JP2003343711A - Belt sliding determination device of belt type non-stage transmission for vehicle and control device of belt type non-stage transmission for vehicle - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To improve durability of a V belt by preventing sliding of the V belt of a non-stage transmission and simultaneously preventing application of nipping pressure more than necessary on the V belt. <P>SOLUTION: Target speed-changing speed is set in accordance with time to stop a vehicle in down-shift motion at the time of decelerating the vehicle (S106), and whether the sliding is caused on the V belt 34 or not is determined in accordance with belt nipping pressure lowering quantity PD<SB>dwn</SB>(n) (S108). In the case when this determination result is YES, whether sliding prevention of the V belt 34 is possible or not is determined (S112) in accordance with the belt nipping pressure lowering quantity PD<SB>dwn</SB>(n) and belt nipping pressure maximum compensating quantity PD<SB>up</SB>(n). In the case when this determination result is NO, the target speed changing speed is reset to a value at which the sliding prevention of the V belt 34 is possible. Finally, a duty ratio DS2(n) is computed in accordance with the target speed changing speed and a physical model (S109). <P>COPYRIGHT: (C)2004,JPO

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a control device for a belt type continuously variable transmission, and more particularly to a device for determining belt slippage of a belt type continuously variable transmission.

[0002]

2. Description of the Related Art Conventionally, the problems to be solved by the invention
A continuously variable transmission is used as a transmission for automobiles and the like. In this continuously variable transmission, in the belt type, the V-belt is wound around the primary sheave on the engine side and the secondary sheave on the wheel side, and the gear ratio is continuously changed by changing the groove widths of the primary sheave and the secondary sheave. Has been changed to.

In this continuously variable transmission, the driving force for changing the gear ratio is generally generated by the hydraulic pressure from a hydraulic actuator. Specifically, by controlling the hydraulic pressure supplied to the oil chamber of the primary sheave by the hydraulic actuator, the radius of gyration of the portion where the V-belt is wound around the primary sheave and the secondary sheave is changed, and the gear shifting operation is performed. On the other hand, the oil pressure of the secondary sheave is supplied with the hydraulic pressure controlled by another hydraulic actuator according to the input torque to the continuously variable transmission. This gives the V-belt proper clamping force, and the sheave and V
The slippage between the belt and the belt is suppressed and the durability of the V-belt is improved. Further, a pump is used as a hydraulic pressure source of the hydraulic pressure generated by the hydraulic actuator, and the pump is driven by the engine.

When the vehicle using the continuously variable transmission is decelerated and stopped, it is necessary to downshift the gear ratio to the maximum gear ratio in order to smoothly restart the vehicle.
Therefore, in a situation where the vehicle decelerates rapidly and then stops, it is necessary to increase the downshift speed. But,
When the downshift speed is fast, the moving speed of the sheave becomes faster, so it may not be possible to supply the hydraulic oil from the pump to the oil chamber of the secondary sheave to provide the proper belt clamping pressure. There will be a slip between and.

In Japanese Unexamined Patent Publication No. 6-109121, there is a stepless method for setting the line pressure so that the flow rate of the pump discharge flow rate that can be used for controlling the line pressure is equal to the flow rate necessary for shifting when the vehicle is decelerating. A hydraulic control device for a transmission is disclosed. In this conventional device, when the vehicle is decelerated, the line pressure is maximized under the condition that the flow rate required for gear shifting can be secured, so that the hydraulic oil supply can be prevented from being insufficient and the downshift speed can be maximized. It is made as fast as possible, and the gear ratio can be downshifted to the maximum gear ratio before the vehicle stops.

However, in this conventional apparatus, when the vehicle is decelerated, the line pressure is constantly increased to the condition that the flow rate available for line pressure control and the flow rate required for gear shifting become equal. Depending on the case, the line pressure may be increased more than necessary. In that case, since the V-belt is given an excessive clamping force, there is a problem that the load on the V-belt increases and the durability of the V-belt deteriorates.

The present invention has been made in view of the above problems, and is a vehicle for improving the durability of a V-belt by preventing the V-belt from slipping and at the same time preventing the V-belt from exerting an excessive clamping force. It is an object of the present invention to provide a belt slippage determination device for a belt type continuously variable transmission and a control device for a vehicle belt type continuously variable transmission including the same.

[0008]

In order to achieve such an object, the belt slip determining device for a vehicle belt type continuously variable transmission according to the first aspect of the present invention receives a drive torque from a prime mover. A device for determining the slippage of a belt of a vehicle belt type continuously variable transmission having a primary sheave, a secondary sheave that outputs the drive torque to a load, and a belt that is wound around the primary sheave and the secondary sheave. A hydraulic pressure source for generating a hydraulic pressure for supplying hydraulic oil to the primary sheave and the secondary sheave, and a shift control means for controlling a gear ratio by adjusting the amount of hydraulic oil supplied from the hydraulic pressure source to the primary sheave. , Clamping pressure control means for controlling the belt clamping pressure by adjusting the amount of hydraulic oil supplied from the hydraulic pressure source to the secondary sheave, and preventing the belt from slipping. In addition, a clamping pressure control command value calculating means for calculating a clamping pressure control command value to be output to the clamping pressure control means, a hydraulic pressure source output calculating means for calculating the discharge amount of hydraulic oil generated by the hydraulic pressure source, and a target shift speed. And a belt slip determining means for determining the slip of the belt based on the target shift speed and the calculated value of the hydraulic power source output calculating means.

By thus determining the slippage of the belt on the basis of the target shift speed and the calculated value of the hydraulic power source output calculating means, the slippage of the belt can be prevented during the shift operation, and at the same time, the clamping force applied to the belt is higher than necessary. Can be given. Therefore, the durability of the belt can be improved.

