil pressure Pl.
2 was temporarily increased, but instead, a certain value is subtracted from Pmec in step SB9, and a certain subtracted value is changed to Pdown in step SB10.
Alternatively, a certain value may be subtracted from IsolL in step SB11.

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;

FIG. 9 is a flowchart illustrating the routine of step SA3 in FIG. 8;

FIG. 10 is a diagram showing relationships used in the flowchart of FIG. 9;

FIG. 11 is a time chart illustrating the range of change in driving wheel rotational acceleration in the embodiment of FIG. 8;

12 is a time chart showing the relationship between the driving wheel rotational acceleration change range and the second line oil pressure obtained by the operation of the embodiment shown in FIG. 8; FIG.

FIG. 13 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 Driving wheel rotational acceleration detection means 322 Variation range determining means 324 Belt clamping force increasing means

Claims (1)

[Claims] 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 that controls a clamping force of the transmission belt according to an input torque and a gear ratio in the continuously variable transmission, the control device continuously controlling the rotational acceleration of the drive wheel. a driving wheel rotational acceleration to be determined; a change range determining means for determining a change width of the rotational acceleration of the drive wheel obtained by the drive wheel rotational acceleration detecting means; and a change range determining means for determining the change width of the drive wheel determined by the change width determination means A belt-type continuously variable transmission for a vehicle, characterized in that the belt-type continuously variable transmission includes a belt-squeezing force increasing means for increasing the belt-squeezing force when the range of change in rotational acceleration exceeds a predetermined criterion value. Control device.