JPH03189455A - Control device for continuously variable transmission - Google Patents

Abstract

PURPOSE:To prevent a delay in control response and hunting of speed change ratio by compensating the line pressure for a drive pulley in accordance with a variation speed by a speed change ratio, and simultaneously by compensating a control gain for the feed-back control of the speed change ratio. CONSTITUTION:If a variation speed by a speed change ratio is higher, the line pressure fed into a hydraulic cylinder 414 for a drive pulley is compensated up to a high value so as to rapidly change the effective radii of drive and driven pulleys 41, 42. Simultaneously, a control gain for the feed-back control of the speed change ratio is compensated to a large value. Meanwhile, if the variation speed by the speed change ratio is low, the line pressure is compensated to a lower value, and simultaneously, the control gain of the feed-back control is compensated to a small value. Thus, the shift speed can be suitable controlled, and the feed-back control of the speed change ratio is allowed to meet the shift speed, thereby it is possible to effectively prevent a delay in control response and hunting of the speed ratio.

Description

DETAILED DESCRIPTION OF THE INVENTION (Field of Industrial Application) The present invention relates to an improvement of a gear ratio control device in a continuously variable transmission.

(Prior Art) Conventionally, as a control device for the gear ratio of a continuously variable transmission, a driving pulley and a driven pulley each having a variable effective radius, a belt wound between the two pulleys, and a hydraulic pressure of the driving pulley are used. A gear ratio adjustment means that adjusts the effective radius of the driving pulley and the driven pulley by supplying and discharging line pressure to and from the cylinder, and continuously and variably adjusts the gear ratio between the two. It is known that the gear ratio adjusting means is feedback-controlled so that the gear ratio becomes a target gear ratio depending on the driving state.

By the way, when controlling the gear ratio between the driving pulley and the driven pulley to the target gear ratio, it is preferable that the gear ratio be configured to be able to quickly change toward the target gear ratio.

Therefore, as disclosed in Japanese Unexamined Patent Publication No. 61-132427, the present applicant has proposed that when the change rate of the gear ratio is large and the gear ratio is to be changed quickly, the oil pressure is supplied to the hydraulic cylinders of the driving pulley and the driven pulley. By increasing the line pressure, the effective radii of the drive pulley and driven pulley can be quickly changed to improve speed change responsiveness, and the tension of the belt between the drive pulley and driven pulley can be controlled to an appropriate tension in a short time. Thus, when the transmission torque increases, such as during downshifting, slippage of the belt is suppressed and damage to the belt is effectively suppressed or prevented.

(Problem to be Solved by the Invention) However, in this case, it has been found that when the control gain of the feedback control of the gear ratio is fixed to a constant value, the following problem arises. In other words, when the line pressure is increased to improve shift responsiveness as described above, the control gain of the feedback control becomes relatively small, and a response delay may occur in the feedback control. On the other hand, when the line pressure is controlled to be low and the shift responsiveness is not improved much, the control gain of the feedback control becomes relatively large, which may cause hunting in the control.

The present invention has been made in view of the above points, and its purpose is to control the line pressure to appropriately improve the shift response, and to reduce the response delay of the feedback control of the gear ratio and the speed change response in the case of i*. The objective is to effectively prevent hunting.

(Means for Solving the Problems) In order to achieve the above object, the present invention adjusts the degree of adjustment of speed change responsiveness by controlling line pressure and the degree of control responsiveness of feedback control of speed ratio. We will respond appropriately.

In other words, the specific solution of the present invention consists of a driving pulley and a driven pulley whose effective radii are changed by supplying and discharging line pressure to and from a hydraulic cylinder, a belt wound between the two pulleys, and the driving pulley. a gear ratio adjusting means for variably adjusting the gear ratio between the drive pulley and the driven pulley by supplying and discharging line pressure to the hydraulic cylinder of the pulley; The present invention is based on a control device for a continuously variable transmission, which is equipped with a control device that performs feedback control of a ratio adjustment device. Further, a change speed detection means for detecting the change speed of the speed change ratio, and a line pressure correction means for correcting the line pressure according to the change speed of the speed change ratio detected by the change speed detection means are provided. Further, gain correction means receives the output of the change speed detection means and corrects the control gain of the feedback control by the control means in accordance with the change speed of the gear ratio in accordance with the line pressure correction by the line pressure correction means. The configuration is such that it is provided.