A belt slip determining device for a vehicle belt type continuously variable transmission according to a second aspect of the present invention is the device according to the first aspect of the present invention, wherein the hydraulic power source output calculating means is the hydraulic power source. It is characterized in that a value is calculated when the discharge amount of the generated hydraulic oil is increased from the present time.

A belt slip determining apparatus for a vehicle belt type continuously variable transmission according to a third aspect of the present invention is the apparatus according to the first or second aspect of the present invention, which predicts a time until the vehicle stops. Further, the shift speed setting means sets the target shift speed based on the predicted value of the stop predicting means.

As described above, by setting the target shift speed based on the time until the vehicle stops, it is possible to reliably perform the downshift operation up to the maximum gear ratio before the vehicle stops during deceleration. .

A belt slip determining device for a vehicle belt type continuously variable transmission according to a fourth aspect of the present invention is the device according to the third aspect of the present invention, wherein the stop predicting means detects a road surface state. It is characterized in that it has a state detecting means, and predicts a time until the vehicle stops based on a detection value of the road surface state detecting means.

As described above, by predicting the time until the vehicle stops based on the detection value of the road surface state detecting means, the time until the vehicle stops can be reduced even in the deceleration condition where the ABS operates. It is possible to predict accurately without an acceleration sensor that detects speed. Therefore,
Cost reduction can be realized.

A belt slip determining device for a vehicle belt type continuously variable transmission according to a fifth aspect of the present invention is the device according to the fourth aspect of the present invention, wherein the road surface state detecting means is a road surface gradient and a road surface. It is characterized in that at least one of the friction coefficients is detected.

A control device for a vehicle belt type continuously variable transmission according to a sixth aspect of the present invention is a control device for a vehicle belt type continuously variable transmission including the device according to the first aspect of the present invention. A shift control command value calculating means for calculating a control command value and outputting the control command value to the shift control means; and a differential pressure detecting means for detecting a pressure difference before and after the hydraulic fluid passes through the shift control means. The shift control command value calculating means calculates the shift control command value based on the target shift speed and the detection value of the differential pressure detecting means when the belt slip determining means determines that there is no belt slip. Characterize.

As described above, when the belt slippage determination means determines that there is no belt slippage, the gearshift control command value is calculated based on the target shift speed and the detection value of the differential pressure detection means, so that the belt slippage and the unnecessary amount are performed. It is possible to prevent the belt clamping pressure from being applied and to perform the gear shift operation with accuracy according to the target value.

A vehicle belt type continuously variable transmission control device according to a seventh aspect of the present invention is a vehicle belt type continuously variable transmission control device including the device according to the second aspect of the present invention. A shift control command value calculating means for calculating a control command value and outputting it to the shift control means, a differential pressure detecting means for detecting a pressure difference before and after the hydraulic fluid passes through the shift control means, and a hydraulic power source. When the belt slip determination means determines that there is no belt slip, the shift control command value calculation means determines the target shift speed and the detection value of the differential pressure detection means. Based on the calculated value of the hydraulic power source output calculation means, the hydraulic pressure source drive means increases the discharge amount of hydraulic oil generated by the hydraulic pressure source from the present time. To do.

As described above, by increasing the discharge amount of the hydraulic oil generated by the hydraulic pressure source from the present time, it is possible to compensate for the shortage of the hydraulic oil discharge amount of the hydraulic pressure source when obtaining the necessary belt clamping pressure. Therefore, the slippage of the belt can be surely prevented, and at the same time, the shift speed can be increased.

A control device for a vehicle belt type continuously variable transmission according to an eighth aspect of the present invention is the device according to the seventh aspect of the present invention, wherein the hydraulic power source drive means is the prime mover. The discharge amount of the hydraulic oil generated by the hydraulic pressure source is increased by increasing the rotation speed.

A vehicle belt type continuously variable transmission control device according to a ninth aspect of the present invention is a vehicle belt type continuously variable transmission control device including the device according to the first or second aspect of the present invention. A shift control command value calculating means for calculating a shift control command value and outputting the shift control command value to the shift control means, and a differential pressure detecting means for detecting a pressure difference before and after the hydraulic fluid passes through the shift control means. If the belt slip determination means determines that belt slip has occurred, the shift speed setting means resets the target shift speed based on the discharge amount of hydraulic oil generated by the hydraulic pressure source, and the shift control is performed. The command value calculating means is characterized by calculating the shift control command value based on the reset target shift speed and the detection value of the differential pressure detecting means.

In this way, when the belt slip determination means determines that belt slip has occurred, the target shift speed is reset based on the discharge amount of hydraulic oil generated by the hydraulic pressure source, and the target shift speed is reset. Since the shift control command value is calculated based on this, it is possible to prevent a situation in which the hydraulic oil discharge amount of the hydraulic power source for obtaining the belt clamping pressure required during the shift operation is insufficient. Therefore, slippage of the belt can be reliably prevented.

A vehicle belt type continuously variable transmission control device according to a tenth aspect of the present invention is the device according to any one of the sixth to ninth aspects of the present invention, and detects the rotational speed of the primary sheave. Input rotation speed detection means, output rotation speed detection means for detecting the rotation speed of the secondary sheave, input torque detection means for detecting the input torque to the primary sheave, and secondary pressure detection for detecting the pressure of the hydraulic oil in the secondary sheave. And a means for supplying the hydraulic pressure based on the pressure of the hydraulic source to the secondary sheave, wherein the differential pressure detecting means is provided for the input rotation. Based on the detection value of the speed detection means, the detection value of the output rotation speed detection means, the detection value of the input torque detection means, and the detection value of the secondary pressure detection means , And detecting the pressure difference before and after passing through the shift control means of the hydraulic oil.

