JPH0548388B2 - - Google Patents

Description

[Detailed description of the invention] [Industrial application field]

The present invention relates to a hydraulic control device for a belt-type continuously variable transmission for a vehicle, and more particularly, to countermeasures against line pressure control abnormalities in electronic control of line pressure.

[Conventional technology]

Regarding the control of this type of continuously variable transmission, for example, there is a basic method shown in Japanese Unexamined Patent Publication No. 55-65755, which describes mechanical speed change and line pressure control. However, with mechanical control, the structures of valves and hydraulic control systems are complicated, and input information and control are inherently limited. Therefore, in recent years, there has been a trend to electrically process various information and requirements to optimally electronically control speed change and line pressure. The continuously variable transmission is based on the premise that power is transmitted through friction between pulleys and belts. For this reason, electronic line pressure control uses appropriate input information to accurately estimate the transmission torque of the pulley and belt.
It is desirable to control this transmitted torque so as to always apply the necessary minimum pulley pressing force that does not cause belt slip. Conventionally, regarding line pressure electronic control of the above-mentioned continuously variable transmission, there is a prior art, for example, Japanese Patent Application Laid-Open No. 161347/1983, which shows that line pressure is controlled in relation to engine torque and the like.

[Problem to be solved by the invention]

By the way, the above-mentioned prior art is applied when the line pressure control is performed normally, and cannot deal with the case where the line pressure control becomes abnormal and decreases. In electronic control of line pressure, there is a risk that abnormalities may occur in the control due to various causes, resulting in an abnormal drop in line pressure. If power is transmitted in the same way as normal when the line pressure is abnormally low, serious damage such as belt slip will occur. For this reason, it is desirable to appropriately detect whether or not there is an abnormal drop in line pressure, and to provide a failsafe so that the vehicle can run without belt slip in the event of an abnormal drop. SUMMARY OF THE INVENTION In view of the above, an object of the present invention is to provide a hydraulic control system for a continuously variable transmission that allows a vehicle to run without causing belt slip when line pressure control is abnormal.

[Means to solve the problem]

In order to achieve this purpose, the configuration of the present invention will be explained with reference to FIG.
A reducing valve 60 that generates a constant reducing pressure is provided in the oil passage 31 branching from the oil passage 31 via the flow rate restriction means 32, and the oil passage 3 of the reducing pressure is
3 is communicated with a solenoid valve 64 that exhausts reducing pressure and converts it into signal hydraulic pressure in response to an electric signal from the control unit 70, an oil passage 37 for the signal oil pressure is communicated with the line pressure control valve 40, and an oil passage 37 for the reducing pressure is communicated with the line pressure control valve 40. 33 is provided with a hydraulic switch 90 for detecting abnormality in line pressure. The control unit 70 also includes a means 94 for calculating the clutch torque at the time of clutch start and direct engagement based on signals of engine speed, slot opening and secondary pulley speed, and means for determining an abnormal drop in line pressure based on a signal from the hydraulic switch 90. 91, means 96 for limiting the clutch torque to below the torque that can be transmitted at the lowest line pressure when line pressure is abnormal, and means 95 for outputting an electric signal to the clutch 2 in accordance with the clutch torque. .

[Effect]

In the present invention having the above configuration, the continuously variable transmission basically converts the reducing pressure into a signal oil pressure as a source pressure using the solenoid valve 65 in response to an electric signal from the control unit 70, and sends this signal oil pressure to the line pressure control valve 40. The line pressure is controlled electronically. The presence or absence of an abnormality in the line pressure control is always detected by the hydraulic switch 90 in the reducing pressure oil passage 33. Therefore, if the reducing pressure abnormally decreases together with the line pressure, it is determined that there is an abnormality and the clutch torque of the clutch 2 is limited to the torque that can be transmitted at the lowest line pressure. As a result, the transmission of engine power is reduced due to slippage of the clutch 2, belt slip is reliably prevented, and the vehicle can be driven at low speeds.