(Function) With the above configuration, in the present invention, when the change speed of the gear ratio is high and a speed change is to be made quickly, the line pressure supplied to the hydraulic cylinder of the drive pulley is corrected to be high, so that the line pressure supplied to the hydraulic cylinder of the drive pulley and The effective radius of the driven pulley changes quickly, increasing the speed of shifting. Moreover, at this time, the control gain of the feedback control of the gear ratio is also greatly corrected at the same time as the line pressure increase correction. As a result, the feedback control amount also becomes a large value, so that no response delay occurs in the feedback control of the gear ratio at the high shift speed, and good responsiveness can be obtained.

In addition, when the change rate of the gear ratio is low and the gear ratio is to be changed relatively slowly, the line pressure supplied to the hydraulic cylinder of the drive pulley is corrected to be low, and the effective radius of the drive pulley and driven pulley changes. Since the speed becomes gentler, the gear shifting speed does not become very high. At this time, the control gain of the feedback control is also corrected to a small value at the same time, so the effective radius of the driving pulley and the driven pulley does not cause hunting in the vicinity of the effective radius corresponding to the target gear ratio, and the gear ratio becomes the target gear ratio. It will converge smoothly.

(Effects of the Invention) As explained above, according to the continuously variable transmission control device of the present invention, the line pressure supplied to the drive pulley is corrected according to the speed of change of the gear ratio, and the line pressure correction At the same time, the control gain of the feedback control of the gear ratio is also corrected, so the gear shifting speed can be controlled appropriately, and the responsiveness of the gear ratio feedback control can be matched to this shifting speed, resulting in a delay in control response. It is possible to effectively prevent hunting in the gear ratio.

(Example) Embodiments of the present invention will be described below with reference to the drawings.

FIG. 1 shows the overall structure of a continuously variable transmission. The continuously variable transmission shown in the figure includes a torque converter 2 connected to an output shaft 11 of an engine 1, a forward/reverse switching mechanism 3, a continuously variable transmission mechanism 4,
It basically consists of a speed reduction mechanism 5 and a differential mechanism 6.

The torque converter 2 includes a pump cover 21 coupled to the engine output shaft 11, a pump impeller 22 fixed to one side of the pump cover 21 and rotating integrally with the engine output shaft 11, and the pump impeller 22. a turbine runner 23 rotatably provided inside the pump cover 21 so as to face the turbine launch 23;
The stator 24 is interposed between the pump impeller 22 and the pump impeller 22 to increase torque, and the turbine shaft 25 is fixed to the turbine runner 23. Above stake 24
is connected to the transmission case 7 via a one-way clutch 26 and a stator shaft 27. A turbine shaft 2 is provided between the turbine runner 23 and the pump cover 21.
A lock-up piston 28 is slidably attached to the lock-up piston 28, and hydraulic pressure is introduced into and discharged from a lock-up engagement chamber 29g and a lock-up opening chamber 29b formed on both sides of the lock-up piston 28. The piston 28 and the pump cover 21 are connected and opened.

The forward/reverse switching mechanism 3 includes a carrier 31, a pinion gear 32°32 supported by the carrier 31, a sun gear 34 spline-coupled to a primary shaft 411 of a continuously variable transmission mechanism 4 (described later) and meshing with the pinion gear 32, and a pinion gear 32, and the ring gear 35 is spline-coupled to the turbine shaft 25 of the torque converter 2. Further, a forward clutch 36 is provided between the ring gear 35 and the carrier 31 to connect and disconnect them, and a forward clutch 36 is provided between the carrier 31 and the transmission case 7 to selectively fix the carrier 31 to the transmission case 7. A reverse brake 37 is provided. With this configuration, when the forward clutch 36 is engaged and the reverse brake 37 is released, the ring gear 3
5 and the carrier 31 are connected to rotate integrally, and the rotation of the turbine shaft 25 is directly transmitted to the primary shaft 4 of the continuously variable transmission mechanism 4.
On the other hand, when the reverse brake 37 is engaged and the forward clutch 36 is released, the carrier 31 is fixed non-rotatably to the case 7, and the rotation of the ring gear 35 is transmitted to the sun gear 34 via the pinion gear 32... The rotation of the turbine shaft 25 is reversed and transmitted to the primary shaft 411 of the continuously variable transmission mechanism 4. Further, when both the forward clutch 36 and the reverse brake 37 are released, the continuously variable transmission mechanism 4 is connected to the turbine shaft 25.
(in neutral and parking states), the driving force of the engine is no longer transmitted to the primary shaft 411 of the vehicle.