As described above, the pressure difference between before and after the hydraulic oil passes through the speed change control means based on the rotational speed of the primary sheave, the rotational speed of the secondary sheave, the input torque to the primary sheave, and the hydraulic oil pressure in the secondary sheave. Since it is detected, the pressure sensor for detecting the hydraulic oil pressure in the primary sheave can be omitted, and the cost can be reduced.

[0025]

BEST MODE FOR CARRYING OUT THE INVENTION Embodiments of the present invention (hereinafter referred to as embodiments) will be described below with reference to the drawings.

FIG. 1 according to the embodiment of the present invention shows an overall configuration diagram in which the present invention is applied to the control of a belt type continuously variable transmission, in which a torque converter 10 connected to an engine output shaft 22 and a forward / reverse switching device. 12, belt type continuously variable transmission 14,
A hydraulic control device 40 that controls the gear ratio of the transmission 14 and an electronic control device 42 that controls the hydraulic pressure of the hydraulic control device 40 are provided. The drive torque output from the engine as the prime mover is the torque converter 10, the forward / reverse switching device 12,
It is transmitted to drive wheels (not shown) via the belt type continuously variable transmission 14 and a differential gear device (not shown).

The torque converter 10 includes a pump impeller 10a connected to the engine output shaft 22 and a pump impeller 10a connected to the torque converter output shaft 24 via a fluid.
A turbine impeller 10b to which a drive torque is transmitted from a fixed impeller 10c fixed to a position-fixed housing 10f via a one-way clutch 10e; and a pump impeller 10
A lock-up clutch 10d that fastens a and the turbine impeller 10b via a damper is provided.

The forward / reverse switching device 12 is provided with a double planetary type gear device and has a sun gear 12s, a carrier 12c and a ring gear 12r. Sun gear 12s
It is connected to the torque converter output shaft 24. The group of carriers 12c is connected to the torque converter output shaft 24 via a clutch 28 and is also connected to a belt type continuously variable transmission input shaft 26. The ring gear 12r is connected to the brake 12b.

The belt type continuously variable transmission 14 includes a primary sheave 30 connected to the input shaft 26, a secondary sheave 32 connected to the output shaft 36, and a V-shaped structure wound around the primary sheave 30 and the secondary sheave 32. Cross section V
The belt 34 is provided, and the primary sheave 3 is connected to the input shaft 26.
The torque transmitted to 0 is transmitted to the output shaft 36 via the V belt 34 and the secondary sheave 32.

The primary sheave 30 is composed of a primary movable sheave half 30a and a primary fixed sheave half 30b which are movable in the direction of the input shaft 26. Similarly, the secondary sheave 32 is composed of a secondary movable-side sheave half 32a and a secondary fixed-side sheave half 32b that are movable in the direction of the output shaft 36. The primary movable-side sheave half 30a moves in the direction of the input shaft 26 by the hydraulic pressure supplied to the primary oil chamber 30c. As a result, the radius of rotation of the portion of the V-belt 34 wound around the primary sheave 30 and the secondary sheave 32 changes, and the gear ratio of the belt type continuously variable transmission 14 continuously changes. Further, the belt clamping pressure is applied to the V belt 34 by the hydraulic pressure supplied to the secondary oil chamber 32c provided in the secondary movable-side sheave half 32a. As a result, slippage that occurs between the sheave and the V-belt 34 is suppressed.

The hydraulic pressures supplied to the primary oil chamber 30c and the secondary oil chamber 32c of the belt type continuously variable transmission 14 are supplied by the hydraulic control device 40, and those hydraulic pressures are controlled by the electronic control device 42.

The electronic control unit 42 has a throttle opening T
A throttle opening sensor 76 for detecting A, an engine rotation speed sensor 78 for detecting engine rotation speed Ne, an input shaft rotation speed sensor 80 for detecting rotation speed N _in of the input shaft 26, and a rotation speed N _out of the output shaft 36. The output shaft rotation speed sensor 82 for detecting, the oil temperature sensor 88 for detecting the oil temperature T _OIL of the hydraulic oil in the hydraulic control device 40, and the pressure for detecting the hydraulic oil pressure P _out in the secondary oil chamber 32c as the belt clamping pressure. A signal from the sensor 74 or the like is input. The electronic control unit 42 processes the input signal, and based on the processing result, the primary oil chamber 3 of the belt type continuously variable transmission 14 is processed.
0c and the hydraulic pressure supplied to the secondary oil chamber 32c are controlled.

Next, the main structure of the hydraulic control device 40 will be described with reference to FIG.

The line pressure control device 90 is provided with a linear solenoid valve (not shown), and regulates the output hydraulic pressure of the pump 52 as a hydraulic pressure source driven by the engine to a line pressure PL by the linear solenoid valve, This line pressure PL is output to the oil passage R1. here,
The pinching control command value for the linear solenoid valve is determined based on the input shaft 26 torque, and the line pressure PL is controlled according to the input shaft 26 torque. Secondary pressure control device 60
Supplies the belt clamping pressure adjusted according to the line pressure PL in the oil passage R1 to the secondary oil chamber 32c through the oil passage R3. Therefore, the belt clamping pressure is the line pressure PL.
Is controlled by a linear solenoid valve for controlling. In this way, the line pressure control device 90 and the secondary pressure control device 60 constitute the clamping pressure control means. Further, the oil passage R1 is provided with a constant pressure control device 70 for adjusting and outputting the line pressure PL so that the line pressure PL always becomes a constant oil pressure. The hydraulic pressure maintained constant by the constant pressure control device 70 is supplied to the speed increasing solenoid valve 66 and the speed reducing solenoid valve 68 described later through the oil passage R7.