【Example】

Embodiments of the present invention will be described below based on the drawings. Referring to FIG. 2, an outline of a continuously variable transmission to which the present invention is applied will be explained. First, the drive system will be described. An engine 1 is connected to a main shaft 5 of a continuously variable transmission 4 via, for example, an electromagnetic clutch 2 capable of controlling clutch torque and a forward/reverse switching device 3. In the continuously variable transmission 4, a subshaft 6 is arranged parallel to a main shaft 5, a primary pulley 7 is provided on the main shaft 5, a secondary pulley 8 is provided on the subshaft 6, and a drive belt 11 is connected to both pulleys 7, 8. Can be wrapped. Both pulleys 7 and 8 are connected to the fixed side and hydraulic cylinders 9 and 10.
and a movable side that is movable in the axial direction, the pulley interval is variable, and the primary cylinder 9 has a larger pressure receiving area than the secondary cylinder 10. and a control unit 70 that outputs electrical signals to control line pressure and conversion;
It is equipped with a hydraulic circuit 20 that changes the primary pressure and the secondary line pressure based on an electric signal, and the line pressure and the primary pressure are electrically controlled. The belt is properly clamped by line pressure, and the ratio of the winding diameter of the drive belt 11 to the pulleys 7 and 8 is changed by the primary pressure to achieve stepless speed change. The subshaft 6 also has a set of reduction gears 12,
It is connected to the output shaft 14 via 13. A drive gear 15 of the output shaft 14 is configured to be transmitted to drive wheels 19 via a final gear 16, a differential gear 17, and an axle 18. The electromagnetic clutch 2 is automatically connected/disconnected and the clutch torque is variably controlled by electric signals from the control unit 70. Referring to FIG. 3, a hydraulic control system including the hydraulic circuit 20 will be explained. First, it has an oil pump 21 driven by the engine 1, and a line pressure oil passage 22 on the discharge side of the oil pump 21 is communicated with the secondary cylinder 10, the lie pressure control pressure 40, and the speed change speed control valve 50, and the speed change control valve A valve 50 is communicated with the primary cylinder 9 via an oil passage 23. Oil passage 24 on the drain side of the speed change control valve 50
is connected to an oil pan 26 and has a check valve 25 that prevents air from entering when draining oil from the primary cylinder 9. Further, the oil passage 27 on the drain side of the line pressure control valve 40 is provided with a regeneration valve 28 that generates a constant lubricating pressure. The oil passages 23 to the primary cylinder 9 are communicated via the prefilling valves 30, respectively. Line pressure control pressure 40 is applied to valve body 41 on spool 4
2 is movably inserted, and this spool 42 communicates the port 41a of the oil passage 22 with the drain port 41b to regulate the pressure. One port 4 of spool 42
Line pressure PL acts on 1c through an oil passage 46 branching from the oil passage 22. Also, on the other side of the spool 42 is a control port 41 that electrically controls line pressure.
d is provided, and a spring 43 that mechanically sets the minimum line pressure PLmin is biased. A duty pressure Pd for line pressure control is supplied to the control port 41d as a signal hydraulic pressure through an oil passage 47. Also, the end of the spring 43 is attached to an adjusting screw 44.
It is supported by a block 45 having a spring 4
By adjusting the set load in step 3, it is possible to adjust the relationship between the duty ratio D and the minimum line PLmin due to variations in each component. Therefore, line pressure PL, effective area SL, duty pressure Pd, effective area Sd, spring load Fs
Then, the following formula holds true. Fs + Pd・Sd=PL・SL PL=(Pd・Sd+Fs)/SL Therefore, line pressure PL is controlled to change continuously as a function of spring load Fs and duty pressure Pd, and increases with duty pressure Pd. Controlled by functions. The variable speed control valve 50 has a valve pair 51 and a spool 5.
2 is movably inserted, and by moving the spool 52 left and right, the oil supply position where the port 51a of the oil passage 22 is communicated with the port 51b of the oil passage 23, and the port 51
b and the oil drain position communicating with the drain port 51c. One port 52 of the spool 52
A constant reducing pressure is applied to d by the oil passage 53.
PR acts on the other port 51e, and the oil passage 54
