JPH0464759A - Control device for continuously variable transmission - Google Patents

Abstract

PURPOSE:To improve a responsiveness for a shift-up from a maximum gear ratio upon acceleration during starting by providing a pressure increasing means for increasing the pressure of hydraulic oil fed to a drive pulley when the gear ratio is maximum. CONSTITUTION:A pressure increasing means increases the pressure of hydraulic oil fed to a drive pulley 41, when the gear ratio is maximum, so as to prevent occurrence of a slip between a primary pulley 41 and a belt 43 at the time when the primary rotational speed reaches a rotational speed at which a shift is initiated upon acceleration during starting. Further, the pressure of hydraulic oil fed to the drive pulley is controlled in such a way that the ratio between a pressing force Fp applied to the belt 43 from the drive pulley 41 and a pressing force Fs applied to the belt 43 from a driven pulley 42 increases corresponding to an increase in input torque applied to the drive pulley, thereby it is possible to prevent occurrence of a slip between the primary pulley 41 and the belt 43 at the time when the primary rotational speed reaches a rotational speed at which the shifting is initiated upon acceleration during starting.

Description

[Detailed description of the invention] (Industrial application field) The present invention relates to a control device for a continuously variable transmission.

(Prior art) Conventional control devices for continuously variable transmissions include belt tension control means that controls the belt tension by controlling the supply and discharge of hydraulic pressure to the hydraulic chamber of the driven pulley, and a belt tension control means that controls the belt tension by controlling the supply and discharge of hydraulic pressure to the hydraulic chamber of the driven pulley. 2. Description of the Related Art A control device for a continuously variable transmission is known, which is equipped with a gear ratio control means for controlling exhaust and a gear ratio.

Shift control by such a control device is performed in the following steps. That is, from the throttle valve opening degree Tvo and the rotational speed Ns of the secondary pulley as a driven pulley, the target value of the gear ratio γ of the continuously variable transmission according to the operating state is determined based on the map illustrated in FIG. Target rotation speed N of the primary pulley as a drive pulley corresponding to the target gear ratio γ1
p+ is calculated, and the gear ratio control means controls the supply and discharge of hydraulic pressure to the hydraulic chamber of the primary pulley to adjust the actual primary rotation speed Np to the target primary rotation speed Np, and the input torque to the primary pulley and Based on the target speed ratio γ1, the belt tension control means controls supply and discharge of hydraulic pressure to the hydraulic chamber of the secondary pulley to control the belt tension to an optimum value.

In this control, at the time of starting acceleration, the primary rotation speed is increased to the rotation speed corresponding to the throttle valve opening along line 11 of the map in FIG. 7, that is, the line of the maximum gear ratio γmax, and the primary rotation speed Np When the rotational speed corresponding to the valve opening degree Tvo, that is, the speed change start rotational speed N+)2 is reached, the speed change control from the maximum speed ratio γmax to the target speed ratio γ1 is started. The increase in primary rotational speed along the line of this maximum gear ratio γmax is
This is achieved by discharging the hydraulic oil from the hydraulic chamber of the primary pulley, minimizing the rotation radius of the belt passed over the primary pulley, setting the speed ratio to the maximum speed ratio γll1aX, and maintaining this state. Then, when the primary rotational speed Np or the shift start rotational speed NJ)2 is reached, the supply of hydraulic oil to the hydraulic chamber of the primary pulley is started, and along with this, the rotation radius of the belt that is passed over the primary pulley. increases, and the gear ratio changes toward the target gear ratio γ1.