Flow rate control device 50 as a shift control means
Controls the flow rate of the hydraulic oil flowing in and out of the primary oil chamber 30c of the primary sheave 30, and increases the flow rate control valve 62 for acceleration.
And a deceleration flow control valve 64, and a speed increasing solenoid valve 66 and a speed reducing solenoid valve 68 that supply control pressure to the speed increasing flow control valve 62 and the speed reducing flow control valve 64, respectively. The speed increasing flow control valve 62 includes four ports 62a, 62b,
62c, 62d, spool 6 which moves in the vertical direction in FIG.
2s, a spring 62 for pressing the spool 62s downward in FIG.
It has a control pressure chamber 62h to which f and control pressure are supplied. The speed increasing solenoid valve 66 has three ports 66a, 66.
b and 66c. When the speed increasing solenoid valve 66 is turned on (on the right side in FIG. 2), the ports 66a and 66b communicate with each other. Then, the speed-up solenoid valve 66 controls the hydraulic pressure, which is adjusted to a constant pressure in the oil passage R7, between the atmospheric pressure and this constant pressure by the duty control in which the valve is repeatedly turned on and off. From the port 62a of the flow control valve 62 to the control pressure chamber 62
supply to h. When the speed increasing solenoid valve 66 is off (on the left side in FIG. 2), the ports 66b and 66c communicate with each other, the hydraulic pressure of the control pressure chamber 62h is discharged from the port 66c to the reservoir 54, and is reduced to atmospheric pressure. .

When the control pressure from the speed increasing solenoid valve 66 is supplied to the control pressure chamber 62h from the port 62a of the speed increasing flow control valve 62, the spool 62s is pressed upward in FIG. 2 by this control pressure. . On the other hand, the spring 62f presses the spool 62s downward in FIG. 2, and the balance of these forces regulates the line pressure PL supplied from the port 62c through the oil passage R4 and the oil passage R from the port 62d.
It is supplied to the primary oil chamber 30c via

Similarly, the deceleration flow control valve 64 includes four ports 64a, 64b, 64c and 64d, a spool 64s that moves vertically in FIG. 2, a spring 64f that presses the spool 64s downward in FIG. It has a control pressure chamber 64h to which pressure is supplied. The deceleration solenoid valve 68 has three ports 68a, 68b, 68c. When the deceleration solenoid valve 68 is on (right side in FIG. 2), the ports 68a and 68b communicate with each other. The deceleration solenoid valve 68 controls the hydraulic pressure, which is adjusted to a constant pressure in the oil passage R7, between the atmospheric pressure and this constant pressure by the duty control in which it is repeatedly turned on and off, and the deceleration flow rate control is used as the control pressure. Port 64a of valve 64
To the control pressure chamber 64h. Also, the deceleration solenoid valve 6
When 8 is off (left side of Figure 2), ports 68b and 68c
, The hydraulic pressure of the control pressure chamber 64h is discharged from the port 68c to the reservoir 54, and is reduced to atmospheric pressure.

When the control pressure from the deceleration solenoid valve 68 is supplied to the control pressure chamber 64h from the port 64a of the deceleration flow control valve 64, the spool 64s is pressed upward in FIG. 2 by this control pressure. On the other hand, the spool 64s is pressed downward in FIG. 2 by the spring 64f, and the communication state between the port 64c and the port 64d is controlled by the balance of these forces, and the oil pressure supplied to the primary oil chamber 30c is controlled by the oil passage. It is discharged from the port 64d to the reservoir 54 through R5.

Next, the main configuration inside the electronic control unit 42 in FIG. 2 will be described.

In the electronic control unit 42, the speed increasing solenoid valve 6
6 and a shift control command value calculating means 124 for calculating a duty ratio of a duty control command value as a shift control command value to the deceleration solenoid valve 68 and a linear solenoid valve in the line pressure control device 90. A pinching control command value calculating means 126 for calculating the pressure control command value is provided. The shift control command value calculation means 124 uses a previously stored duty ratio-
The duty ratio for obtaining the desired speed change speed is calculated based on the orifice area characteristic, and the duty ratio of this duty ratio is calculated.
The tee control command value is output to the speed increasing solenoid valve 66 or the speed reducing solenoid valve 68. Further, in the present embodiment, the electronic control unit 42 includes the stop prediction means 130 for predicting the time until the vehicle stops, the shift speed setting means 134 for setting a target shift speed by a method described later, and the operation generated by the pump 52. Hydraulic power source output calculating means 136 for calculating the discharge flow rate of oil,
A differential pressure detecting means 128 for detecting a pressure difference before and after passing through the speed increasing flow control valve 62 or the decelerating flow control valve 64 of the hydraulic oil and a belt slip determining means 138 for determining slip of the V belt 34 are provided. . The stop predicting means 130 has a road surface state detecting means 132 for detecting the road surface gradient and the road surface friction coefficient.

Next, a vehicle deceleration downshift control routine executed in the electronic control unit 42 will be described with reference to the flow chart shown in FIG. The execution of the downshift control routine at the time of vehicle deceleration is repeated every predetermined time during vehicle deceleration.

First, in step (hereinafter referred to as S) 101, the input torque T _in (n) and the gear ratio γ (n) at the current sampling time n are calculated. The input torque T _in (n) is the engine torque Te (n) and the torque ratio t of the torque converter 10.
It can be calculated based on (n) and the input inertia torque. Here, the engine torque Te (n) can be calculated, for example, from the throttle opening TA (n) and the engine rotation speed Ne (n), and the torque ratio t (n) can be calculated by (the input shaft 26 rotation speed N _in (n ) / Engine rotation speed Ne (n)), and the input inertia torque can be calculated from the time variation of the input shaft 26 rotation speed N _in (n). Also, the gear ratio γ
(n) can be calculated by (input shaft 26 rotation speed N _in (n) / output shaft 36 rotation speed N _out (n)).