Accordingly, duty pressure Pd for shift speed control acts as a signal hydraulic pressure. Further, a spring 55 for initial setting is urged on the spool 52 at the port 51e, and the spool 52 is configured to operate toward the oil sealing side when the duty pressure Pd is turned on. The duty pressure Pd of the signal hydraulic pressure is the duty ratio D of the electric signal using the reducing pressure PR as the source pressure.
It is generated in a pulse shape according to the Therefore, by changing the on/off ratio (duty ratio) of the duty pressure Pd, the oil supply and drain times, that is, the inflow and outflow flow rates, are changed, making it possible to control the speed change. That is, the shift speed di/dt is the primary cylinder 9.
Since the flow rate Q is a function of the duty ratio D, line pressure PL, and primary pressure Pp, the following equation holds. di/dt=f(Q)=f(D, PL, Pp) Here, line pressure PL is controlled by gear ratio i and engine torque T, and primary pressure Pp is determined by line pressure PL and gear ratio i, so T Assuming that is constant, di/dt=f(D,i). On the other hand, since the gear change speed di/dt is determined based on the deviation between the target gear ratio is and the actual gear ratio i in steady state, the following equation holds true. di/dt=k(is-i) From this, at each gear ratio i, the target gear ratio
If is is determined and the speed change speed di/dt is determined, the duty ratio D can be determined from the relationship between the speed change speed di/dt and the speed change ratio i. Therefore, it can be seen that if the shift speed control valve 50 is operated at this duty ratio D, the shift speed can be controlled over the entire shift range. Next, a circuit that generates the duty pressure Pd as the signal oil pressure will be explained. First, as a circuit for generating a constant reducing pressure PR, a line pressure oil passage 22 branches into an oil passage 31 via an orifice 32 that restricts the flow rate, and this oil passage 31 is communicated with a reducing valve 60. In the reducing valve 60, a spool 62 is movably inserted into a valve pair 61, and one of the spools 62 is biased by a spring 63. In addition, an inlet port 61a, an outlet port 61b, which communicates with the oil passage 31,
A drain port 61c is provided, and reducing pressure from the outlet port 61b acts through the oil passage 33 on the port 61d of the spool 62 on the opposite side from the spring 63. One block 64 of the spring 63
can be moved using an adjustment screw or the like to adjust the reducing pressure PR along with the spring load. As a result, the line pressure PL is restricted by the orifice 32 and supplied to the port 61a, and when the reducing PR of the oil path 33 decreases, the spring 6
3, the spool 62 communicates the ports 61a and 61b and introduces the line pressure PL. Further, due to the rise in the oil pressure in the port 61d, the spool 62 is returned to communicate the ports 61b and 61c to reduce the reducing pressure PR. Thus reducing pressure
The line pressure PL is replenished by the amount of decrease in PR, and a constant reducing pressure PR equal to the load of the spring 63 is always generated. The oil passage 33 for this reducing pressure PR is connected to a line pressure control solenoid valve 65, an accumulator 66 and a duty pressure Pd via an orifice 34.
The oil passage 37 is connected to the oil passage 37. and solenoid valve 6
5 intermittently drains a constant reducing pressure PR according to the duty signal to generate pulse-like oil pressure, smooths this oil pressure to a predetermined level of duty pressure Pd with an accumulator 66, and controls the line pressure from the oil path 37. Supplied to valve 40. Further, an oil passage 53 branches off from the upstream side of the orifice 34 of the oil passage 33, and has an orifice 35 in the middle of the oil passage 53 to communicate with a solenoid valve 67 for speed change control and the oil passage 54. and solenoid valve 67
Similarly, the data pressure Pd is increased by the data signal.
This duty pressure Pd is directly supplied to the speed change control valve 50 through the oil passage 54. The solenoid valve 65 is configured to drain oil when the duty signal is turned on. For this reason, the duty pressure Pd is generated with respect to the duty ratio D with the characteristic of a decreasing function. Since the solenoid valve 67 has a similar configuration, when the duty ratio D is large, the shift speed control valve 50