(Problem to be solved by the invention) The belt pressing force applied to the belt via the secondary pulley by the belt tension control means (hereinafter referred to as secondary clamping force) Fs is determined by the input torque to the primary pulley and the target gear ratio. , the larger the input torque and the larger the target gear ratio. This is because the larger the input torque, the greater the frictional force between the belt and pulley, and the larger the target gear ratio, the greater the torque increasing effect. This is due to the need to increase the frictional force between them. On the other hand, when the input torque increases at a predetermined gear ratio, the belt pressing applied to the belt via the primary pulley by the gear ratio control means increases the force (hereinafter referred to as primary clamping force) Fp in proportion to the secondary clamping force Fs. By increasing the rotation radius, the primary rotation radius corresponding to the predetermined speed ratio, that is, the rotation radius of the belt on the primary pulley can be maintained. However, due to the occurrence of slip between the primary pulley and the belt, even if the primary rotation radius is maintained at a value corresponding to the predetermined gear ratio, the actual gear ratio is not maintained at the predetermined value and is lower than this. becomes a large value. In order to avoid such a situation and maintain a predetermined gear ratio, it is necessary to set the ratio Fp/Fs between the primary clamping force Fp and the secondary clamping force Fs to an increasing function of input torque T)1(T).

However, conventionally, when accelerating from a start, hydraulic oil is drained from the hydraulic chamber of the primary pulley until the primary rotation speed Np or the shift start rotation speed Np+ is reached, and the radius of rotation of the belt passed through the primary pulley is maintained at the minimum. Because I was
During this time, the primary clamping force Fp is maintained at a constant value until the primary rotational speed Np reaches the shift start rotational speed Np+, and the secondary clamping force Fs is increased in response to an increase in throttle opening and, ultimately, an increase in input torque. , the ratio Fp/Fs of the primary clamping force Fp and the secondary clamping force Fs was smaller than the required value 1 ((T). Therefore, slippage occurred between the primary pulley and the belt, and the primary rotation speed When reaches the shift start rotation speed Np, the actual gear ratio is even larger than the maximum gear ratio, and from this point on, even if hydraulic oil is supplied to the hydraulic chamber of the primary pulley, the primary clamping force Fp Ratio Fp to secondary clamping force Fs
There is a problem in that the shift is not started until /Fs reaches the required value H(T), that is, the responsiveness of upshifting from the maximum gear ratio at the time of starting acceleration deteriorates.

SUMMARY OF THE INVENTION Therefore, an object of the present invention is to provide a control device for a continuously variable transmission that can improve the responsiveness of upshifting from the maximum gear ratio during start acceleration.

(Means for Solving the Problems) In order to solve the above problems, the present invention provides belt tension adjustment that controls the belt tension by controlling the supply and discharge of hydraulic pressure to the hydraulic chamber of the driven pulley of the continuously variable transmission. and a gear ratio adjusting means for controlling supply and discharge of hydraulic pressure to and from a hydraulic chamber of a drive pulley to control a gear ratio, in which the gear ratio is set to a maximum gear ratio. A control device for a continuously variable transmission is provided, characterized in that it is equipped with a pressure increasing means for increasing the pressure of oil pressure supplied to a drive pulley.

In a preferred embodiment of the present invention, the pressure increasing means increases the input torque to the driving pulley such that the ratio of the pressing force Fp applied to the belt from the driving pulley to the pressing force Fs applied to the belt from the driven pulley increases. This is a means for controlling the amount to increase in accordance with the amount.

(Function) According to the above configuration of the present invention, the pressure increasing means increases the pressure of the oil pressure supplied to the drive pulley when the gear ratio is the maximum gear ratio, and the primary rotational speed reaches the shift start rotational speed at the time of starting acceleration. This prevents slippage between the primary pulley and the belt when the Further, the pressure increasing means is configured such that the ratio of the pressing force Fp applied to the belt from the driving pulley and the pressing force Fs applied to the belt from the driven pulley increases in response to an increase in the input torque to the driving pulley. The hydraulic pressure supplied to the drive pulley is increased and controlled to prevent slippage between the primary pulley and the belt when the primary rotational speed reaches the shift start rotational speed during start acceleration.

(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 an actuation 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 runner 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.

The stator 24 is connected to the transmission case 7 via a one-way clutch 26 and a stator shaft 27. A lock-up piston 28 slidably attached to the turbine shaft 25 is provided between the turbine runner 23 and the pump cover 21, and lock-up fastening chambers 2 are formed on both sides of the lock-up piston 28.
The lock-up piston 2B and the pump cover 21 are connected and opened by introducing and discharging hydraulic pressure into the lock-up opening chamber 29a and the lock-up opening chamber 29b.

The longitudinal switching mechanism 3 includes a carrier 31, a pinion gear 32, 32 supported by the carrier 3I, and a sun gear 34 spline-coupled to a primary shaft 411 of the continuously variable transmission mechanism 4, which will be described later, and meshing with the pinion gear 32.
A ring gear 35 that meshes with the pinion gear 32,
The ring gear 35 is connected to the turbine shaft 2 of the torque converter 2.
5 with spline connection. 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 35
and the carrier 31 are connected to rotate integrally, and the turbine shaft 2
The primary shaft 41 of the continuously variable transmission mechanism 4 keeps the rotation of 5 as it is.
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... tell the
The rotation of the turbine shaft 25 is reversed and transmitted to the primary shaft 411 of the continuously variable transmission mechanism 4. Furthermore, when both the forward clutch 36 and the reverse brake 37 are released, the driving force of the engine is not transmitted from the turbine shaft 25 to the primary shaft 411 of the continuously variable transmission mechanism 4 (neutral and parking states).

Further, 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 fixed conical plate 412 disposed opposite to the fixed conical plate 412 and attached to the primary shaft 411. It has a movable conical plate 413 that is slidably supported. When the movable conical plate 413 moves, the clamping position of the V-belt 43 changes, and the effective pitch (effective radius) changes. That is, the movable conical plate 41
When movable conical plate 413 moves away from fixed conical plate 412, the effective pitch diameter becomes large.

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 pulleys 41, 42
．． 423, each movable conical plate 413.4
Hydraulic cylinders 414, 424 for sliding 23 are provided. The pressure receiving area of the hydraulic cylinder 414 of the primary pulley 41 is set to about two degrees of 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 always maintains the appropriate tension of the V-belt 43. 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 secondary pulley 42, the clamping position of the V-belt 43 on the primary pulley 41 moves to the outside, and the effective pitch diameter of the primary pulley 41 increases.
The clamping position 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 thicker. Large gear ratio (in the direction of deceleration)
It's about to change. Note that the transmission of this embodiment is configured such that the speed ratio when the hydraulic pressure is completely discharged from the hydraulic cylinder 414, that is, the maximum speed ratio is 2.47.

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 the engine 1. 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. The main spool 82 and the sub spool 822 constituting the spool 820 are connected so that one end of the sub spool 822 is in 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 port 82 through which oil discharged from the oil pump 81 is introduced.
3 and a drain port 824 communicating 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 port 823 and a drain port 824 are provided.
When the main spool 821 moves to the right side in the figure, the pressure regulating port 823 and the drain port 824 are communicated with each other. A first pilot chamber 825 is formed at a position corresponding to the connecting portion between the main spool 821 and the sub-spool 822, and a spring 826 that biases the main spool 821 to the left in the figure is interposed in the first pilot chamber 825. has been done. Further, a second pilot chamber 827 communicating with the first pilot chamber 825 is formed in the large diameter portion 822a of the sub spool 822.

These first pilot chamber 825 and second pilot chamber 8
27, after branching from line 102, line 103
The hydraulic oil is reduced in pressure to a predetermined pressure by the reducing valve 83 while passing through the pilot passage 103a, and is introduced as a pilot pressure adjusted by the first duty solenoid valve 19. and,
While this pilot pressure acts in the same direction as the biasing force of the spring 826, the line 101 is connected to the other end of the main spool 821 to counteract the biasing force and the pilot pressure.
The hydraulic pressure inside acts, and due to these force relationships, the spool 8
20 moves to open the pressure regulating port 823 and drain port 824.
By communicating and cutting off between the
It is controlled to a value according to the pilot pressure regulated by the duty solenoid valve 91.

The line pressure adjustment valve 82 constitutes a belt tension adjustment means 200 that adjusts the tension of the belt 43 of the continuously variable transmission mechanism 4 by supplying the regulated line pressure to the hydraulic cylinder 424 of the secondary pulley 42 .