The continuously variable transmission mechanism 4 includes a primary pulley 41 as a driving pulley, a secondary pulley 42 as a driven pulley, and a V-belt 43 wound between these pulleys 41 and 42.

The primary pulley 41 includes a primary shaft 411 disposed coaxially with the turbine shaft 25, a fixed conical plate 412 fixed to the primary shaft 411, and a primary pulley 41 disposed opposite to the fixed conical plate 412. It has a movable conical plate 413 that is slidably supported by *411. When the movable conical plate 413 moves, the clamping position of the V-belt 43 changes, and the effective pitch diameter (effective radius) changes. That is, when the movable conical plate 413 approaches the fixed conical plate 412, the effective pitch diameter increases, and the movable conical plate 413 approaches the fixed conical plate 41.
2, the effective pitch diameter becomes smaller.

Further, the secondary pulley 42 basically has the same configuration as the primary pulley 41 described above.

That is, it has a secondary shaft 421 arranged parallel to the primary shaft 411, a fixed conical plate 422 fixed to the secondary shaft 421, and a movable conical plate 423 slidably supported. The effective pitch diameter is changed.

Each movable conical plate 413 in each of these boos 941.42
, 423 are provided with movable conical plates 413, 4, respectively.
Hydraulic cylinders 414 and 424 for sliding the 23 are provided. The pressure receiving area of the hydraulic cylinder 414 of the primary pulley 41 is set to about twice the pressure receiving area of the hydraulic cylinder 424 of the secondary pulley 42.

Hydraulic pressure is introduced and discharged into the hydraulic cylinder 414 of the primary pulley 41 in order to change the gear ratio between both pulleys 41 and 42, and the hydraulic cylinder 424 of the secondary pulley 42 is used to maintain the appropriate tension of the V-belt 43 at all times. Hydraulic pressure is introduced and discharged to maintain the And the hydraulic cylinder 41 of the primary pulley 41
When hydraulic pressure is introduced into the primary pulley 41, the gripping position of the V belt 43 on the primary pulley 41 moves outward, and the effective pitch diameter of the primary pulley 41 increases.
Along with this, the V belt 4 at the secondary pulley 42
3 moves inward, the effective pitch diameter of the secondary pulley 42 becomes smaller, and the gear ratio between the primary shaft 411 and the secondary shaft 421 changes to a smaller value (in the direction of speed increase). Conversely, when the hydraulic pressure is discharged from the hydraulic cylinder 414, the effective pitch diameter of the primary pulley 41 becomes smaller and the effective pitch diameter of the secondary pulley 42 becomes larger, and the gear ratio between the primary shaft 411 and the secondary shaft 421 becomes larger. (in the direction of deceleration)
It's about to change.

Further, the speed reduction mechanism 5 and the differential mechanism 6 have a known structure, and are adapted to transmit the rotation of the secondary shaft 421 to the axle shaft 61.

Next, the torque converter 2 in the above-mentioned continuously variable transmission
, the forward clutch 36 and reverse brake 37 of the forward/reverse switching mechanism 3, and the primary pulley 41 and secondary pulley 4 of the continuously variable transmission mechanism 4.
The hydraulic circuit that controls each operation of the 2 and 2 will be explained based on FIG. 2.

The hydraulic circuit shown in the figure includes an oil pump 81 driven by an engine l. The hydraulic oil discharged from the oil pump 81 is first adjusted to a predetermined line pressure by the line pressure regulating valve 82 and then supplied to the hydraulic cylinder 424 of the secondary pulley 42 via the line 101. Finally, the hydraulic cylinder 4 of the primary pulley 41 is connected to the hydraulic cylinder 4 of the primary pulley 41 via a line 102 branched from the
14.