If it takes a long time to switch to the oil supply position, it will be shifted up, and vice versa, it will take a long time to switch to the oil drain position and it will be shifted down. The larger the deviation of is-i, the larger the change in duty ratio D, so that the shift speed is controlled faster. Here, the solenoid valve 65 is configured to drain oil when the duty signal is on, and therefore, the greater the duty ratio is, the smaller the duty pressure is. As a result, the line pressure has a characteristic that changes as a decreasing function with respect to the duty ratio. Further, referring to FIG. 4, an electric control system including a control unit 70 will be explained. First, the primary pulley rotation speed sensor 71 detects the primary pulley rotation speed Np, and the secondary pulley rotation speed Ns
a secondary pulley rotation speed sensor 72 that detects
It has an engine rotation speed sensor 74 that detects the engine rotation speed Ne and a throttle opening sensor 73 that detects the throttle opening degree θ. These sensor signals are input to a control unit 70. The transmission speed control system in the control unit 70 will be explained. It has an actual gear ratio calculation unit 75 into which the primary pulley point number Np and the secondary pulley rotation speed Ns are input, and calculates the actual gear ratio i by i=Np/Ns. Also, the secondary pulley rotation speed Ns
and the throttle opening θ are input to the target gear ratio search section 76, and the target gear ratio is is searched by a table of Ns-θ based on the gear shift pattern. The actual gear ratio i, the target gear ratio is in steady state, and the coefficient k of the coefficient setting unit 77 are calculated by the gear shifting speed calculating unit 78.
is input, and the shifting speed di/dt is calculated as di/dt=k(is-i). Further, if the sign of the shift speed di/dt is positive, it is determined to be a downshift, and if it is negative, it is determined to be a shift up. The shift speed di/dt and the actual gear ratio i are further input to a duty ratio search section 79, and a duty ratio D is set as a manipulated variable according to the shift speed di/dt and the actual gear ratio i. Here, the duty ratio D is D=f(di/dt, i)
According to the relationship, a table is set based on ±di/dt and i. In other words, the shift-up -di/
In the table of dt and i, for the value of D = 50 to 100%,
In the table of +di/dt and i for downshifting, D
It is divided into values from =50 to 0%. In the shift-up table, the duty ratio D is set as a decreasing function for i and as an increasing function for -di/dt. Furthermore, in the shift down table, the neutrality ratio D is set as an increasing function for i and as a decreasing function for di/dt. So, referring to this table, di/dt
The duty ratio D corresponding to i is searched. Then, an electric signal having a duty ratio D is outputted to the solenoid valve 67 via the drive section 80. Next, the line pressure control system will be explained. First, the control principle will be explained. This type of belt-type continuously variable transmission transmits torque through friction between the pulley and the belt while the belt is clamped by line pressure at the secondary pulley. Therefore, allowable input torque (torque that can be transmitted without slipping) Tmax, belt clamp Fs at secondary pulley, friction coefficient μ between pulley and belt, belt pitch radius R1 at primary pulley, belt pitch radius R2 at secondary pulley, pulley If the belt pinch angle α, line pressure PL, secondary piston effective pressure-receiving area As, pulley ratio (conversion ratio) i, pulley angle β at the primary pulley, and distance between pulley centers C, then the torque transmission capacity is the mechanics at the secondary pulley. From the viewpoint of balance, it can be simply expressed by the following formula. Tmax=Fs・2・μ・R1/cosα
...(1) Also, Fs=PL・As R2=R1−C・sinβ i=R2/R1 Therefore, R1=1/f(1−i), and R1 becomes a function of i. Also, in equation (1), 1/f'(i)=As・2・μ・/cosα・f(1−
i) Then, Tmax=PL/f′(i). Therefore, if we calculate the required run-in pressure PLu per unit torquer, assuming that it is the minimum line pressure that can transmit a constant torque without belt slip, then PLu=f'(i) becomes a function of i. That is, the required line pressure PLu per unit torque is set to be high at low speeds where the gear ratio i is large, and to decrease as the gear ratio i becomes smaller. Therefore, the required line pressure PLu per unit torque
By calculating the target run-in pressure PLt from and engine torque T, it becomes possible to accurately calculate the minimum necessary target run-in pressure PLt corresponding to the transmission torque of the pulley and belt. Therefore, in the run-in pressure control system, the throttle opening θ
and the engine rotation speed Ne are determined by the engine torque setting section 8.