Next, in S102, the input torque T _in (n)
And the target belt clamping pressure PD based on the gear ratio γ (n)
Calculate _act (n). The target belt clamping pressure PD _act (n) is expressed by the following equation (1).

[0044]

## EQU1 ## PD _act (n) = K × T _in (n) × cos α / (2 × μ × RIN (n) × S _out ) (1)

Here, K is a coefficient of 1 or more and is set as a safety factor. RIN (n) is the suspension radius of the V belt 34 on the primary sheave 30, and is a function of the gear ratio γ (n). Further, α is an inclination angle of the sheave, μ is a coefficient of friction between the V belt 34 and the sheave, and S _out is a pressure receiving area of the secondary movable-side sheave half 32a.

Next, in S103, the road surface condition detecting means 1
At 32, the road surface condition of the road slope and the road surface friction coefficient is detected. The road gradient is calculated, for example, from the difference between the reference acceleration and the actual acceleration immediately before deceleration. The reference acceleration is calculated, for example, from the engine torque Te (n), and the actual acceleration is calculated, for example, from the time change of the output shaft 36 rotational speed N _out (n) when an acceleration sensor that detects the vehicle body deceleration is not used. Further, the road surface friction coefficient is calculated, for example, based on the wheel speed difference between the drive wheel and the driven wheel immediately before deceleration. Next, in S104, the vehicle stop deceleration is calculated by the stop prediction unit 130. When the acceleration sensor is not used, the vehicle deceleration is calculated from, for example, the time change of the output shaft 36 rotation speed N _out (n).

Next, in S105, the stop prediction means 130.
At, the time until the vehicle stops is calculated. The time until the vehicle stops is calculated from the vehicle deceleration, the road gradient and the road surface friction coefficient. Thus, by considering not only the vehicle deceleration but also the road gradient and the road surface friction coefficient, even in a vehicle deceleration situation where the drive wheels are almost locked (the ABS operates), the acceleration sensor is not used. It is possible to accurately calculate the time until the vehicle stops. Next, in S106, the shift speed setting means 134 sets the target shift speed required for downshifting from the time until the vehicle stops and the current gear ratio γ (n) to the maximum gear ratio before the vehicle stops. .

Next, in S107, the belt slippage determination means 138 calculates the belt clamping pressure decrease amount PD _dwn (n). Here, the lowering of the belt clamping pressure occurs when the flow rate of the working oil that has flowed into the secondary oil chamber 32c during the downshift becomes less than the required inflow rate. Therefore, the belt clamping pressure decrease amount PD _dwn (n) can be calculated based on the difference between the required inflow flow rate into the secondary oil chamber 32c and the actual inflow flow rate into the secondary oil chamber 32c. Here, the required inflow flow rate into the secondary oil chamber 32c can be calculated based on the target shift speed. Regarding the actual flow rate into the secondary oil chamber 32c, of the discharge flow rate of the pump 52, the secondary oil chamber 3
Calculated from the rate of consumption as the flow rate into 2c,
Calculated based on the hydraulic oil temperature T _OIL (n), the engine rotation speed Ne (n) (or the input shaft 26 rotation speed N _in (n)), and the control command value to the linear solenoid valve in the line pressure control device 90. be able to.

Next, in S108, the V-belt 3 is detected by the belt slippage determination means 138 due to the decrease in the belt clamping pressure.
It is determined whether or not slippage occurs at No. 4. Here, by determining whether the following expression (2) is satisfied,
It is determined whether the V-belt 34 slips.

[0050]

## _EQU00002 ## PD _dwn (n)> (K−L) × T _in (n) × cos α / (2 × μ × RIN (n) × S _out ) (2)

Here, L is a predetermined value of 1 or more and K or less. Further, here, the V belt 34 is determined by determining whether or not the belt clamping pressure decrease amount PD _dwn (n) is larger than a threshold value.
The slip determination of may be simplified. If the determination result in S108 is NO, the process proceeds to S109 described later. On the other hand, S10
If the determination result in 8 is YES, the process proceeds to S110.

In S110, the belt slippage judging means 138.
Belt clamping pressure maximum compensation amount PD_upCalculate (n)
It Here, the discharge flow rate of the pump 52 is the secondary oil chamber 3
If the flow rate is higher than the required inflow rate to 2c,
The clamping pressure can be compensated. Therefore, the belt
Clamping pressure maximum compensation PD_up(n) is the discharge flow rate of the pump 52
And the required flow rate into the secondary oil chamber 32c.
Can be calculated. The discharge flow rate of the pump 52 is
In the hydraulic power source output calculating means 136, the hydraulic oil temperature T
_OIL(n) and engine speed Ne (n) (or input shaft
26 rotation speed N_in(n))
The target inflow rate into the secondary oil chamber 32c
It can be calculated based on speed. At this time
Use the current value for the discharge flow rate of the pump 52
The discharge flow rate of the pump 52, that is, the engine
Rotational speed Ne (n) (input shaft 26 rotational speed N_in(n)))
You may use the value at the time of making it increase from the present time.
However, the engine rotation speed Ne (n) (input shaft 26 rotation speed
Degree N_in(n))) is increased, output shaft 36 torque
Increase within the range that does not affect the increase of. Well
In addition, here, the determination result of S112 described later is YES.
Engine rotation speed Ne (n) (input shaft 26 rotation speed for
Degree N_in(n))) may be calculated backward. Then S111
The belt slip determination means 138 advances to the target bell.
Clamping force PD_act(n), belt clamping pressure drop PD_{d wn}(n)
And belt clamping force maximum compensation amount PD_upPD based on (n)
_upThe belt clamping pressure after the compensation of (n) is calculated.