The line pressure regulating valve 82 has a spool 8 composed of a main spool 821 and a sub spool 822 arranged in series.
It has 20. A main spool 821 and a sub-spool 822 constituting the spool 820 are connected such that one end of the sub-spool 822 is brought into contact with one end of the main spool 821. At the other end of the sub spool 822,
A large diameter portion 822a is provided which has a larger cross-sectional area than the contact area (cross-sectional area of the connecting portion) with the main spool 821. At a position corresponding to the center of the main spool 821 is a pressure regulating boat 82 to which oil discharged from the oil pump 81 is guided.
3 and a drain boat 824 that communicates with the suction side of the oil pump 81. When the main spool 821 moves to the left side in the figure, a pressure regulating boat 823 and a drain boat 824 are provided.
When the main spool 821 moves to the right side in the figure, the pressure regulating boat 823 and the drain boat 824 are cut off, and when the main spool 821 moves to the right side in the figure, the pressure regulating boat 823 and the drain boat 824 are cut off. 824 are communicated with each other. A first pilot chamber 825 is located at a position corresponding to the connecting portion between the main spool 821 and the sub spool 822.
A spring 826 is interposed in the first pilot chamber 825 to urge the main spool 821 to the left in the figure. In addition, the large diameter portion 822 of the sub spool 822
A second pilot chamber 827 communicating with the first pilot chamber 825 is formed in a. These first pilot chamber 825 and second pilot chamber 827 have a line 10
2, the hydraulic oil is reduced to a predetermined pressure by the reducing valve 83 while passing through the line 103, and is introduced as a pilot pressure adjusted by the first duty solenoid valve 91 while passing through the pilot passage 103a. It has become so. While this pilot pressure acts in the same direction as the biasing force of the spring 826,
The hydraulic pressure in the line 101 acts on the other end of the main spool 821 to counteract the biasing force and pilot pressure, and the spool 820 moves due to the force relationship between the pressure regulating boat 823 and the drain boat 824. By communicating and cutting off the line pressure, the line pressure is controlled to a value corresponding to the pilot pressure regulated by the first duty solenoid valve 91.

A speed ratio control valve 85 is provided in the line 102 . The gear ratio control valve 85 includes a spool 851 and a spring 8 that urges the spool 851 to the right in the figure.
52, a line pressure boat 853 connected to the upstream part of the line 102, a drain boat 854, and a spring 85.
2. A reverse port 855 opened on the installation side and connected to the shift valve 87 via the line 104, and a spring 852.
It has a pilot chamber 856 formed on the opposite side to the installation side and into which pilot pressure is introduced. Pilot room 856
is connected via a pitot valve 86 to a second duty solenoid valve 92 and a pitot pressure generating means 90 that generates a pitot pressure corresponding to the rotation speed of the engine 1. Therefore, the pitot pressure generated by the pitot pressure generating means 90 and the pressure adjusted by the second duty solenoid valve 92 can be selectively introduced into the pilot chamber 856 as pilot pressure by the pitot valve 86,
Even if the second duty solenoid valve 92 fails, the pitot pressure generating means 90 can
Pitot pressure can be introduced as pilot pressure.

When the gear ratio control valve 85 is moving forward (when the shift valve 87 is in any of the shift positions 2.1), hydraulic pressure is drained from the reverse boat 855 through the shift valve 87. Therefore, the spool 851 moves due to the force relationship between the pilot pressure introduced into the pilot chamber 856 and the biasing force of the spring 852, and the line pressure boat 8
53 and the drain boat 854 are selectively communicated with the hydraulic cylinder 414 of the primary pulley 41.

In this way, when moving forward, the pilot chamber 856
Primary pulley 4
A gear ratio adjustment system that variably adjusts the gear ratio between a primary pulley 41 and a secondary pulley 42 of a continuously variable speed mechanism 4 by controlling the supply and discharge of hydraulic pressure to and from a hydraulic cylinder 414 of 1. It constitutes means 100.

In addition, when moving backward (when the shift valve 87 is in the R shift position)
Hydraulic pressure (operating pressure to be described later) is introduced from the reverse port 855, and the spool 851 is fixed in a state where it is pressed to the right side in the figure by this operating pressure. Therefore, when the vehicle moves backward, the hydraulic cylinder 414 of the primary pulley 41 and the drain boat 854 are constantly communicated with each other, and the gear ratio is fixedly maintained at the maximum gear ratio.

In addition, even in neutral and parking (when the shift valve 87 is in each of the N and P shift positions), where the driving force of the engine 1 is not transmitted to the axle 61 by the forward/reverse switching mechanism 3, the same state as in reverse occurs. .

The hydraulic oil whose pressure is regulated by the line pressure regulating valve 82 is
In addition to line 101, it is also sent out to line 105. The hydraulic fluid sent to the line 105 is adjusted to a predetermined operating pressure by the operating pressure regulating valve 88 and then supplied to the lines 106 and 107.