1 and set the engine torque T by referring to the θ-Ne table. Further, the required line pressure setting section 82 sets the required line pressure PLu per unit torque according to the gear ratio i using a table of PLu-i. The engine torque T and the required line PLu per unit torque are input to the target line pressure calculating section 83, and the target line pressure PLt is calculated by PLt=PLu·T. The target run-in pressure PLt is input to the duty ratio setting section 84, and the duty ratio D as a manipulated variable corresponding to the target line pressure PLt is set in a decreasing function manner. Here, the line pressure control valve 40 is a pressure regulating valve, and the oil pump 21 is driven by the engine 1. Therefore, if the pump discharge amount changes depending on the engine rotation speed Ne, the actual line pressure will change even if the operation amount of the electric signal is the same, so it is necessary to correct the operation amount with respect to the engine rotation speed Ne. Therefore, the duty ratio D of the manipulated variable is the engine speed
The duty ratio D is set based on the relationship between Ne and the target line pressure PLt, and the duty ratio D is calculated by referring to this table.
Then, the electric signal of duty D is outputted to the solenoid valve 65 via the drive section 85. Furthermore, countermeasures against line pressure control abnormalities will be explained. First, a hydraulic switch 90 is provided in the reducing pressure oil passage 33 on the output side of the reducing valve 60. The signal from the oil pressure switch 90 is input to a line pressure abnormal drop determination section 91 of the control unit, and if the line pressure has dropped below a certain reducing pressure, it is determined that the line pressure has fallen abnormally. The control system of electromagnetic clutch 2 is based on the engine speed
Start detection section 9 where Ne and throttle opening θ are input
2, and a clutch direct connection part 93 to which the secondary pulley rotation speed Ns is input. Then, the start and direct connection signals are input to the calculation unit 94, which gradually increases the clutch current IC when starting, determines the clutch current Ic to be directly connected or disconnected depending on the vehicle speed when the clutch is directly connected, and drives this clutch current Ic. 95 to the clutch coil 2a. Therefore, a clutch torque limiting section 96 is attached to the output side of the calculating section 94, and a line pressure abnormal decrease determining section 9
1 signal is input. When an abnormal decrease signal is input, the clutch current Ic is corrected to be significantly reduced. In other words, in the worst case of line pressure control abnormality, the line pressure is controlled by the line pressure control valve 40.
PL is the lowest line pressure Pmin in FIG. 5, which is controlled only by the load of the spring 43. Furthermore, the gear ratio is maximum when the vehicle is running at low speed. Therefore, the transmission torque that can be transmitted without belt slip under the minimum line pressure and maximum transmission ratio is calculated in advance, and the clutch torque Tc is limited to below this transmission torque by the clutch current Ic. Next, the operation of this embodiment will be explained. First, the oil pump 2 is activated by the operation of the engine 1.
1 is driven, the line pressure PL of the oil passage 22 is supplied only to the secondary cylinder 10, and the gear ratio is set to the lowest gear, which is the maximum. At this time, a constant reducing pressure PR is taken out by the reducing valve 60 using the line pressure PL, and this reducing pressure PR is guided to each solenoid valve 65, 67, making it possible to electronically control the line pressure and speed change. Become. Further, signals of the primary pulley rotation speed Np, the secondary pulley rotation speed Ns, the throttle opening θ, and the engine rotation speed Ne are input to the control unit 70. In the line pressure control system, the actual gear ratio i is calculated from the primary pulley rotation speed Np and the secondary pulley rotation speed Ns, and the required line pressure PLu per unit torque is set according to this gear ratio i. In addition, the engine torque T is estimated from the engine speed Ne and the throttle opening θ, and the required line pressure PLu per unit torque is calculated as the engine torque T.
The target line pressure PLt is calculated by multiplying by Therefore, when the vehicle is stopped and idling, the maximum gear ratio is reached when the secondary pulley rotation speed Ns reaches zero, and the required line pressure PLu per unit torque is set to be large. Therefore, even if the engine torque T is small, the target line pressure PLt is calculated to be relatively large, and a signal with a small duty ratio D is output to the solenoid valve 65. Therefore, the amount of oil discharged from the solenoid valve 65 decreases and is converted into a large duty pressure Pd, and this duty pressure Pd is introduced into the port 41d of the line pressure control valve 40. Therefore, in the line pressure control valve 40, the load of the spring 43 that sets the minimum line pressure PLmin and the duty pressure Pd act in a direction to increase the line pressure PL, so that the line pressure PL is controlled to be high. Furthermore, when starting by pressing the accelerator, the electromagnetic clutch 2 is automatically connected while slipping by the control unit 70, the engine output is input to the continuously variable transmission 4, and the shifting power is transmitted to the drive wheels 19 to drive the vehicle. . Thereafter, the electromagnetic clutch 2 is completely directly engaged, and is automatically disconnected just before the vehicle stops during deceleration to prevent the engine from stalling. When the engine torque T increases at the time of starting or accelerating during driving, the target line pressure PLt is calculated to be even larger. Therefore, the duty ratio D is set smaller, and the line pressure control valve 40 controls the line pressure.
PL is controlled to be high by the amount of engine torque T. Further, as the vehicle speed increases, shift control is started, the gear ratio i becomes smaller, and the engine torque T also becomes smaller, so the duty ratio D becomes larger. Therefore, the duty pressure Pd in the solenoid valve 65 decreases due to an increase in the amount of discharged oil, and the line pressure PL in the line pressure control valve 40 is sequentially controlled to decrease. In this case, at the minimum gear ratio and the minimum engine torque, the duty pressure Pd becomes approximately zero and is controlled to the minimum line pressure PLmin by the spring load. In this way, as shown in FIG. 5, the line pressure PL is continuously electronically controlled to be lower as the gear ratio i is smaller and higher as the engine torque T is larger. Then, the required line pressure PLu corresponding to the gear ratio i makes the target line pressure PLt correspond to the transmission torque between the pulley and the belt, and the line pressure PL always maintains the necessary minimum pulley pressing force that does not cause belt slip. Granted. On the other hand, after the vehicle has started, the control unit 70 further calculates the shift speed di/dt using is-i.
A signal of the duty ratio D set according to the relationship between di/dt and i is output to the solenoid valve 67, and the duty ratio is
Converted to Pd. This duty pressure Pd is then introduced into the conversion speed control valve 50, which operates to supply and drain line pressure PL to and from the primary cylinder 9 at a predetermined flow rate to change the primary pressure Pp. Therefore, the actual gear ratio i follows the target gear ratio is, and the gear change is controlled appropriately according to the driving and driving conditions. In this case, as in the transient state, the larger is-i is, the faster the gear change speed is. . In the above-mentioned electronic control of line pressure and speed change, the reducing valve 60 is always controlled by the hydraulic switch 90.
The reducing pressure PR is detected. Therefore, if the line pressure PL itself decreases abnormally for some reason, the reducing pressure PR also decreases abnormally. Furthermore, if the line pressure control valve 40 acts in a direction to decrease the line pressure PL due to an abnormal decrease in the reducing pressure PR or duty pressure Pd, the reducing pressure PR will also decrease.
is abnormally reduced and detected by the oil pressure switch 90. At this time, an abnormal decrease in line pressure is determined by the line pressure abnormal decrease determination section 91 based on the signal from the oil pressure switch 90, and this abnormal decrease signal is input to the clutch torque limiter 96. Then, the clutch current Ic supplied to the clutch coil 2a by the calculating section 94 is immediately and significantly reduced. Therefore, the clutch torque Tc of the electromagnetic clutch 2 decreases according to the clutch current Ic, and even if the engine output is large, the transmission torque of the pulleys 7, 8 and the belt 11 will not cause belt slip due to the slippage of the electromagnetic clutch 2. limited to things. Therefore, belt slip is reliably prevented when line pressure control is abnormal. Also, electromagnetic clutch 2
Since the power is transmitted to the continuously variable transmission 4 according to the clutch torque Tc of the clutch torque Tc, the vehicle can be driven at a low speed at the maximum speed change. Although one embodiment of the present invention has been described above,
It is not limited to this only.