A gear 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 port 853 connected to the upstream part of the line 102, a drain port 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 rotational 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 introduced into the selective pilot chamber 856 by the pitot valve 86 as pilot pressure, and in the unlikely event that Even when the second duty solenoid valve 92 fails, pitot pressure can be introduced from the pitot pressure generating means 90 into the pilot chamber 856 as pilot pressure.

In the above speed ratio control valve 85, when moving forward (
When the shift valve 87 is in either shift position D or 2.1), hydraulic pressure is drained from the reverse port 855 through the shift valve 87, so the pilot pressure and spring introduced into the pilot chamber 856 are The spool 851 moves due to the force relationship with the biasing force of the spool 852, so that the line pressure port 853 and the tren port 854 are selectively communicated with the hydraulic cylinder 414 of the primary pulley 41. In this way, when moving forward, the gear ratio control valve 8
5 is a hydraulic cylinder 414 of the primary pulley 41 in response to the pilot pressure introduced into the pilot chamber 856.
By controlling the supply and discharge of hydraulic pressure to the continuously variable transmission mechanism 4,
Gear ratio adjusting means 1 configured to variably modulate the gear ratio between the primary pulley 41 and the secondary pulley 42.
Configure 00.

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 traveling in reverse, the hydraulic cylinder 414 of the primary pulley 41 and the drain port 854 are always communicated with each other, and the gear ratio γ is fixedly maintained at the maximum gear ratio γmax=2.47.

In addition, even in neutral and parking (when the shift valve 87 is in the N and P shift positions) where the forward/reverse switching mechanism 3 does not transmit the driving force of the engine 1 to the axle 61, the solenoid valve 92 operates to shift the primary The hydraulic cylinder 414 of the pulley 41 and the drain port 854 are always in communication, and the gear ratio γ is the maximum gear ratio γmax=2.
It is held fixed in the state of 47.

The operating pressure 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;
It has a first pressure regulation port 884 connected to line 105, a second pressure regulation port 885 connected to line 107, and a drain port 886. Pilot chamber 882 is connected to first duty solenoid valve 91 via pilot passage 103a. Therefore, the pilot chamber 882 includes the first duty solenoid valve 9.
The hydraulic oil whose pressure was regulated in step 1 is introduced as pilot pressure. This pilot pressure acts in the same direction as the biasing force of the spring 883, while the spool 888 acts in the same direction as the biasing force of the spring 883.
The 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 spools.
By communicating and disconnecting between the pressure regulating ports 884 and 885 and the drain port 886, the operating pressure of the forward clutch 36 and the reverse brake 37 becomes the pilot pressure that is regulated by the first duty solenoid valve 91. The value is controlled accordingly.

The hydraulic oil supplied to the line 106 is transferred to the shift valve 87
When the gear is in the shift position of 2, l, line 10
9 to the hydraulic chamber 36'a of the forward clutch 36 of the forward/reverse switching mechanism 3, and is discharged from the hydraulic chamber 37a of the reverse brake via line 108.
04 to the reverse port 855 of the gear ratio control valve 85
When the shift valve 87 is in the R shift position, it is discharged from the hydraulic chamber 36a of the forward clutch 36 via line 109, and is discharged from the hydraulic chamber 36a of the reverse brake 37 of the forward/reverse switching mechanism 3 via line 108. 37a and is also supplied to the reverse port 855 of the gear ratio control valve 85 via the line 104. On the other hand, the hydraulic fluid in each hydraulic chamber 36a, 37a of the forward clutch 36 and reverse brake 37 of the forward/reverse switching mechanism 3 is
Line 1 when the shift valve 87 is in the N or P shift position
09, it is designed to be exhausted through IO2. Therefore, the forward clutch 36 and the backward brake 37 of the forward/reverse switching mechanism 3 are engaged and released according to the shift position of the shift valve 87, and as described above, the forward clutch 36 and the reverse brake 37 are engaged and released in accordance with the shift position of the shift valve 87, and as described above, the forward clutch 36 and the reverse brake 37 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 low, the spool 891 moves to the right side in the figure and connects the line 107 to the lock-up 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.

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 the shift position (Dll, 2, R, N, P) operated by the driver, and a primary shaft. The primary pulley rotation speed signal from the primary rotation speed sensor 112 that detects the rotation speed of the secondary shaft 411, the secondary pulley rotation speed signal from the secondary rotation speed sensor 113 that detects the rotation speed of the secondary shaft 421, and the throttle valve opening of the engine 1. a throttle valve opening signal from a throttle opening sensor 114 that detects the rotational speed of the engine 1, an engine rotational speed signal from an engine rotational speed sensor 115 that detects the rotational speed of the engine 1, and a turbine shaft 2 of the torque converter 2.
A turbine rotation speed signal from a turbine rotation speed sensor 116 that detects the rotation speed of the engine 5, and a primary oil pressure signal from a primary oil pressure sensor 117 that detects the oil pressure of the primary pulley are 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 controlling the line pressure regulating valve 82, the operating pressure regulating valve 88, and the gear ratio. control valve 8
5 and the lock-up control valve 89 are 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 second duty solenoid valve 92 for the gear ratio control valve 85 of the gear ratio adjusting means 100, the actual primary rotation speed is changed to the target primary rotation. The speed change control means 150 is configured to adjust the speed change ratio to the target speed change ratio by adjusting the speed change ratio.

Next, the shift control during start acceleration by the control unit 110 will be explained based on the flowchart of FIG. 4. Note that in the following explanation, S indicates a control step.

Shift control is started on condition that the engine is ON and the shift valve 87 is at the shift position D12.1.

First, the engine rotation speed Ne, primary rotation speed Np, secondary rotation speed NS, and throttle valve opening degree T■0 are read (Sl), and act on the primary pulley 41 based on the engine rotation speed Ne and throttle valve opening degree TVO. Calculate the input torque Tin (S2), calculate the necessary pressing force to be applied from the secondary pulley 42 to the belt 43 based on the calculated input torque Tin and the maximum gear ratio γmaX, that is, calculate the secondary clamping force Fs (S3), and calculate the secondary clamping force. Line pressure Po required to obtain Fs
is calculated based on the pressure receiving area of the hydraulic cylinder 424 of the secondary pulley 42 (S4), and the first duty solenoid valve 91 is controlled so that the line pressure becomes Po (S4).
S5). Next, based on the throttle valve opening TVO, the sixth
From the control map shown in the figure, find the primary rotational speed Npo at which the gear shift starts, that is, the primary rotational speed Npo at which the gear ratio starts to decrease from the maximum gear ratio γmax, and then calculate the actual primary rotational speed Npo.
It is determined whether p has reached Npo (S6).

At the beginning of acceleration, the primary rotation speed Nl) has not reached Npo, so next the second duty solenoid valve 9
2 is duty controlled to gradually supply hydraulic oil to the hydraulic cylinder 414 of the primary pulley 41 (S7). As a result, the effective pitch of the primary pulley 41 increases, and the actual gear ratio γ is still the maximum gear ratio γmax = 2.4.
Although it is larger than 7, it decreases and the maximum gear ratio γmaX
approach. Next, the actual gear ratio γ (γ=Np2/Ns
), and determine whether the actual gear ratio γ is less than γmax = 2.47 (S8), and determine whether the actual gear ratio γ is less than γm
If it is not less than ax = 2.47, read the primary rotation speed Np again (S9) and then proceed to step S6.
Return to step 86 and repeat steps S9. If the actual gear ratio γ is less than γmax = 2.47, the second duty solenoid valve 92 is duty-controlled to gradually drain the hydraulic oil from the hydraulic cylinder 414 of the primary pulley 4. S10). As a result, the effective pitch of the primary pulley 41 decreases, and the actual gear ratio γ increases and approaches the maximum gear ratio γmax=2.47.

Next, it is determined again whether the actual gear ratio γ is less than γmax = 2.47 (S8), and based on the determination result, the process proceeds to step S6 or S9 as described above. By repeating steps 86 to S9, the primary rotation Np increases while the gear ratio γ is maintained near the maximum gear ratio γmax = 2.47, and the shift start primary rotation speed Np
approach o. Then, when the primary rotation Np reaches the shift start primary rotation speed Npo, the maximum gear ratio γ
Normal speed change control based on the speed change map shown in FIG. 6 is started from the maximum speed toward the target speed change ratio. As mentioned above, this normal speed change control sets the target primary rotation speed corresponding to the target gear ratio of the continuously variable transmission according to the operating condition based on the input throttle valve opening and the secondary rotation speed as shown in FIG. In addition to calculating from the map, 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 for feedback control is calculated based on this rotation speed deviation and the set control gain. , a proportional term, and a differential term, and use the total value as a feedback control amount for the second duty solenoid valve 92 for the gear ratio control valve 85 of the gear ratio adjusting means 100.
This is done by adjusting the actual primary rotation speed to the target primary rotation speed by controlling the duty. Since the actual gear ratio is maintained near the maximum gear ratio γmax = 2.47 at the start of the normal gear change control, the gear change is actually performed at the same time as the start of the gear change control. Therefore, the responsiveness of shifting up from the maximum gear ratio during start acceleration is improved.

Another example of shift control by the control unit 110 will be described based on the flowchart of FIG. 5. Note that in the following explanation, R indicates a control step.

Shift control is started on the condition that the engine is ON, the shift valve 87 is open, and the shift position is 2.1.

First, the engine speed Ne, primary speed Np, secondary speed NS, and throttle valve opening TVO are read (R1), and the input torque acting on the primary pulley 41 is determined based on the engine speed Ne and throttle valve opening TVO. Calculate Tin (R2), calculate the secondary clamping force Fs based on the calculated input torque Tin and maximum gear ratio γmax (R3), and calculate the line pressure Po required to obtain the secondary clamping force Fs of the secondary pulleys 4 and 2. It is calculated based on the pressure receiving area of the hydraulic cylinder 424 (R4), and the first duty solenoid valve 91 is controlled so that the line pressure becomes Po (R5). Then,
The actual gear ratio γ is calculated from the primary rotation speed Np and the secondary rotation speed Ns, and the actual gear ratio γ is the maximum gear ratio γm.
It is determined whether the actual gear ratio γ is larger than aχ-2.47 (R6), and if the actual gear ratio γ is less than the maximum gear ratio γ [1IaX, normal gear shift control based on the gear shift map shown in FIG. 6 is possible. , so the shift control is started. If the actual gear ratio γ is larger than the maximum gear ratio γmax = 2.47, the actual gear ratio γ is set to the maximum gear ratio γmax
Primary clamping force F required to make = 2.47
p is determined by the following formula (R7).

Fp = H(Tin) It is a given, and it is determined experimentally. Function H(Tin
) is an increasing function with respect to the input torque to the primary pulley.

The fact that the actual gear ratio γ is larger than the maximum gear ratio γmax = 2.47 means that the actual primary clamping force F
Since ly' means less than the required primary clamp force Fp given by the above formula, next, the second duty solenoid knob 92 is duty-controlled,
Hydraulic oil is gradually supplied to the hydraulic cylinder 414 of the primary pulley 41 to increase the hydraulic pressure (R8). As a result, the actual primary clamping force Fp' increases and approaches the required primary clamping force Fp. Next, the primary hydraulic pressure pp is detected (R9), the actual primary clamping force Fp' is calculated based on the hydraulic pressure pp and the pressure receiving area of the hydraulic cylinder 414 of the primary pulley 41, and the required primary clamping force Fp and the actual Difference between primary clamping force Fp' and required primary clamping force F
Determine whether the ratio with p is less than a predetermined value of 8 (
RIO).

If the result of the determination is negative, the process returns to step R8 and the actual primary clamping force Fp' is further increased. The absolute value of the ratio of the difference between the primary clamping force Fp required in step RIO and the actual primary clamping force Fp' to the required primary clamping force Fp is taken as follows.
This is to deal with the case where the actual primary clamping force Fp' becomes larger than the necessary primary clamping force Fp due to the supply of hydraulic oil to the hydraulic cylinder 414 of the primary pulley 41. In this case, the stem tube R8 Then, the hydraulic oil is discharged from the hydraulic cylinder 414 of the primary pulley 41 to reduce the hydraulic pressure. If the ratio of the difference between the required primary clamping force Fl) and the actual primary clamping force FIG' and the required primary clamping force Fp becomes less than the predetermined value 8 by repeating steps R8 to RIO, the actual gear ratio γ is the maximum gear ratio γm
Since ax has already reached a value close to 2.47 and the shift control based on the shift map shown in FIG. 6 is already possible, the shift control is started. In this way, the actual gear ratio γ at the start of the gear change control is the maximum gear ratio γmax = 2.4
Since the value is close to 7, the shift is performed at the same time as the start of the shift control, and the responsiveness of shifting up from the maximum gear ratio during start acceleration is improved.

(Effects) As can be seen from the above description, in the present invention, the pressure increasing means increases the pressure of the oil pressure supplied to the drive pulley when the gear ratio is the maximum gear ratio, and when accelerating from a start, the primary rotational speed increases to the gear shift start rotation. To prevent the gear ratio from increasing from the maximum gear ratio due to slip between the primary pulley and the belt when the number reaches the maximum gear ratio.

Further, the pressure increasing means is configured such that the ratio of the pressing force Fp applied to the belt from the driving pulley and the pressing force Fs applied to the belt from the driven pulley increases in response to an increase in the input torque to the driving pulley. The hydraulic pressure supplied to the drive pulley is increased and controlled to prevent the gear ratio from increasing from the maximum gear ratio due to slip between the primary pulley and the belt when the primary rotational speed reaches the shift start rotational speed during start-up acceleration. .

As a result, the shift is started immediately when the primary rotational speed reaches the shift start rotational speed, and the responsiveness of shifting up from the maximum gear ratio at the time of starting acceleration is improved.

[Brief explanation of the drawing]

FIG. 1 is an overall configuration diagram of a continuously variable transmission, FIG. 2 is a hydraulic control circuit diagram of a control device according to an embodiment of the present invention, and FIG. 3 is a diagram of an electrical control system of a control device according to an embodiment of the present invention. Block diagram, FIG. 4 is a flowchart of an example of speed change control by the control device according to the embodiment of the present invention, FIG. 5 is a flowchart of another example of speed change control by the control device according to the embodiment of the present invention, and FIG. 6 is a target An example of a control map for setting the primary rotation, FIG. 7 is a diagram showing a conventional speed change control procedure for a continuously variable transmission. 41...Primary pulley (drive pulley), 42...
- Secondary pulley (driven pulley), 43... Belt, 414.424... Hydraulic cylinder, 82... Line pressure adjustment valve, 85... Gear ratio adjustment valve, 100... Gear ratio adjustment means, 150... Control means, 200... Belt tension adjustment means. S L taste rm! ! '1 + n r and twice four sources 22 post concave! ! 'rXen

Claims (2)

[Claims] (1) A belt tension adjustment means that controls the belt tension by controlling the supply and discharge of hydraulic pressure to the hydraulic chamber of the driven pulley of the continuously variable transmission, and a belt tension adjustment means that controls the supply and discharge of hydraulic pressure to the hydraulic chamber of the drive pulley to change speed. A control device for a continuously variable transmission equipped with a gear ratio adjusting means for controlling a ratio, characterized in that the control device includes a pressure increasing means for increasing the pressure of oil pressure supplied to a drive pulley when the gear ratio is at a maximum gear ratio. A control device for a continuously variable transmission. (2) The boosting means is driven such that the ratio of the pressing force Fp applied to the belt from the driving pulley to the pressing force Fs applied to the belt from the driven pulley increases in response to an increase in input torque to the driving pulley. 2. The control device for a continuously variable transmission according to claim 1, wherein the control device is a means for boosting and controlling the hydraulic pressure supplied to the hydraulic chamber of the pulley.