The operating pressure regulating valve 88 has a spool 881 and a spool 88.
a pilot chamber 882 formed on one end side of 1; a spring 883 interposed in this pilot chamber 882;
A 121st pressure boat 884 connected to the line 105,
It has a second pressure regulating boat 885 connected to the line 107 and a drain boat 886. Pilot room 882
is connected to the first duty solenoid valve 91 via the pilot passage 103a. Therefore, hydraulic oil whose pressure is regulated by the first duty solenoid valve 91 is introduced into the pilot chamber 882 as pilot pressure. While this pilot pressure acts in the same direction as the biasing force of the spring 883, the spool 888 acts against the biasing force and the pilot pressure.
Hydraulic pressure in the line 105 acts on the other end of the spool 881, and the spool 881 moves due to the force relationship between the first and second pressure regulating boats 884, 885 and the drain boat 886 to communicate with and disconnect from the drain boat 886. Accordingly, the forward clutch 36
The operating pressure of the reverse brake 37 is controlled to a value corresponding to the pilot pressure regulated by the first duty solenoid valve 91.

The hydraulic oil supplied to the line 106 is transferred to the shift valve 87
line 10 when in shift position 2.1.
9 to the hydraulic chamber 36a of the forward clutch 36 of the forward/reverse switching mechanism 3, and when the shift valve 87 is in the R shift position, the hydraulic pressure of the reverse brake 37 of the forward/reverse switching mechanism 3 is supplied via the line 108. It is supplied to the chamber 37a and also to the reverse boat 855 of the speed ratio control valve 85 via the line 104. On the other hand, when the shift valve 87 is in the R, N, or P shift position, the hydraulic fluid in each hydraulic chamber 35a, 37a of the forward clutch 36 and reverse brake 37 of the forward/reverse switching mechanism 3 flows through the line 109.
, 108. Therefore, the forward clutch 36 and the backward brake 37 of the forward/reverse switching mechanism 3 are engaged and released in accordance with the shift position of the shift valve 87, and as described above, the Rl
At the N and P shift positions, the gear ratio of the continuously variable transmission mechanism 4 is held fixed at the maximum gear ratio.

Further, the hydraulic oil supplied to the line 107 is supplied to the torque converter 2 through the lock-up control valve 89.
The lockup closing chamber 29a or the lockup opening chamber 29b is supplied with the lockup. The lock-up control valve 89 is configured such that the operation of the spool 891 is controlled by the pilot pressure regulated by the third duty solenoid valve 93. Then, when the pilot pressure becomes lower, the spool 891 moves to the right in the figure and connects the line 107 to the lockup fastening chamber 29.
As hydraulic oil is supplied to the lock-up opening chamber 29b, and the pilot pressure increases, the spool 891 moves to the left in the figure, and the hydraulic oil in the lock-up opening chamber 29b begins to drain. The hydraulic oil is supplied to the lockup opening chamber 29b, and the hydraulic oil in the lockup engagement chamber 29a is drained.

Note that 94 indicates that the first duty solenoid valve 91 is O.
An accumulator valve is used to prevent the pilot pressure in the pilot passage 103a from pulsating when the N/OFF state is turned off, 95 and 96 are accumulators that cushion the shock when the forward clutch 36 and reverse brake 37 are engaged, respectively, and 97 is a relief. It's a valve. Further, 98 is a pressure holding valve that maintains a predetermined low constant pressure when draining the pressure oil in the hydraulic cylinder 414 of the primary pulley 41.

FIG. 3 shows the electric control circuit of the above-mentioned continuously variable transmission. In this figure, a control unit 110 containing a microcomputer etc. receives a shift position signal from a shift position sensor 111 that detects a shift position (D, 1.2°R, N, P) operated by a driver;
The primary pulley rotation speed signal from the primary rotation speed sensor 112 that detects the rotation speed np of the primary shaft 411, the secondary pulley rotation speed signal from the secondary rotation speed sensor 113 that detects the rotation speed ns of the secondary shaft 421, and the engine. The throttle valve opening signal from the throttle opening sensor 114 that detects the throttle valve opening TVO of the engine 1, the engine rotation speed signal from the engine rotation speed sensor 115 that detects the rotation speed Ne of the engine 1, and the engine rotation speed signal of the torque converter 2. A turbine rotation speed signal from a turbine rotation speed sensor 116 that detects the rotation speed Nt of the turbine shaft 25 is input.