Effects of the Invention As described above, according to the present invention, in a hydraulic control device that electronically controls line pressure in a continuously variable transmission, the control unit determines the actual gear ratio based on the primary pulley rotation speed and the secondary pulley rotation speed. The required line pressure per unit torque is determined according to the gear ratio, and the target line pressure is calculated from the required line pressure per unit torque and engine torque, so line pressure is adjusted according to the transmitted torque to prevent belt slip. Can be controlled to the minimum necessary. Also, since hydraulic switches are used in oil passages that generate constant reducing using line pressure,
Abnormalities in the line pressure control system can be accurately detected. When line pressure control is abnormal. Since the clutch torque is corrected to be limited to the torque that can be transmitted without belt slip at the lowest line pressure, belt slip of the continuously variable transmission and accompanying damage to the belt etc. can be reliably prevented. In addition, since the transmission of engine output ensures that the vehicle can run at low speeds, the driver can drive himself and go to repairs.

[Brief explanation of the drawing]

FIG. 1 is a functional block diagram showing a hydraulic control system for a continuously variable transmission according to the present invention, FIG. 2 is a block diagram schematically showing a continuously variable transmission to which the present invention is applied, and FIG. 3 is a hydraulic control system according to the present invention. FIG. 4 is a hydraulic circuit diagram showing an embodiment of the control device, FIG. 4 is a block diagram of the electric control system, and FIG. 5 is a characteristic diagram of line pressure control. 2...Electromagnetic clutch, 4...Continuously variable transmission, 5
... Main shaft, 11 ... Drive belt, 6 ... Subshaft, 7
...Primary pulley, 8...Secondary pulley, 9...Primary cylinder, 10...Secondary cylinder, 21...Oil pump, 22,2
3, 31, 33, 37... Oil path, 32... Orifice, 40... Line pressure control valve, 50... Shift speed control valve, 60... Reducing valve, 65...
Solenoid valve, 70... Control unit, 90...
Hydraulic switch, 91...Line pressure abnormal drop determination section, 94...Calculation section, 95...Drive section, 96...
Clutch torque limiter.

Claims (1)

[Claims] 1. A main shaft is connected to the engine via a clutch that can control clutch torque, a primary pulley with variable pulley spacing is provided on this main shaft, and a subshaft on the wheel side arranged parallel to the main shaft has variable pulley spacing. A variable secondary pulley is provided, and a drive belt is wound between both pulleys to control the line pressure in the oil path from the hydraulic source and supply that line pressure to the cylinder of the secondary pulley to apply pulley pressing force. A line pressure control valve is provided to change the primary pressure by supplying and discharging line pressure to the oil path to the cylinder of the primary pulley, and a shift speed control valve is provided to change the primary pressure. In a continuously variable transmission that changes the speed steplessly by changing the ratio of the winding diameter of the drive belt, reducing pressure is generated in an oil passage 31 branching from a line pressure oil passage 22 via a flow rate restriction means 32 to generate a constant reducing pressure. A valve 60 is provided, and a reducing pressure oil passage 33 communicates with a solenoid valve 65 that discharges reducing pressure in response to an electric signal from a control unit 70 and converts it into signal oil pressure, and a signal oil pressure oil passage 37 is connected to a line pressure control valve. A hydraulic switch 90 is provided in the reducing pressure oil passage 33 to detect an abnormality in the line pressure. means 94 for calculating the clutch torque at the time of starting or direct connection; means 91 for determining an abnormal drop in line pressure based on a signal from the hydraulic switch 90; A hydraulic control device for a continuously variable transmission, comprising means 96 for limiting and means 95 for outputting an electric signal to the clutch 2 in accordance with clutch torque.