Next, in S112, the belt slip determining means 138 determines whether or not the V belt 34 can be prevented from slipping. The discharge flow rate of the pump 52 is the secondary oil chamber 32.
If the flow rate is smaller than the required flow rate into c, the belt clamping pressure cannot be compensated, so here the hydraulic oil temperature T
Pump 52 calculated based on _OIL (n) and engine speed Ne (n) (or input shaft 26 speed N _in (n))
It is determined whether or not the discharge flow rate value of is greater than or equal to the required inflow flow rate value into the secondary oil chamber 32c calculated based on the target shift speed. Further, when the belt clamping pressure can be compensated, that is, P
Even if D _up (n)> 0, if the belt clamping pressure value after compensating for PD _up (n) is smaller than (L / K) × PD _act (n), the belt clamping pressure is still sufficiently compensated at that time. Since it has not been completed, the belt clamping pressure value after compensating for PD _up (n) is (L
/ K) × PD _act (n) or more, it is possible to more accurately determine whether slip prevention is possible. If at least one of the above two determinations is NO, S112
The result of the determination is NO, and the process proceeds to S113. If both of the above two determinations are YES, the determination result of S112 is YE.
It becomes S and proceeds to S109.

At S113, at the present target shift speed, V
When it is judged that the slippage of the belt 34 cannot be prevented, the target speed change speed is reset by the speed change speed setting means 134 and S1 is set.
Go to 09. The target shift speed to be reset is calculated as follows. First, PD _up (n) such that the belt clamping pressure value after compensation is (L / K) × PD _act (n) or more
_Is calculated based on the target belt clamping pressure PD _act (n) and the belt clamping pressure decrease amount PD _dwn (n). Next, this P
The flow rate difference between the discharge flow rate of the pump 52 and the required inflow flow rate into the secondary oil chamber 32c, which makes it possible to compensate the belt clamping pressure by the value of D _up (n), is calculated, and the required inflow flow rate into the secondary oil chamber 32c is calculated. The target shift speed is set so as to be smaller than the discharge flow rate of the pump 52 by this flow rate difference or more. With respect to the discharge flow rate of the pump 52, the hydraulic power source output calculation means 136 calculates the hydraulic oil temperature T _OIL (n) and the engine rotation speed Ne (n).
(Or calculated based on the rotation speed N _in (n) of the input shaft 26), and either the current value or the value when the discharge flow rate is increased from the current time may be used. Further, here, by resetting the target shift speed such that the required inflow flow rate into the secondary oil chamber 32c is smaller than the discharge flow rate of the pump 52 (either the current value or the value when increasing) by a predetermined amount. The reset of the target shift speed may be simplified.

In S109, the shift control command value calculation means 1
24, the duty ratio DS2 (n) of the duty control command value to the deceleration solenoid valve 68 is calculated to calculate the deceleration solenoid valve 6
8 and the execution of this routine is completed. Here, the duty ratio DS2 (n) is calculated as follows. First, from the set target shift speed, gear ratio γ (n) and sheave half position-gear ratio characteristic, the primary movable sheave half 3 is moved.
Calculates a target speed of movement of 0a, calculates the target outflow rate Q _{ou t} (n) in the primary oil chamber 30c from the target moving velocity. Next, the target orifice area A (n) of the deceleration flow control valve 64 is calculated using the physical model shown in the equation (3).

[0056]

## _EQU00003 ## Q _out (n) = C × A (n) × (2 × P _in (n) / ρ) ^0.5 (3)

Here, C is the flow coefficient, ρ is the density of the hydraulic oil, and P _in (n) is the pressure in the primary oil chamber 30c. The flow coefficient C is determined by the orifice area A (n) and the hydraulic oil temperature T.
It is set experimentally from _OIL (n) and the like. When the pressure sensor is not used, the pressure P _in (n) of the primary oil chamber 30c can be calculated by the differential pressure detection means 128 using the equation (4).

[0058]

## EQU00004 ## P _in (n) = (W _in (n) −k _in × N _in (n) ² ) / S _in (4)

Here, k _in is the primary sheave centrifugal oil pressure coefficient, and S _in is the pressure receiving area of the primary movable sheave half 30a. W _in (n) is the thrust of the primary movable-side sheave half 30a at time n, and is represented by equation (5).

[0060]

## EQU00005 ## W _in (n) = W _out (n) / (a + b × log ₁₀ γ (n) + c × T _in (n) + d × N _in (n)) (5)

Here, the coefficients a, b, c and d are obtained by experiments. W _out (n) is the thrust of the secondary movable-side sheave half 32a at time n, and is expressed by equation (6).

[0062]

## _EQU6 ## W _out (n) = P _out (n) × S _out + k _out × N _out (n) ² (6)

Here, P _out (n) is the secondary oil chamber 32
The pressure of c (detected by the pressure sensor 74), k _out is the secondary sheave centrifugal hydraulic pressure coefficient.

Next, from the orifice area A (n) and the duty ratio-orifice area characteristic map, the duty ratio DS2
Calculate the value of (n). At this time, the duty ratio DS2 (n)
Between the orifice area and the orifice area A (n) (operating oil temperature T
It is preferable to calculate the value of the duty ratio DS2 (n) in consideration of _OIL (n) function.

The duty control command value is set to the deceleration solenoid valve 68.
At the same time, the clamping pressure control command value is output to the linear solenoid valve in the line pressure control device 90. If the determination result in S108 is NO, the clamping pressure control command value for obtaining the target line pressure set based on the input torque T _in (n) is output. On the other hand, the determination result of S108 is YES
If it is, the belt clamping pressure value after compensation is (L /
K) × PD _act (n) or more to output the pinching control command value for ensuring the above.

Further, at the same time as outputting the duty control command value to the deceleration solenoid valve 68, the drive torque of the engine is controlled as necessary to increase the engine rotation speed Ne (n). Specifically, if the determination result of S108 is NO, or the determination result of S112 is YES.
When the current value is used to calculate the discharge flow rate of the pump 52 in S110, or when the current value is used to calculate the discharge flow rate of the pump 52 in S113,
The control for increasing the engine rotation speed Ne (n) is not performed. On the other hand, if the determination result in S112 is YES and a value obtained by increasing the discharge flow rate from the present time is used in the calculation of the discharge flow rate of the pump 52 in S110, or the discharge flow rate is increased from the present time in the calculation of the discharge flow rate of the pump 52 in S113. When using the value thus set, control is performed to increase the engine rotation speed Ne (n) so that the discharge flow rate of the pump 52 increases to that value.

In this embodiment, the V-belt 34 is based on the target belt clamping pressure PD _act (n), the belt clamping pressure decrease amount PD _dwn (n) and the belt clamping pressure maximum compensation amount PD _up (n).
Whether or not slipping is prevented is determined, and the target shift speed is set according to the determination result. Here, since the capacity of the secondary oil chamber 32c increases during the downshift operation, the secondary oil chamber 32 for ensuring the necessary belt clamping pressure.
If the pump 52 cannot generate the required inflow flow rate into c, the V-belt 34 will slip. However, in the present embodiment, it is possible to prevent the V-belt 34 from slipping at the same time by determining whether or not the belt clamping pressure required during the downshift operation can be secured by the discharge flow rate of the pump 52. It is possible to prevent applying an excessive clamping force. Therefore, the durability of the belt can be improved.

The target shift speed when the V-belt 34 can be prevented from slipping is set on the basis of the time until the vehicle stops. Therefore, during deceleration, the maximum gear ratio is reached before the vehicle stops. The downshift operation can be reliably performed. Further, since the time until the vehicle stops is calculated based on the road surface gradient and the road surface friction coefficient, the time until the vehicle stops in a deceleration situation where the ABS operates can be accurately calculated without the acceleration sensor that detects the vehicle deceleration. It can be calculated and cost reduction can be realized.

On the other hand, regarding the target shift speed when the V-belt 34 cannot be prevented from slipping, downshift is performed by resetting the target shift speed for ensuring the necessary belt clamping pressure by the discharge flow rate of the pump 52. It is possible to prevent a situation where the discharge flow rate of the pump 52 for obtaining the belt clamping pressure required during the operation is insufficient. Therefore, slippage of the V-belt 34 can be reliably prevented, and the downshift operation can be performed at the maximum speed change speed within the range in which the V-belt 34 does not slip.

When the duty ratio DS2 (n) is calculated based on the target shift speed, the physical model relating to the deceleration flow control valve 64 shown in the equation (3) is used. Therefore, it is possible to accurately control the shift speed according to the target value. Then, the optimum adjustment of the duty ratio DS2 (n) and the pinching control command value in the downshift operation during vehicle deceleration can be easily performed. Further, when calculating the duty ratio DS2 (n), the hydraulic oil pressure P _in (n) _in the primary oil chamber 30c is obtained using the physical model shown in equations (4) to (6). The pressure sensor for detecting the hydraulic oil pressure in the primary oil chamber 30c can be omitted, and the cost can be reduced.

Further, by increasing the engine rotation speed Ne (n) during the downshift operation to increase the discharge flow rate of the pump 52, it is possible to compensate for the shortage of the discharge flow rate of the pump 52 when the necessary belt clamping pressure is obtained. You can Therefore, the maximum value of the downshift speed that can prevent the V-belt 34 from slipping can be increased, and the downshift operation can be reliably performed up to the maximum gear ratio before the vehicle stops even at high deceleration. .

[0072]

As described above, according to the present invention,
Since the belt slippage is determined based on the target shift speed and the calculated value of the hydraulic power source output calculating means, it is possible to prevent the belt slippage during the gear shift operation and at the same time prevent the belt from being subjected to an excessive clamping force. . Therefore, the durability of the belt can be improved.

[Brief description of drawings]

FIG. 1 is a diagram showing a configuration of a vehicle power transmission device including a vehicle belt type continuously variable transmission control device according to an embodiment of the present invention.

FIG. 2 is a diagram showing a schematic configuration of a hydraulic control device and an electronic control device according to an embodiment of the present invention.

FIG. 3 is a flowchart showing a downshift control routine during vehicle deceleration according to the embodiment of the present invention.

[Explanation of symbols]

10 torque converter, 12 forward / reverse switching device, 14
Belt type continuously variable transmission, 30 primary sheaves, 32
Secondary sheave, 34 V belt, 40 hydraulic control device, 42 electronic control device, 50 flow control device, 52
Pump, 60 secondary pressure control device, 62 speed increase flow control valve, 64 deceleration flow control valve, 66 speed increase solenoid valve, 68 speed reduction solenoid valve, 90 line pressure control device, 1
24 shift control command value calculation means, 126 pinching control command value calculation means, 128 differential pressure detection means, 130 stop prediction means, 132 road surface state detection means, 134 shift speed setting means, 136 hydraulic power source output calculation means, 138 belt slip Judgment means.

─────────────────────────────────────────────────── ─── Continuation of the front page (51) Int.Cl. ⁷ Identification code FI theme code (reference) F16H 59:42 F16H 59:42 59:66 59:66 59:68 59:68 63:06 63:06 ( 72) Inventor Kenji Matsuo 1 Toyota Town, Toyota City, Aichi Prefecture, Toyota Motor Corporation (72) Inventor Masato Terashima 1 Toyota Town, Toyota City, Aichi Prefecture Toyota Motor Corporation, (72) Inventor, Hiroki Kondo Aichi Prefecture Toyota-City, Toyota-City Toyota Motor Co., Ltd. F-term (reference) 3J552 MA07 MA12 NA01 NB01 PA12 RA02 RB18 RB22 RB23 RB26 SA32 SA36 SA59 TA10 TB03 VA18W VA32W VA34W VA37W VA47Z VA05WZW04Z01V01Z01

Claims (10)

[Claims] 1. A belt-type vehicle-less vehicle having a primary sheave to which drive torque is input from a prime mover, a secondary sheave that outputs the drive torque to a load, and a belt that is wound around the primary sheave and the secondary sheave. A device for determining the slippage of a belt of a variable speed transmission, the hydraulic source generating hydraulic pressure for supplying hydraulic oil to a primary sheave and a secondary sheave, and the amount of hydraulic oil supplied from the hydraulic source to the primary sheave. To control the gear ratio, the clamping pressure control means for controlling the belt clamping pressure by adjusting the amount of hydraulic oil supplied from the hydraulic source to the secondary sheave, and the belt slip prevention. Clamping pressure control command value calculation means for calculating a clamping pressure control command value to be output to the clamping pressure control means, and discharge of hydraulic oil generated by the hydraulic pressure source. A hydraulic pressure source output calculating means for calculating the amount, a shift speed setting means for setting a target shift speed, and a belt slippage judging means for judging belt slippage based on the target shift speed and the calculated value of the hydraulic pressure source output calculating means. A belt slip determination device for a belt type continuously variable transmission for a vehicle, comprising: 2. The belt slip determination device for a vehicle belt type continuously variable transmission according to claim 1, wherein the hydraulic power source output calculation means determines the discharge amount of the hydraulic oil generated by the hydraulic pressure source from the present time. A belt slip determination device for a belt type continuously variable transmission for a vehicle, which calculates a value when the number is increased. 3. The belt slip determination device for a vehicle belt type continuously variable transmission according to claim 1, further comprising stop prediction means for predicting a time until the vehicle stops. The belt slippage determination device for a belt type continuously variable transmission for a vehicle, wherein the speed setting means sets the target shift speed based on a predicted value of the stop predicting means. 4. The belt slip determination device for a vehicle belt type continuously variable transmission according to claim 3, wherein the stop predicting means includes road surface state detecting means for detecting a road surface state. A belt slippage determination device for a belt type continuously variable transmission for a vehicle, which predicts a time until the vehicle stops based on a detection value of a detection means. 5. The belt slip determination device for a vehicle belt type continuously variable transmission according to claim 4, wherein the road surface condition detecting means detects at least one of a road surface slope and a road surface friction coefficient. A belt slip determination device for a belt type continuously variable transmission for a vehicle. 6. A control device for a vehicle belt type continuously variable transmission, including the belt slippage determination device for a vehicle belt type continuously variable transmission according to claim 1, wherein a shift control command value is calculated and the shift control command value is calculated. The shift control command value calculating means; and a differential pressure detecting means for detecting a pressure difference before and after the hydraulic oil passes through the shift control means. When the belt slip determination means determines that there is no belt slip, the shift control command value is calculated based on the target shift speed and the detection value of the differential pressure detection means. Gearbox control device. 7. A vehicle belt type continuously variable transmission control device including the belt slip continuously variable transmission device for a vehicle belt type continuously variable transmission according to claim 2, wherein a shift control command value is calculated to obtain the shift control command value. A shift control command value calculating means for outputting to the shift control means, a differential pressure detecting means for detecting a pressure difference between hydraulic oil before and after passing through the shift control means, and a hydraulic source driving means for driving the hydraulic source. Further, when the belt slip determination means determines that there is no belt slip, the shift control command value calculation means calculates the shift control command value based on the target shift speed and the detection value of the differential pressure detection means. A belt-type continuously variable transmission for a vehicle, characterized in that the hydraulic pressure source drive means increases the discharge amount of the hydraulic oil generated by the hydraulic pressure source from the present time based on the calculated value of the hydraulic pressure source output calculation means. Machine control Location. 8. The control device for a vehicle belt type continuously variable transmission according to claim 7, wherein the hydraulic power source drive means is the prime mover, and the hydraulic power source is increased by increasing a rotational speed of the prime mover. A control device for a belt-type continuously variable transmission for a vehicle, which increases a discharge amount of hydraulic oil generated by the above. 9. A control device for a vehicle belt type continuously variable transmission including the belt slip determination device for a vehicle belt type continuously variable transmission according to claim 1, wherein a shift control command value is calculated. Further comprising: a shift control command value calculating means for outputting to the shift control means, and a differential pressure detecting means for detecting a pressure difference before and after the hydraulic oil passes through the shift control means. If it is determined that belt slippage has occurred, the shift speed setting means resets the target shift speed based on the discharge amount of hydraulic oil generated by the hydraulic pressure source, and the shift control command value calculation means resets the target shift speed. A control device for a belt type continuously variable transmission for a vehicle, wherein the shift control command value is calculated based on the target shift speed and the detection value of the differential pressure detection means. 10. A control device for a belt type continuously variable transmission according to claim 6, wherein the input rotation speed detection means detects a rotation speed of the primary sheave, and the secondary sheave of the secondary sheave. An output rotation speed detecting means for detecting a rotation speed; an input torque detecting means for detecting an input torque to the primary sheave; and a secondary pressure detecting means for detecting a pressure of the hydraulic oil in the secondary sheave, further comprising: The source is a control device for a vehicle belt type continuously variable transmission that supplies hydraulic pressure based on the pressure of the hydraulic pressure source to a secondary sheave, wherein the differential pressure detection means is a detection value of the input rotation speed detection means, the output rotation. Based on the detection value of the speed detection means, the detection value of the input torque detection means, and the detection value of the secondary pressure detection means, the shift control operation of the hydraulic oil is performed. A control device for a belt-type continuously variable transmission for a vehicle, which detects a pressure difference before and after passing through a step.