The control unit 110 performs duty control on the first to third duty solenoid valves 91, 92, and 93 based on these input signals, thereby increasing the line pressure yJ3 valve regulator 82 and operating pressure: J3fi! Each pilot pressure introduced into the valve 88, the gear ratio control valve 85, and the rotor up control valve 89 is adjusted.

Then, the control unit 110 calculates a target primary rotation speed corresponding to a target gear ratio of the continuously variable transmission according to the operating state based on the input throttle valve opening degree and secondary rotation speed from the map shown in FIG. At the same time,
The deviation between this target primary rotation speed and the actual primary rotation speed detected by the primary rotation speed sensor 112 is calculated, and the integral term, proportional term, and differential term of feedback control are calculated based on this rotation speed deviation and the set control gain. By calculating the total value as a feedback control amount and controlling the duty of the second duty solenoid valve 92 for the gear ratio control valve 85 of the gear ratio:A difference means 100, the actual primary rotation speed is set as the target. A control means 150 is configured to adjust the speed ratio to the target speed ratio by adjusting the rotation speed to the primary rotation speed.

Next, line pressure control by the control unit 110 will be explained based on the control flow shown in FIG. 4.

After starting, at step SA+, the engine speed Ne,
After reading the primary rotational speed np1 secondary rotational speed ns and throttle valve opening TVO signals, step S
In step A2, the input torque T1n acting on the primary pulley 41 is calculated based on the engine speed Ne and throttle valve opening TVO read above, and in step SA3, the target gear ratio R is calculated as the secondary engine speed ns and throttle valve opening. Calculate based on TVO. And step SA4
Based on the above-calculated human torque Tin and target gear ratio R, the necessary pressing force to be applied from the secondary pulley 42 to the belt 43 is calculated, and the necessary pressing force against the belt calculated above in step SA5 is necessary to obtain the force. The target line pressure Po is calculated based on the pressure receiving area of the hydraulic cylinder 424 of the secondary pulley 42, etc.

After that, in step SA6, the change speed ΔR of the gear ratio is set to the previous and current target primary rotation speeds np.

(1) , n po (1-1) and secondary rotation speed n
After calculating 7 using the formula % formula % based on s, the correction coefficient k of the target line pressure is
Set p to a large value greater than or equal to "1" and proceed to step SA.
7, the target line pressure Po is multiplied and corrected by this correction coefficient kp, and this corrected target line pressure Po (Po-Po
The process ends by controlling the first Denitey solenoid valve 91 so that txkp).

Further, FIG. 5 shows a control gain correction flow for feedback control of the gear ratio by the gear ratio control valve 85. In the same figure, after starting, at step SBI, the engine speed Ne, the throttle valve opening TVO, the primary rotation speed n
After reading the signals of the throttle valve opening TVO and the secondary rotation speed ns, the primary pulley 41 is adjusted in step SB2 based on the read throttle valve opening TVO and the secondary rotation speed ns.
At step 5Bff, the speed change rate ΔR of the gear ratio is calculated based on the same calculation formula as in step S6 of the target line pressure setting flow in FIG. 4.

Then, in step SB4, the larger the change speed ΔR of the gear ratio calculated above is, and the faster the gear shift is attempted, the feedback control gain correction amount kP is set to a larger value of "1" or more, and in step S8, The control gain kO of the feedback control that has been set is adjusted by the above correction E1k.
Multiplication correction is performed by F, and this corrected control gain k (k−k
oXkF) to calculate the integral term, proportional term, and differential term of the feedback control, and use the total value as the feedback control amount to feedback control the second duty solenoid valve 92 for the gear ratio control valve 85. finish.

Therefore, step S of the target line pressure setting flow in FIG.
A6 and step SR2+s of the correction flow in FIG.
s3 constitutes a change rate detection means 151 that detects the change rate ΔR of the gear ratio, and steps SA2 to SA7 of the target line pressure setting flow in FIG.
The line pressure correction means 152 is adapted to correct the line pressure higher according to the change speed ΔR of the speed change ratio detected by the change speed detection means 151, the larger the change speed ΔR of the speed change ratio is and the faster the shift is attempted. It consists of

Also, step S of the correction flow in FIG.

4 and SBS, upon receiving the output of the rate of change detection means 151, the control gain of the feedback control by the control means 150 is adjusted according to the rate of change ΔR of the gear ratio in accordance with the correction of the line pressure by the line pressure correction means 152. ,
The gain correction means 153 is configured to correct the change speed ΔR of the gear ratio to a larger value as the speed of change ΔR increases and the speed is changed faster.

Therefore, in the above embodiment, when the change rate of the gear ratio is large and the gear ratio is to be shifted quickly, such as during a downshift when the throttle valve opening is wide open, the line pressure is corrected as the change rate increases. The coefficient kp is “
1” value, the target line pressure p
o is multiplied by this correction coefficient kp. As a result, when the line pressure regulating valve 82 is actuated and controlled by the first duty solenoid valve 91 so as to reach the increased target line pressure pO, the actual line pressure increases as the speed of change in the gear ratio increases. As the torque increases, the tension of the belt 43 of the continuously variable transmission mechanism 4 also quickly increases to a large value, and slipping of the belt 43 when the transmitted torque increases is effectively suppressed or prevented.

Also, as mentioned above, since the line pressure is high and corrected,
Even if the control duty rate of the second duty solenoid valve 92 for speed ratio control is the same value, the hydraulic cylinder 41 of the primary pulley 41 is adjusted as much as the actual line pressure is higher.
The operating pressure acting on 4 also increases. As a result, the effective radii of the primary pulley 41 and the secondary pulley 4242 change over time, resulting in a higher gear shift speed, which can better cope with situations where fast gear shifts are required.

In that case, regarding the feedback control of the gear ratio, that is, the duty control of the second duty solenoid valve 92 for gear ratio control, the correction coefficient kF for the control gain increases to a value of "1" as the speed of change of the gear ratio increases. Since the control gain kO of feedback control set in advance is multiplied by this correction coefficient kF, the feedback control amount is multiplied by the correction coefficient kF, and the second
Since the control duty rate of the duty solenoid valve 92 becomes a correspondingly large value, no response delay occurs in the feedback control of the gear ratio.

As a result, the operating pressure adjusted by the speed ratio control valve 85 increases more quickly, and the effective radius of the primary pulley 41 and the secondary pulley 42 changes in a shorter time, resulting in good speed change responsiveness.

In the above embodiment, the line pressure correction coefficient kp was set to a value of "1" or more according to the speed of change of the gear ratio, and the line pressure was corrected only to the pressure increasing side, but the speed of change of the gear ratio was small. When a slow shift is performed, set the correction coefficient kp to "1".
The line pressure may be reduced by setting it to a value less than the above value. In this case, the control gain of the feedback control of the gear ratio is also corrected to a small value at the same time. As a result, the feedback correction amount is not too large and is an appropriate value, so that the gear ratio can be satisfactorily converged to the target gear ratio without hunting.

[Brief explanation of drawings]

The drawings show an embodiment of the present invention, in which Fig. 1 is an overall configuration diagram of a continuously variable transmission, Fig. 2 is a hydraulic control circuit diagram, Fig. 3 is a block diagram of an electrical control system, and Fig. 4 is a line pressure correction diagram. FIG. 5 is a flowchart showing correction of control gain of gear ratio feedback control, and FIG. 6 is a diagram showing target primary rotation speed characteristics. 41...Primary pulley (drive pulley), 414...
...Hydraulic cylinder, 42...Secondary pulley (driven pulley), 43...Belt, 82...Line pressure adjustment valve, 85...Speed ratio control valve, 100...Speed ratio: J
! J adjustment means 50...control means, 151...change speed detection means, 152...line pressure correction means, 153
...gain correction means. Figure 5

Claims (1)

[Claims] (1) A driving pulley and a driven pulley whose effective radii are changed by supplying and discharging line pressure to and from the hydraulic cylinder, a belt wound between the two pulleys, and a line pressure supplied to the hydraulic cylinder of the driving pulley. a gear ratio adjusting means for variably adjusting the gear ratio between the driving pulley and the driven pulley, and a control means for feedback controlling the gear ratio adjusting means so that the gear ratio becomes a target gear ratio according to the operating state. A continuously variable transmission comprising: a change speed detection means for detecting a change speed of the speed ratio; and a line pressure correction means for correcting line pressure according to the change speed of the speed ratio detected by the change speed detection means. , gain correction means for receiving the output of the change speed detection means and correcting the control gain of the feedback control by the control means in accordance with the change speed of the gear ratio in accordance with the line pressure correction by the line pressure correction means. A control device for a continuously variable transmission characterized by: