JPH07217713A - Control device of continuously variable transmission - Google Patents

Abstract

PURPOSE:To restrain and prevent slippage of a belt and pulley when brakes are applied on a road in low friction coefficient condition by outputting a signal to increase fluid pressure supplied between pulley sides to a specified quantity when brake application is detected. CONSTITUTION:A soft position is read (S1). Then the soft position is judged whether it is in P range or in N range (S2) and if it is in P range or in N range, it is restored to a main program but if not in P range or N range, judgment is made as to whether the output signal is off or not (S3) and, if it is off, the solenoid duty ratio DtSOL of a line pressure regulating electromagnetic valve is set to 100% (S4). If it is not off, a control signal SSOL corresponding to the respective duty ratio DtSOL is formed after setting the solenoid duty ratio DtSOL to 0% (S5) and after outputting the control signal SSOL to the line pressure regulating solenoid valve solenoid (S6), it is returned to the main program.

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 continuously variable transmission, and particularly to a low friction coefficient state (this friction coefficient state is also referred to as μ) such as an ice and snow road surface or a wet tile road surface. It is suitable for sudden braking in.

[0002]

2. Description of the Related Art A belt-type continuously variable transmission that changes the input / output gear ratio by changing the so-called pulley ratio, which changes the contact point radius between the belt and the pulley, has recently been in view of its performance. It is said that it is preferable not to interpose a power transmission direction restricting means such as a one-way clutch interposed between the automatic transmission using the torque converter and the gear transmission mechanism and the output side thereof. Here, the upstream side of the belt type continuously variable transmission, that is, the engine side is defined as the input side, and the downstream side, that is, the power transmission system such as the propeller shaft and the differential device and the wheel side are defined as the output side.
In order to control the change of the pulley ratio of such a belt type continuously variable transmission, a fluid pressure, an electronically controlled mechanical mechanism, or the like is usually used, and specifically, a piston-shaped movable pulley piece ( By moving the movable conical plate relative to the fixed pulley piece (fixed conical plate), the width of the pulley groove formed between the two is changed and controlled. Here, among the continuously variable transmissions of this type, Japanese Patent Application Laid-Open No. 61-1053 previously proposed by the present applicant.
To cite the continuously variable transmission and its control device described in Japanese Patent No. 53, a stepping motor is mainly used to control the gear ratio of the continuously variable transmission, and the rotation angle of the stepping motor is controlled. By doing so, the width of the pulley groove formed between the movable pulley piece (movable conical plate) and the fixed pulley piece (fixed conical plate) is changed and controlled. On the other hand, a predetermined fluid pressure, specifically, a hydraulic pressure is applied to the movable pulley piece (movable conical plate) that is made into a piston, and this hydraulic pressure causes both pulley pieces (both conical plates) to move.
While sandwiching a belt that is interposed between and that rotates,
Even if the transmitted rotational driving force (torque) fluctuates, the width of the pulley groove does not change. Further, focusing on the continuously variable transmission and its control device described in JP-A-61-105353, among the input side and output side pulleys, particularly the output side pulley (hereinafter, also simply referred to as a secondary pulley). The movable pulley piece (movable conical plate) is a fluid pressure obtained by, for example, adding a throttle pressure according to the opening of the throttle valve (hereinafter, simply referred to as throttle opening) to the reference line pressure. That is, hydraulic pressure is being supplied, and a clamping force sufficient to prevent the belt from slipping is generated between both pulley pieces of the secondary pulley in accordance with the rotational driving force (torque) transmitted from the input side to the output side. I am doing it.

On the other hand, in general, in this type of continuously variable transmission, the speed change pattern for controlling the speed change ratio has a vehicle speed and an engine speed or an engine speed (hereinafter, these are also collectively referred to as an engine speed state). And, specifically, for example, the shift pattern is controlled with the vehicle speed and the throttle opening degree as variables. Therefore, since the throttle opening is reduced during braking, it is conceivable that the shift pattern of the continuously variable transmission may continue to be set to a certain gear ratio regardless of the actual engine rotation state. If the throttle opening is reduced and the vehicle speed is high to some extent, it is considered that the normal shift pattern is equal to the coast state, that is, the coasting state. With a certain possible gear ratio, the reduction ratio is the smallest in the actual gear ratio of the vehicle.

[0004] Here, in a high friction coefficient state (hereinafter, also simply referred to as μ) road surface such as a normal dry asphalt road surface or a concrete road surface, a case where a sudden braking to the extent that the wheels are locked, that is, a so-called sudden braking is performed. Suppose. In such a sudden braking on a high μ road surface, the deceleration of the wheel from the normal rotation state to the lock of the wheel is very large, and the grip force between the tire and the road surface is large. It is very large, and the detected vehicle speed decreases rapidly. Therefore, as described above, the gear ratio immediately before braking is relatively small in terms of the vehicle speed reduction ratio. Therefore, in the gear ratio control of the continuously variable transmission, the gear ratio is considered to rapidly increase during the sudden braking. . However, the gear ratio of the continuously variable transmission suddenly increases during the sudden braking operation as described above, and if the accelerator pedal is depressed after that, a sudden driving force is applied to the driving wheels, which causes the vehicle speed to change. Running stability may be impaired.

In order to solve such a problem, the applicant of the present invention has previously proposed a control device for a continuously variable transmission disclosed in Japanese Patent Laid-Open No. 4-254054. According to this control device for a continuously variable transmission, when the above-described sudden braking operation is performed, the gear ratio of the continuously variable transmission is fixed to the gear ratio immediately before that, that is, a relatively small gear ratio. After that, when the sudden braking operation is released and the drive wheels start to rotate, the gear ratio is released. Therefore, as long as the drive wheels are rotated by the sudden braking operation, the gear ratio is fixed to a small gear ratio. Immediately after the sudden braking operation, the accelerator pedal is depressed to apply the driving force to the driving wheels. However,
At least at that moment, there is no fear that traveling stability will be impaired. Further, when the accelerator pedal is stepped on and the vehicle is started after the vehicle is stopped by the sudden braking operation, the gear ratio fixing control is released in accordance with the rotation of the drive wheels. The gear is controlled to a large gear ratio according to the vehicle speed immediately after starting, that is, a small vehicle speed, and a sufficient rotational driving force is transmitted to the drive wheels to enable a smooth start. Note that it may be considered that the release of the gear ratio fixed control is performed only in accordance with the driving wheel rotation state such as the rotation speed or the number of rotations of the driving wheel. Specifically, the driving wheel rotation state value is “0”. The fixed gear ratio control is released at the moment when the speed increases in the positive direction.

[0006]

DISCLOSURE OF INVENTION Problems to be Solved by the Invention The speed ratio control by the control device for a continuously variable transmission described in the above-mentioned Japanese Patent Laid-Open No. 4-254054 is therefore performed on low μ road surfaces such as ice and snow road surfaces and wet tile road surfaces. Let's consider it. Here, it is assumed that the vehicle has no existing anti-skid control device in the braking system.

On such a low μ road surface, the state of the friction coefficient between the tire and the road surface is also small. Therefore, when a braking force is applied to the tire on the low μ road surface, the wheels are likely to be locked, and the slip ratio in automobile engineering is called. Grows larger. With the present tire characteristics, the grip force (equivalent to frictional force) of the tire involved in steering, driving and braking is secured within the slip ratio range of 10 to 30%. If the slip ratio (hereinafter also simply referred to as the actual slip ratio) becomes large, it becomes difficult to secure the steering effect and the braking distance. Conversely speaking, the range of the slip ratio of the wheel that can secure the grip force of this tire is set as the target slip ratio, and the wheel that satisfies the target slip ratio is the vehicle speed, that is, the target slip ratio calculated from the vehicle speed. When the speed range is set as the target wheel speed, the fact that the actual wheel speed is within the target wheel speed range can ensure the steering effect and the braking distance as the vehicle. At this time, the fact that the actual slip ratio is large beyond the range of the target slip ratio means that the grip force of the tire itself has decreased as described above.

This is described in the above-mentioned Japanese Patent Laid-Open No. 4-254054.
In the control device for a continuously variable transmission described in Japanese Patent Publication, a state in which a large braking force of a braking system is applied to wheels will be applied. At this time, the input to rotate the wheel against the large braking force applied to the wheel of the braking system is the vehicle speed (vehicle speed) due to the running inertia of the tire gripping on the road surface unless the accelerator pedal is depressed. The only input is to rotate the wheel to follow.

However, when the wheels, which are the output side ends of the actual gear shifting and power transmission system of such a continuously variable transmission, are rotated by the grip force of the tire and the vehicle body speed, during this period, the throttle is opened as described above. As the speed ratio of the continuously variable transmission is fixed to the speed ratio immediately before the sudden braking, the input for rotating this wheel reaches the input side of the continuously variable transmission. It is transmitted as rotation fluctuation.
Here, in a continuously variable transmission that does not have a power transmission direction regulating means such as the one-way clutch on the output side, the input transmission system from the wheels necessary for rotating the wheels has the continuously variable transmission at its end. It can be said that it has the inertial weight of the rotating system. Considering this in the input transmission system of the rotational driving force input to the wheels from the road surface (that is, also referred to as the road surface rotational driving force to the wheels because the road surface is the driving force to rotate the wheels), the road surface rotational driving force The deceleration of the wheel speed of the drive wheels (strictly speaking, the wheel speed decelerates itself) generated from the deviation between the braking force and the braking force causes the fluctuation of the inertia torque with respect to the continuously variable transmission in which the gear ratio is fixedly controlled. Will be granted. This is because the actual wheel speed is in the range of the target wheel speed,
Alternatively, it can be considered that the inertia torque applied to the continuously variable transmission increases as the wheels (driving wheels) exhibit a locking tendency, that is, as the deceleration of the wheel speed increases within a range slightly smaller than that.

Here, since the road surface rotational driving force is transmitted from the output side to the input side of the continuously variable transmission whose gear ratio is fixedly controlled, the driving wheel speed is reduced during the sudden braking on the low μ road surface. The inertia torque that increases with increasing speed is
The above-mentioned Japanese Patent Application Laid-Open No. 61-105353 discloses that the belt acts as a large rotational movement (blocking) force for blocking the rotational movement of the belt, and that the throttle opening is considered to be reduced during braking. It is considered that the fluid pressure applied to the output-side pulley (secondary pulley) is also reduced by the control device for the continuously variable transmission described above. There is a problem that the moving force becomes large and the belt slips between the pulley pieces. This is because the belt clamping force of the secondary pulley is just a magnitude necessary for transmitting the rotational driving force (torque) from the input side to the output side of the continuously variable transmission. This is because it is not assumed that the inertia torque of the continuously variable transmission is increased due to the increase in the drive wheel speed deceleration caused by the deviation between When the belt slips in this way, the durability of the belt is reduced.

The present invention has been developed in view of these problems, and particularly during braking on a low μ road surface, even if a large deceleration of the wheel speed (driving wheel speed) occurs, the belt-pulley can be driven. An object of the present invention is to provide a control device for a continuously variable transmission that can prevent slippage and prevent the durability of the belt from increasing.

[0012]

Means for Solving the Problems As a result of intensive studies to solve the above problems, the present inventor has obtained the following findings and developed the present invention. That is, as described above, in order to prevent the slippage of the belt that occurs during braking, including when braking on a low μ road surface, braking of the wheels by the braking system is performed by, for example, a foot brake switch. When
It suffices to increase the fluid pressure supplied to the pulley to increase the clamping force of the belt. At this time, in order to prevent the slippage of the belt from being suppressed to the maximum under any circumstances, the fluid pressure to be supplied to the pulley should be increased to the maximum during all braking. Here, increasing the fluid pressure supplied to the pulley to the maximum would also increase the fluid pressure in a sense by a predetermined increment. However, as described above, the slip between the belt and the pulley is caused by the deceleration of the driving wheel speed generated from the difference between the rotational driving force (road surface rotational driving force) due to the grip force between the road surface and the tire and the braking force. It can be said that it is an energy loss to increase the fluid pressure supplied to the pulley at any braking, if it is considered to be only when it becomes larger than a certain level. Therefore, if the deceleration of the drive wheel is detected and the fluid pressure supplied to the pulley is increased only when the drive wheel deceleration is larger than a predetermined value set in advance, it is possible to reduce the low μ road surface. It is equivalent to determining that the tire locks or tends to lock due to sudden braking at, while surely preventing the slippage of the belt that occurs during braking on the low μ road surface,
Energy loss can be reduced. Further, since the increase in the inertia torque for sliding the belt, that is, the rotational movement (blocking) force applied to the belt depends on the magnitude of the deceleration of the driving wheel speed, the fluid pressure supplied to the pulley is If the predetermined increase amount is set according to the magnitude of the deceleration of the drive wheel speed, it is possible to further reduce the energy loss while increasing the clamping force of the belt by the pulley as much as necessary. Further, since the deceleration of the driving wheel speed depends on the deviation between the road surface rotational driving force and the braking force, specifically, the braking force is detected by detecting the braking fluid pressure of the braking system. Is detected, and a predetermined increase amount of the fluid pressure supplied to the pulley is set according to the magnitude of the braking force,
The energy loss can be further reduced while increasing the clamping force of the belt by the pulley as much as necessary.

Thus, the control device for a continuously variable transmission according to the first aspect of the present invention is provided between a pair of relatively movable pulley pieces as shown in the basic configuration diagram of FIG. A control device for a continuously variable transmission, comprising: a fluid pressure adjusting means capable of changing and controlling a fluid pressure to change and control a clamping force of a belt clamped between the pair of pulley pieces, the wheel comprising a braking system. Detection means for detecting that the braking to the pulley has been executed, and a signal for increasing the fluid pressure supplied between the pulley pieces by a predetermined increment when the braking detection means detects the execution of the braking. And a fluid pressure increasing means for outputting to the pressure adjusting means.

The control device for a continuously variable transmission according to claim 2 of the present invention, as shown in the basic configuration diagram of FIG. 1, is a fluid supplied between a pair of relatively movable pulley pieces. A control device for a continuously variable transmission, comprising a fluid pressure adjusting means capable of changing and controlling the clamping force of a belt clamped between the pair of pulley pieces by increasing / decreasing the pressure. Braking detection means for detecting that the braking of the wheel has been executed, wheel deceleration detection means for detecting the deceleration of the wheel rotation, and the braking detection means detects the execution of the braking and is detected by the wheel deceleration detection means. When the detected wheel rotation deceleration value is greater than or equal to a predetermined wheel rotation deceleration value, a fluid pressure for outputting a signal for increasing the fluid pressure supplied between the pulley pieces by a predetermined increment toward the fluid pressure adjusting means. It is characterized by having an increasing means.

A control device for a continuously variable transmission according to a third aspect of the present invention is the control device for a continuously variable transmission according to the first aspect, as shown in the basic configuration diagram of FIG. Wheel deceleration detecting means for detecting deceleration of wheel rotation is provided, and the fluid pressure increasing means is supplied between the pulley pieces according to the wheel rotation deceleration detection value detected by the wheel deceleration detecting means. It is characterized in that a predetermined increase amount of the fluid pressure is set.

A control device for a continuously variable transmission according to a fourth aspect of the present invention is the control device for a continuously variable transmission according to the second aspect, as shown in the basic configuration diagram of FIG. The fluid pressure increasing means sets a predetermined increase amount of the fluid pressure supplied between the pulley pieces in accordance with a wheel rotation deceleration detection value detected by the wheel deceleration detecting means. Is.

The control device for a continuously variable transmission according to a fifth aspect of the present invention is, as shown in the basic configuration diagram of FIG. 1, for braking for detecting a fluid pressure for generating a braking force of a braking system. A fluid pressure detecting means is provided, and the fluid pressure increasing means determines a predetermined increase amount of the fluid pressure supplied between the pulley pieces in accordance with the braking fluid pressure detection value detected by the braking fluid pressure detecting means. It is characterized by setting.

[0018]

In the control device for a continuously variable transmission according to the first aspect of the present invention, as shown in the basic configuration diagram of FIG. 1, for example, the continuously variable transmission disclosed in Japanese Patent Laid-Open No. 61-105353 is disclosed. In the control device, a solenoid valve or the like that can increase or decrease the line pressure of the fluid pressure supplied between the pulley pieces separately from the throttle pressure of the fluid pressure supplied between the pulley pieces is used. An actuator is provided as a fluid pressure adjusting means, and on the other hand, a braking detecting means such as a foot brake switch detects that braking of a wheel by a braking system is performed, and this braking detecting means brakes the wheel by the braking system. When the execution of is detected, a signal for increasing the fluid pressure supplied between the pulley pieces by a predetermined amount is output from the fluid pressure increasing means toward the fluid pressure adjusting means. Including the case where the tire locks or tends to lock due to sudden braking on a low μ road surface due to an increase in the supply fluid pressure and an increase in the clamping force of the pulley with respect to the belt, the road surface rotation driving force and the braking force It is possible to prevent the slip between the belt and the pulley from being suppressed against the rotational movement (prevention) force of the belt caused by the magnitude of the deceleration of the drive wheels generated from the deviation.

Further, in the control device for a continuously variable transmission according to the second aspect of the present invention, as shown in the basic configuration diagram of FIG. 1, for example, the continuously variable transmission described in Japanese Patent Laid-Open No. 61-105353. In a control device of a machine, a solenoid valve capable of increasing or decreasing the line pressure of the fluid pressure supplied between the pulley pieces separately from the throttle pressure of the fluid pressure supplied between the pulley pieces. And the like are provided as fluid pressure adjusting means. On the other hand, for example, the braking detection means such as a foot brake switch detects that braking of the wheels by the braking system is performed, and the wheel deceleration detection means
For example, the deceleration of the wheel rotation is detected from the differential value of the wheel speed change amount of the drive wheels. The fluid pressure increasing means detects the execution of braking of the wheel by the braking system by the braking detecting means and the wheel rotation deceleration detection value detected by the wheel deceleration detecting means is a predetermined wheel rotation deceleration value. At the above time, for example, the wheel speed of the drive wheel is decelerated beyond the target wheel speed that satisfies the range of the target slip ratio on the low μ road surface, and as a result, the drive wheel is locked or tends to be locked. Then, based on this, a signal for increasing the fluid pressure supplied between the pulley pieces by a predetermined increment is output toward the fluid pressure adjusting means, so that the tire locks due to sudden braking on a low μ road surface. Alternatively, when there is a tendency to lock, the fluid pressure supplied to the pulley increases and the gripping force of the pulley with respect to the belt increases, and as a result, deceleration of the drive wheel occurs due to the difference between the road surface rotational drive force and the braking force. It is possible to prevent the slip between the belt and the pulley from being suppressed and to reduce the energy loss against the rotational movement (blocking) force of the belt caused by the degree of the degree.

Further, as shown in the basic configuration diagram of FIG. 1, particularly in the control device for a continuously variable transmission according to claim 3, the deceleration of the wheel rotation is detected from, for example, the differential value of the wheel speed change amount of the drive wheels. In the control device for a continuously variable transmission according to claim 3 and claim 4 of the present invention, the fluid pressure increasing means causes the wheel rotation detected by the wheel deceleration detecting means. Since the predetermined increase amount of the fluid pressure supplied between the pulley pieces is set according to the deceleration detection value, the inertia torque for sliding the belt increases, that is, the rotational movement (blocking) given to the belt. A belt clamping force sufficient to resist the force can be applied between the pulley pieces, and the energy loss can be further reduced.

Further, in the control device for a continuously variable transmission according to claim 5 of the present invention, as shown in the basic configuration diagram of FIG. Braking fluid pressure detecting means for detecting the fluid pressure of the braking fluid pressure detecting means, and the fluid pressure increasing means is provided between the pulley pieces according to the braking fluid pressure detection value detected by the braking fluid pressure detecting means. Since the predetermined increase amount of the fluid pressure to be set is set, it is possible to apply the belt clamping force corresponding to the rotational deceleration of the drive wheels generated by the deviation between the road surface rotational drive force and the braking force between the pulley pieces. Energy loss can be further reduced.

[0022]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS A first embodiment in which the control device for a continuously variable transmission according to the present invention is applied to an actual vehicle will be described with reference to FIGS. The basic vehicle structure of the first embodiment is described in Japanese Patent Application Laid-Open No. 61-105353 previously proposed by the present applicant, including a continuously variable hydraulic control circuit, a microcomputer as a controller, etc., which will be described later. The control device of the continuously variable transmission described above is the same as or almost the same, and it is explained that the same parts are the same in the description parts of each structure, and then the detailed description will be given with reference to the publication. Sometimes omitted. In this embodiment, it is assumed that the driving wheels that are rotationally driven by the rotational driving force of the engine (engine) are front left and right wheels, that is, a so-called FF (front engine front drive) vehicle. The braking force of each wheel is supplied by a fluid pressure wheel cylinder provided on the wheel, and the working fluid pressure to each wheel cylinder is the master cylinder pressure in the master cylinder connected to the brake pedal. It shall be divided equally. Also,
Basically, it is assumed that the anti-skid control device as described above is not installed.

FIG. 2 shows a power transmission mechanism of a continuously variable transmission. This continuously variable transmission is a fluid coupling 1.
2, forward / reverse switching mechanism 15, V-belt type continuously variable transmission mechanism 2
9, the differential device 56, etc., and rotates the output shaft 10a of the engine 10 at a predetermined gear ratio and rotation direction.
6 and 68. This continuously variable transmission has a fluid coupling 12 (lockup oil chamber 12
a, a pump impeller 12b, a turbine liner 12c, a lockup clutch 12d, and the like), a rotary shaft 1
3, drive shaft 14, forward-reverse switching mechanism 15, drive pulley 16
(Fixed cone member 18, drive pulley cylinder chamber 20 (chamber 2
0a, chamber 20b), movable cone member 22, groove 22a, etc.), planetary gear mechanism 17 (comprising sun gear 19, pinion gear 21, pinion gear 23, pinion carrier 25, internal gear 27, etc.), V-belt 24, driven A pulley 26 (including a fixed conical member 30, a driven pulley cylinder chamber 32, a movable conical member 34, etc.), a driven shaft 28, a forward clutch 40, a drive gear 46, an idler gear 48,
Reverse brake 50, idler shaft 52, pinion gear 5
4, a final gear 44, a differential device 56 (consisting of a pinion gear 58, a pinion gear 60, a side gear 62, a side gear 64, etc.), an output shaft 66, an output shaft 68, etc., but a detailed description thereof will be given. Omit it. Regarding the detailed structure of the parts whose description is omitted, the above-mentioned Japanese Patent Laid-Open No. 61-105353 previously proposed by the present applicant was proposed.
See the publication. The pressure receiving area of the cylinder chamber 32 of the driven pulley 26 is set to about 1/2 of the pressure receiving area of each of the chambers 20a and 20b of the cylinder chamber 20 of the drive pulley 16. Three
A line pressure as a common working hydraulic pressure is supplied to 2 from a hydraulic control device to be described later, and a working hydraulic pressure controlled by the hydraulic control device is supplied to each of the chambers 20a and 20b of the cylinder chamber 20 of the drive pulley 16. When the width of the V-shaped pulley groove of the drive pulley 16 is changed to be narrower and the contact position radius between the V belt 24 and the drive pulley 16 is changed and controlled, the rotational driving force applied to the V belt 24 from the engine 10 is changed. On the contrary, the driven pulley 2 is sandwiched so that the V belt 24 and the driven pulley 26 do not slip and are in inverse proportion to the amount of change in the width of the V-shaped pulley groove of the drive pulley 16.
The width of the V-shaped groove 6 is changed to be narrower to control the contact position radius between the driven pulley 26 and the V belt 24, thereby achieving a desired pulley ratio between the two pulleys 16 and 26. Is configured to be a gear ratio between the input and output of the continuously variable transmission.

FIG. 3 shows a hydraulic control device for a continuously variable transmission according to this embodiment. This hydraulic control device includes an oil pump 101,
Line pressure regulating valve 102, manual valve 104, shift control valve 106, regulating pressure switching valve 108, step motor 11
0, shift operation mechanism 112, throttle valve 114, constant pressure pressure regulating valve 116, solenoid valve 118, coupling pressure regulating valve 1
20, a lock-up control valve 122, etc., which are connected to each other as shown in the figure, and further, a forward clutch 40, a reverse break 50, a fluid coupling 12, a lock-up oil chamber 12a, a drive pulley. The cylinder chamber 20 and the driven pulley cylinder chamber 32 are also connected as shown. A detailed description of these valves and the like is the same as or substantially the same as that described in the above-mentioned Japanese Patent Laid-Open No. 61-105353, and therefore the description thereof will be omitted here as a reference. At the switching stop position of the spool 136 of the manual valve 104, a so-called 2 range is interposed between the L range and the D range, and the spool 136 is stopped at a total of 6 positions. There is no specific change in the operating hydraulic pressure in the hydraulic control device related to the increase in the stop position, and the calculation processing in the microcomputer described later is slightly different.
Each reference numeral in FIG. 3 indicates the following member. Pinion gear 110a, reservoir tank 130, strainer 13
1, oil passage 132, relief valve 133, valve hole 134, ports 134a to 134e, spool 136, land 136
a to 136b, oil passage 138, one-way orifice 139,
Oil passage 140, oil passage 142, one-way orifice 143, valve hole 146, ports 146a to 146g, spool 14
8, lands 148a to 148e, sleeve 150, spring 152, spring 154, gear ratio transmission member 15
8, oil passage 164, oil passage 165, orifice 166, orifice 170, valve hole 172, ports 172a to 172
e, spool 174, lands 174a to 174c, spring 175, oil passage 176, orifice 177, lever 178, oil passage 179, pin 181, rod 182, lands 182a and 182b, rack 182c, pin 183,
Pin 185, valve hole 186, ports 186a-186d,
Oil passage 188, oil passage 189, oil passage 190, valve hole 192, ports 192a to 192g, spool 194, land 19
4a, 194e, negative pressure diaphragm 198, orifice 199, orifice 202, orifice 203, valve hole 2
04, ports 204a to 204e, spool 206, lands 206a and 206b, spring 208, oil passage 20.
9, filter 211, orifice 216, port 22
2, solenoid 224, plunger 224a, spring 225, valve hole 230, ports 230a to 230e, spool 232, lands 232a and 232b, spring 234, oil passage 235, orifice 236, valve hole 240,
Ports 240a to 240h, spool 242, land 2
42a to 242e, oil passage 243, oil passage 245, orifice 246, orifice 247, orifice 248, orifice 279, choke type throttle valve 250, relief valve 251, pressure holding valve 252, choke type throttle valve 253, oil passage 254, cooler 256. , A cooler pressure holding valve 258, an orifice 259, and a changeover detection switch 278.

Further, in the present embodiment, the throttle pressure acting as the throttle valve 114 can be increased / decreased together with the throttle negative pressure diaphragm 198, so that the line pressure supplied to the drive pulley cylinder chamber 20 and the driven pulley cylinder chamber 32 as a common hydraulic pressure. Oil passage 2 to change and control the
A solenoid valve 270 for adjusting the throttle pressure of the oil passage 140 by using the pilot pressure of 09 as a base pressure is added. In addition to this, there is no oil passage communicating the third port 186c of the adjustment pressure switching valve 108 and the oil passage 209, and the third port 186c of the adjustment pressure switching valve 108 is closed. The solenoid valve 118 will be referred to as a lockup solenoid valve 118, and the solenoid valve 270 will be referred to as a line pressure adjusting solenoid valve 270, based on the application described later.

The line pressure adjusting solenoid valve 270 is an oil passage that connects the seventh port 192g of the throttle valve 114, the third port 204c of the constant pressure adjusting valve 116 and the eighth port 240h of the lockup control valve 122. Of the choke type throttle valve 253, it is provided at an end of an oil passage 189 extending from the throttle valve 114 side. This line pressure adjusting solenoid valve 270 is used for the oil passage 1
The amount of hydraulic oil 89 discharged from the port 272 is adjusted by the plunger 274a. The plunger 274a is biased by a spring 275 in a direction to close the port 272, and the spring 275 is opened in a direction to open the port 272 according to the magnetic field strength formed by the excitation of the solenoid 274.
Moved against the urging force of. Since the solenoid 274 is duty ratio controlled by the shift control device (microcomputer) 300 described later,
The amount of hydraulic oil discharged from the oil passage 189 increases in proportion to the energization amount, so that the pilot pressure of the oil passage 189, that is, the oil passage 209 is controlled so as to be inversely proportional to the energization amount to the solenoid 274. become.

This action will be described as the throttle pressure characteristic of the entire throttle valve 114 with reference to FIG. As described above, the engine is operated by the negative pressure diaphragm 198.
As the intake negative pressure of 10 increases, the output oil pressure to the oil passage 140 decreases. Therefore, the throttle pressure from the throttle valve 114 to the oil passage 140 is controlled by the control oil pressure of the negative pressure diaphragm 198 and the line pressure adjusting solenoid valve 170. It may be considered that it is the sum of the control oil pressure and the control oil pressure. Here, the duty ratio of the drive signal to the line pressure adjusting solenoid valve solenoid 274 is 10
When it becomes 0%, the pilot pressure in the oil passage 209 disappears, so the throttle pressure from the throttle valve 114 is only the control oil pressure by the negative pressure diaphragm 198, and the line pressure added to this is the line pressure in the normal traveling state and the creep traveling state. From the oil passage 132 to the shift control valve 10
It is supplied to the drive pulley cylinder chamber 20 and the driven pulley cylinder chamber 32 via 6. On the other hand, when the duty ratio of the drive signal to the line pressure adjusting solenoid valve 274 becomes 0%, the pilot pressure in the oil passage 109 becomes maximum, so the throttle pressure from the throttle valve 114 depends on this maximum pilot pressure. It is possible to increase the sum of the maximum output hydraulic pressure and the control hydraulic pressure by the negative pressure diaphragm 198. From the above, the line pressure of the oil passage 132 supplied to the drive pulley cylinder chamber 20 and the driven pulley cylinder chamber 32 is determined by the solenoid 27 of the line pressure adjusting solenoid valve 270.
Therefore, the lock-up solenoid valve 118 described in the above-mentioned Japanese Patent Laid-Open No. 61-105353 has the lock-up control by the fluid coupling 12 in the calculation process of the gear ratio control described later. It will be understood that it is a solenoid valve that exclusively controls only.

FIG. 5 shows an electronic control unit (microcomputer) 300 for controlling the operations of the step motor 110 and the solenoid 224. The microcomputer 300 includes an input interface 311,
Reference pulse generator 312, central processing unit (CPU)
313, a read only memory (ROM) 314, a random access memory (RAM) 315 and an output interface 316, which are address bus 31.
9 and the data bus 320. The microcomputer includes an engine speed sensor 3
01, vehicle speed sensor 302, throttle opening sensor 30
3, shift position switch 304, turbine rotation speed sensor 305, engine cooling water temperature sensor 306, brake sensor 307, changeover detection switch 298, left drive wheel speed (ie front left wheel speed) sensor 402, right drive wheel speed (ie front right wheel) Speed) sensor 404 and master cylinder pressure sensor 40
The signal from 6 directly or the waveform shapers 308, 309, 3
22, 412 and 414, and AD converters 310 and 41
6 is input, while the amplifier 317 and the signal line 3 are input.
A signal is output to the step motor 110 via 17a to 317d, and a signal is also output to the lock-up solenoid valve solenoid 224 and the line pressure adjusting solenoid valve solenoid 274. Since it is the same as or almost the same as that described in Japanese Laid-Open Patent Publication No. 61-105353, it will be omitted as a reference. In addition, as described above, the shift position includes the L range and the D range, which are not described in the publication.
Since two ranges are newly provided between the ranges,
From the shift position switch 304, a total of six position signals including these two ranges are input to the microcomputer 300. Further, the left drive wheel speed sensor 402 and the right drive wheel speed sensor 404 respectively output sine wave output signals corresponding to the drive wheel speeds, and the drive wheel speeds are respectively transmitted via the waveform shapers 412 and 414. Wheel speed detection values V _WL and V _WR (which are _collectively referred to as respective wheel speeds, and the sign in that case is also _referred to as V _Wj . Therefore, j corresponds to L or R). It is read into the microcomputer 300 via the input interface 311. Further, the master cylinder pressure sensor 40
6 outputs an output signal according to the magnitude of the master cylinder pressure of the braking system. This output signal is the A / D
Via the converter 416, as a master cylinder pressure detection value (also simply referred to as master cylinder pressure) P corresponding to the increase in the master cylinder pressure that has increased above atmospheric pressure, via the input interface 311 the microcomputer 30
Read within 0. The brake sensor 307 is equivalent to a so-called foot brake switch, and when the output signal from the brake sensor 307 is in the ON state,
When the brake pedal is depressed and the output signal from the brake sensor 307 is in the OFF state, the brake pedal is not depressed. Further, as described above, the control signal S output via the output interface 316 is directed toward the line pressure adjusting solenoid valve solenoid 274.
_SOL constitutes the duty ratio Dt _SOL set by the microcomputer 300. In addition, the sensors 402, 404 and 406 and the waveform shaper 4 described above are added.
12, 414 and A / D converter 416 are, in this embodiment,
Since each detected value via them is not used directly, it does not need to be an essential requirement for the configuration.

Then, the microcomputer 300
Thus, the gear ratio control of the continuously variable transmission is executed according to the reference calculation process shown in the flowchart of FIG. The basic logic system of this arithmetic processing is described in JP-A-61-105.
Although it is almost the same as the one described in Japanese Patent No. 353, the shift pattern corresponding to the two ranges is added as the searched shift pattern in the relationship that the two ranges are added to the shift position. Further, since the lock-up solenoid valve 118 and the line pressure adjusting solenoid valve 270 are divided into two parts, the solenoid drive signal that has been both the lock-up control and the line pressure adjusting control is the solenoid valve 118.
Alternatively, they are individually output to the solenoid valves 270. Specifically, the solenoid to be controlled by the duty ratio when the shift position is in the D, 2, L, R range is the solenoid valve 224 for lockup, and the control by the duty ratio when the shift position is in the P, N range. The target solenoid is the solenoid valve solenoid 274 for line pressure adjustment.
Becomes

To briefly describe the reference calculation process of the gear ratio control, the calculation process of FIG. 6 is executed by a timer interrupt every predetermined time (ΔT), and first, step 502.
At step 504, the shift position is read from the shift position switch 304, and if it is determined at step 504 that the shift position is in the D, 2, L, R range, then the process proceeds to step 508, and if not, step 506. Move to. In step 508, the throttle opening TH is read based on the signal from the throttle opening sensor 303, then in step 510, the vehicle speed V is read based on the signal from the vehicle speed sensor 302, and then in step 512, the engine speed sensor 301 is read. The engine rotation speed N _E is read based on the signal from the turbine rotation speed N _E , and then the turbine rotation speed N _t is read based on the signal from the turbine rotation speed sensor 305 in step 514. Next, the routine proceeds to step 516, where the engine speed N _E
And calculates the rotational difference N _D of the turbine rotation speed N _t, then in step 518, searches the lock-up vehicle speed V _ON and drop up off vehicle speed V _OFF according to the control map stored in advance.

Next, the process proceeds to step 520, and if the drop-up flag LUF is set, step 5
44, otherwise, to step 522. In step 544, if the vehicle speed V is lower than the lock-up off vehicle speed V _OFF , the process proceeds to step 540, and if not, the process proceeds to step 546. On the other hand, if it is determined in step 522 that the vehicle speed V is higher than the lockup vehicle speed V _{ON, the process} proceeds to step 524, and if not, the process proceeds to step 540. In Step 524, the first target value Nm ₁ is subtracted from the rotation deviation N _D to calculate the rotation target value deviation e, and then in Step 526, the rotation target value deviation e is calculated from the control map stored in advance. First
Search for the feedback gain G ₁ of, then step 5
At 28, if the rotation deviation N _D is smaller than the control system switching threshold value N _{0, the} process proceeds to step 530, and if not, the process proceeds to step 538. In step 530, a minute predetermined value α is added to the previous duty ratio of the lockup solenoid valve solenoid 224 to set the current duty ratio of the lockup solenoid valve solenoid 224.
Next, if it is determined in step 532 that the current duty ratio of the lockup solenoid valve solenoid 224 is smaller than 100%, the process proceeds to step 602, and if not, the process proceeds to step 534. Step 5
34, the current duty ratio of the lockup solenoid valve solenoid 224 is corrected to 100%, and then step 53 is performed.
In step 6, the drop-up flag LUF is set, and the flow advances to step 602. On the other hand, in step 538, the current duty ratio is calculated based on an arithmetic expression having the rotation target value deviation e and the first feedback gain G ₁ as variables, and the process proceeds to step 602. On the other hand, in step 540, the lockup solenoid valve solenoid 224 is operated.
This time, the duty ratio is set to 0%, and then step 54
In step 2, the lockup flag LUF is calculated, and the process proceeds to step 602. In step 546, the current duty ratio of the lockup solenoid valve solenoid 224 is set to 100%, and the process proceeds to step 602.

If it is determined in step 602 that the vehicle speed V is smaller than the gear ratio control start threshold value V _{0, the process} proceeds to step 604, and if not, step 62.
Go to 4. In step 604, the throttle opening T
If it is determined that H is smaller than the idle determination threshold TH _{0, the process} proceeds to step 610, and if not, the process proceeds to step 606. In step 606, the current duty ratio of the lockup solenoid valve solenoid 224 is set to 0%, then in step 608, the target pulse P _D to the step motor 110 is set to the maximum gear ratio pulse P ₁ and step 630. Move to. On the other hand, in step 506, the solenoid valve solenoid 274 for line pressure adjustment is used.
This time the duty ratio is set to 0% and the above step 63 is performed.
Move to 0.

On the other hand, in step 624, when the shift position is in the D range, the process shifts to step 626 to search the gear ratio corresponding to the vehicle speed V and the throttle opening TH from the gear shift pattern corresponding to the D range, and Go to step 630. If the shift position is not in the D range, the process proceeds to step 639, and if the shift position is in the two range, the process proceeds to step 640 and the shift pattern corresponding to the two ranges corresponds to the vehicle speed V and the throttle opening TH. The gear ratio to be used is searched and the process proceeds to step 630. When the shift position is not in the 2 range, the process proceeds to step 642, and when the shift position is in the L range, the process proceeds to step 628 and the vehicle speed V is changed from the shift pattern corresponding to the L range.
Then, the gear ratio corresponding to the throttle opening TH is searched, and the process proceeds to step 630. If the shift position is not in the L range, the process proceeds to step 644, the gear ratio corresponding to the vehicle speed V and the throttle opening TH is searched from the gear shift pattern corresponding to the shift position R range, and the process proceeds to step 630.

On the other hand, if it is determined in step 610 that the changeover detection switch 298 is on, step 6
12, otherwise, to step 620. In step 612, the second target value Nm ₂ is subtracted from the rotation deviation N _D to calculate the rotation target value deviation e, and then in step 614, the rotation target value deviation e is calculated from the control map stored in advance. The second feedback gain G ₂ is searched, and then, in step 616, the current duty ratio of the lockup solenoid valve solenoid 224 is set to the rotation target value deviation e and the second feedback gain G 2.
It is calculated based on an arithmetic expression having ₂ as a variable, and then at step 618, the current pulse number P _A to the step motor 110 is calculated.
Is set to "0" and the process proceeds to step 636. on the other hand,
If it is determined in step 630 that the current pulse number P _A is equal to the target pulse number P _D , step 636 is performed.
Move to. When it is determined in step 630 that the current pulse number P _A is smaller than the target pulse number P _D , the process proceeds to step 632, the step motor drive signal is moved in the upshift direction, and then in step 634. Add "1" to the current pulse number P _A to obtain a new current pulse number P
After being updated and stored as _A , the process proceeds to step 636. On the other hand, if it is determined in step 630 that the current pulse number P _A is larger than the target pulse number P _D , the process moves to step 620 to move the step motor drive signal in the downshift direction, and then step 622. Then, subtract 1 from the current pulse number P _A to obtain a new current pulse number P _A
After updating and storing as, the process proceeds to step 636.

In step 636, the step motor drive signal is output, and then in step 638, the solenoid valve solenoid drive signal is output, and then the process returns to the main program. In the present embodiment, steps 626 and 62 excluding the R range equivalent shift pattern search in step 644 described above.
It can be considered that the speed change pattern searched in 8, 640 is such that the speed ratio of the continuously variable transmission is set according to the speed change pattern as shown in FIG. That is, the gear ratio in each gear shift pattern is uniquely set by searching according to these variables on the control map having the vehicle speed V and the throttle opening TH as variables. Assuming that FIG. 6 is a comprehensive control map of a shift pattern having the vehicle speed V as the horizontal axis, the engine rotation speed Ne as the vertical axis, and the throttle opening TH as parameters, a straight line with a constant inclination passing through the origin is a gear ratio. Can be considered to be constant. For example, the straight line with the largest inclination in the entire region of the shift pattern has the largest reduction ratio of the entire vehicle, that is, the maximum gear ratio C _Hi , and conversely, the straight line with the smallest inclination is
It can be considered that the reduction ratio of the entire vehicle is the smallest, that is, the D range minimum transmission ratio C _DLO , and the reduction ratio of the entire vehicle having a larger inclination than the D range minimum transmission ratio is the 2 range minimum transmission ratio C _2LO. . Therefore, specifically, the gear shift pattern of the L range is fixed to the maximum gear ratio C _Hi regardless of the vehicle speed V and the throttle opening TH, and the gear shift pattern of the two ranges is the maximum gear ratio C _Hi and the two range minimum. Gear ratio C
_In the region between _2LO and V, it becomes a control curve consisting of a time-dependent locus of the gear ratio set according to the vehicle speed V and the throttle opening TH, and the shift pattern of the D range is the maximum gear ratio C.
_It will be a control curve consisting of a time-dependent locus of the gear ratio set in accordance with the vehicle speed V and the throttle opening TH in the region between _Hi and the D range minimum gear ratio C _DLO . In the region where the vehicle speed V is smaller than the gear ratio control start threshold value V ₀ , the gear ratio (that is, the gear shift pattern) is fixed to the maximum gear ratio C _Hi regardless of the range of each shift position. That is,
It can be considered that the gear ratio control start threshold value V ₀ is the control upper limit value of the creep state that occurs in the vehicle equipped with the automatic transmission. Here, the throttle opening TH at the gear ratio control start threshold V ₀ at the maximum gear ratio C _Hi is defined as the gear ratio control start threshold TH _1, and the gear ratio control start throttle opening threshold TH ₁ has two ranges. The vehicle speed V that becomes the minimum gear ratio C _2LO is the two-range minimum gear ratio vehicle speed V ₂₁ , and the vehicle speed V that becomes the D range minimum gear ratio C _{DLO at the} gear ratio control start throttle opening threshold TH ₁ is the D range minimum gear ratio vehicle speed V. _It is defined as _D1 and these minimum range gear ratio vehicle speeds V ₂₁ and V _D1 are also simply described as minimum range gear ratio vehicle speeds V _j1 .

Then, the state of the line pressure supplied from the oil passage 132 to the drive pulley cylinder chamber 20 and the driven pulley cylinder chamber 32 due to such a gear ratio control is shown in FIG.
Follow the instructions below. At this time, it can be considered that the rotational driving force to be transmitted (which can be easily understood by considering the rotational driving force from the consumed engine) becomes smaller as the gear ratio becomes smaller. Therefore, the engine rotational speed Ne and the turbine rotational speed N
In the calculation process of FIG. 6 for keeping the deviation from t constant, the hydraulic pressure supplied to the fluid coupling 12 may be decreased as the gear ratio of the continuously variable transmission becomes smaller. The solenoid valve 118 increases the amount of hydraulic oil discharged from the oil passage 190 when the gear ratio decreases, and when the gear ratio decreases, the oil passage 190 and the oil passage 188 increase.
The reference line pressure of the oil passage 132 also decreases due to the decrease in the operating hydraulic pressure of. Therefore, the line pressure of the oil passage 132 becomes smaller as the gear ratio of the continuously variable transmission becomes smaller, regardless of the throttle pressure of the throttle valve 114. to this,
The line pressure of the oil passage 132 with respect to the gear ratio is obtained by adding the throttle pressure characteristic shown in FIG. Therefore, irrespective of the control oil pressure by the negative pressure diaphragm 198, if the duty ratio of the line pressure adjusting solenoid valve 270 is changed and controlled, the entire line pressure can be increased or decreased, whereby the drive pulley cylinder chamber 20 and the driven valve are driven. By changing and controlling the hydraulic pressure supplied to the pulley cylinder chamber 32, it is possible to change and control the belt clamping force by them.

Next, in a vehicle equipped with the continuously variable transmission and the shift control device thereof as described above, the slippage between the belt and the pulley, which occurs especially on a low μ road surface such as an icy snow road surface or a wet tile road surface. The problem and the basic principle of this embodiment for solving the problem will be briefly described. On such a low μ road surface, since the friction coefficient state between the tire and the road surface is small, the actual slip ratio of the wheel during the sudden braking such as the sudden braking is the target slip ratio capable of ensuring the steering effect and the braking distance, that is, 10 to 30. The slip ratio range of% is easily exceeded, and the grip force of the tire itself is further reduced.
In the control device for a continuously variable transmission disclosed in Japanese Patent Laid-Open No. 4-254054, the gear ratio of the continuously variable transmission is fixed to a gear ratio immediately before the start of sudden braking, that is, a relatively small gear ratio. When the driver releases the brake pedal and the wheels start to rotate, the gear ratio fixed control is released, so the gear ratio of the continuously variable transmission is changed to a large gear ratio. It will be.

However, in the continuously variable transmission which does not have the power transmission direction regulating means such as the one-way clutch on the output side, the input (that is, road surface rotational driving force) transmission from the road surface necessary for rotating the wheels is transmitted. The system is provided with an inertial weight that the rotary system of the continuously variable transmission has at its end, and the deceleration of the wheel speed of the drive wheels generated from the deviation between the road surface rotational driving force and the braking force is the fixed gear ratio control. The inertia torque is applied to the continuously variable transmission that has been set. This increasing inertial torque acts as a large rotational movement (blocking) force for blocking the rotational movement of the belt, and moreover, during braking where the throttle opening is considered to be reduced, the above-mentioned Japanese Patent Laid-Open No. SHO 61-61 is used. Since it is considered that the fluid pressure applied to the output side pulley (secondary pulley, driven pulley 26 of the present embodiment) is also reduced in the control device for a continuously variable transmission described in Japanese Patent Laid-Open No. 105353, The rotational movement force applied to the belt becomes larger than the belt holding force of the secondary pulley, the belt slides between the pulley pieces, and as a result, the durability of the belt deteriorates.

Therefore, in the present embodiment, in order to prevent the slip between the belt and the pulley from being suppressed, the simplest way is to start the braking by depressing the brake pedal. (Pressure) is increased to the maximum value to increase the belt clamping force of the pulley. Although this method still has some points to be solved such as energy loss, there is an advantage that the calculation load executed by the microcomputer 300 can be made extremely small.

FIG. 9 shows a calculation process for changing and controlling the supply line pressure to the drive pulley 20 and the driven pulley 32 in an actual vehicle based on the above-described principle of the invention. This arithmetic processing is executed by a timer interrupt at every predetermined time (ΔT) in the microcomputer 30 which is the control device of the continuously variable transmission, and the control signal S _SOL calculated and set here is the line pressure adjusting solenoid valve. The solenoid valve 274 for line pressure adjustment, which is output to the solenoid 274, changes and controls the movement amount of the plunger 274a according to the duty ratio Dt _SOL configured in the control signal S _SOL . Here, when the P range or the N range is selected as the shift position, it is supposed that the vehicle is not originally traveling for sudden braking, and hunting with the control signal by the arithmetic processing of FIG. 6 is prevented. Therefore, the line pressure is not adjusted and controlled. (The problem is when the N range is selected during braking.
We do not recommend selecting a range).

In the arithmetic processing of FIG. 9, first, in step S
At 1, the shift position is read based on the signal from the shift position switch 304. Then step S2
Then, it is determined whether the shift position read in step S1 is the P range or the N range,
The selected shift position is in the P range or N
If it is in the range, the process returns to the main program, and if not, the process proceeds to step S3.

In step S3, it is determined whether or not the output signal of the brake sensor 307 is in the OFF state. If the output signal of the brake sensor 307 is in the OFF state, the process proceeds to step S4, and if not. Then, the process proceeds to step S5. In step S4, the duty ratio Dt _SOL of the line pressure adjusting solenoid valve solenoid 274 is set to 100%, and then the process proceeds to step S6.

On the other hand, in step S5, the duty ratio Dt of the line pressure adjusting solenoid valve solenoid 274 is set.
_After setting _SOL to 0%, the process _proceeds to step S6. In step S6, a control signal S _SOL corresponding to each duty ratio Dt _SOL set in step S4 or step S5 is formed, and this control signal S _SOL is output to the line pressure adjusting solenoid valve solenoid 274. Then return to the main program.

Next, the operation of the arithmetic processing of FIG. 9 will be described. In this calculation process, if the brake pedal is depressed with the shift position selected in any of the D, 2, L, and R ranges in steps S1 to S3, the same step S5, regardless of the road surface μ. The control signal S _SOL whose duty ratio is set to 0% in S6 is output to the solenoid 274 of the line pressure adjusting solenoid valve 270.
At this time, unless the special pedal operation is performed, it is considered that the throttle opening is reduced and the control oil pressure by the negative pressure diaphragm 198 in FIGS. 4 and 8 is reduced. Therefore, the output hydraulic pressure of the line pressure adjusting solenoid valve 270 to which the control signal S _SOL is input, that is, the throttle pressure of the oil passage 140 is increased to the maximum except the control hydraulic pressure of the negative pressure diaphragm 198 shown in FIG. As a result, the line pressure in the oil passage 132 shown in FIG. 8 is also increased to the maximum value in the gear ratio. As a result, the hydraulic pressure supplied to the drive pulley cylinder chamber 20 and the driven pulley cylinder chamber 32 is the same as the negative pressure diaphragm 1 at the gear ratio.
The pressure is increased to the maximum value excluding the control oil pressure of 98, and therefore the belt clamping force between the pulleys 16 and 26 becomes extremely large. On the other hand, when the brake pedal is released, the control signal with the duty ratio set to 100% is output to the solenoid 274 of the line pressure adjusting solenoid valve 270, so that regardless of the throttle opening state. The output hydraulic pressure of the line pressure adjusting solenoid valve 270 is reduced to atmospheric pressure or almost atmospheric pressure, and the throttle pressure shown in FIG. 4 and the line pressure shown in FIG. The line pressure supplied to the pulley cylinder chamber 20 and the driven pulley cylinder chamber 32 causes each belt to generate a sufficient belt holding force sufficient to transmit the rotational driving force of the engine, so that slippage occurs between the pulley and the belt. do not do.

This is applied to ice and snow roads, wet tile roads, etc.
On a low μ road surface, the stepless method disclosed in the above-mentioned JP-A-4-254054
The transmission control unit controls the gear ratio to be fixed during sudden braking.
The time chart of Fig. 10
Consider. In the figure, during the low μ road surface traveling, at time t
₁To start sudden braking by pressing the brake pedal abruptly
Then, after that, the wheel speed V of the drive wheels_WjIs decelerating rapidly
t₃And target wheel speed V^* _WjWill soon fall below time t_FiveComplete
It is assumed that all locks have occurred. This and
The vehicle speed V is such that the grip force between the wheels and the road surface is extremely small.
In such a situation, there is almost no deceleration, and therefore the time t_FiveThen the car body
Only the glider has transitioned to a gliding state. This simulator
The wheel speed of the driven drive wheel is set to the wheel speed V on the low μ road surface._WjLFWhen
In order to facilitate understanding, the vehicle speed V is expressed as
Pseudo vehicle speed adopted by Chiskid control device, that is, maximum vehicle
Wheel speed V_WHiIt was adopted. Further, the arithmetic processing shown in FIG.
The normal control line pressure by P_LnIn this example,
Then, the control line pressure by the arithmetic processing of FIG. 9 is set to P_LRepresented by
It was Also, the wheel speed V of the drive wheels_WjLODeceleration of
The belt against the rotational movement (blocking) force of the belt.
Generate the clamping force required for clamping at each pulley
Is converted to the required line pressure, and the required line pressure P
_LOExpressed as Note that the shift position is the normal
It is the D range or engine braking range
It is assumed that 2 or L range is selected. By the way
For dry asphalt roads and concrete roads
Wheel speed V on high μ road surface_WjHFIs the target wheel speed V^* _WjTo
The actual vehicle speed V will not decrease and the deceleration slope will be larger.
It is thought that it will become a good thing.

In the calculation process of FIG. 9, the brake sensor signal is turned on at the same time when the brake pedal is depressed, so the control signal S whose duty ratio is set to 0% in steps S5 and S6 through steps S1 to S3. _SOL is output toward the solenoid 274 of the line pressure adjusting solenoid valve 270. Therefore, as described above, after the time t ₁ when the depression of the brake pedal is started, the drive pulley cylinder chamber 2
0 and the line pressure P _L supplied to the driven pulley cylinder chamber 32 increases up to the maximum line pressure P _LMAX excluding the control hydraulic pressure of the negative pressure diaphragm 298 shown in FIG. 8, and thereafter, until the brake pedal is released. The maximum line pressure P _LMAX is maintained. On the other hand, as described above, the inertia torque to the continuously variable transmission, which is brought about by the magnitude of the wheel deceleration, is within the range where the wheel speed V _WjLF is the target wheel speed V ^* _Wj or is slightly smaller than that range. It increases according to the magnitude of the deceleration, and then gradually decreases as the wheel speed V _WjLF further decreases. Therefore, the required line pressure P _LO required to prevent the slippage between the pulley and the belt in accordance with the change in the inertia torque gradually increases after time t _{1 as} shown in FIG. It gradually decreases over the locking time t ₅ . Here, in the normal line pressure P _Ln in which the calculation process of FIG. 9 is not executed, the required line pressure P _LO that increases and decreases in this way exceeds the normal line pressure P _Ln in the period from time t ₂ to time t _3. I will end up. That is, this means that the rotational movement (blocking) force to the belt at the time t _{2 to} t ₃ exceeds the belt clamping force of each pulley, which causes slippage between the pulley and the belt. It is thought that. On the other hand, at the line pressure P _L for which the calculation process of FIG. 9 has been executed, the required line pressure P _LO does not exceed this line pressure P _L , so it is considered that slippage does not occur between the pulley and the belt. Therefore, the durability of the belt can be improved.

In the above embodiment, the operation of this embodiment has been described only for the vehicle not equipped with the anti-skid control device. However, the same problem occurs in a vehicle equipped with the anti-skid control device. This may occur, and can be solved almost in the same way by executing the arithmetic processing of FIG. 9 in combination with the operation signal of the anti-skid control device. Further, in this case, if the slip ratio is set as the determination requirement in accordance with the determination of the wheel acceleration / deceleration, the configuration becomes more reliable.

Further, the control signal S _SOL and its duty ratio Dt _SOL output toward the solenoid 274 of the line pressure adjusting solenoid valve 270 result in that the belt clamping force of each pulley is increased to increase the belt clamping force. Since the fluid pressure supplied to the motor, that is, the line pressure is increased, the belt clamping force increase amount of the pulley, that is, the line pressure increase amount can be said to depend on the duty ratio Dt _SOL of the control signal S _SOL. . On the other hand, the amount of increase in the inertia torque of the continuously variable transmission, that is, the rotational movement (blocking) force to the belt depends on the deceleration of the drive wheels, and the deceleration of the drive wheels is equal to the road surface rotational drive force. Since it depends on the deviation from the braking force, for example, the deceleration of the driving wheels and the braking fluid pressure related to the braking force are used as parameters, and the duty ratio Dt _{SOL of the} control signal is determined according to the magnitude of these parameters. That is, it is the amount of increase in the line pressure and the amount of increase in the belt clamping force of the pulley. However, if the duty Dt _SOL is increased, it is necessary and sufficient to prevent slippage between the pulley and the belt. It is possible to reduce the above-mentioned energy loss at the same time while obtaining the clamping force.

As described above, the present embodiment is the first aspect of the present invention.
It is considered that the control device for a continuously variable transmission according to the present invention is embodied, and step S3 of the calculation process of FIG. 9 and the brake sensor 307 correspond to the braking detection means of the present invention, and the calculation process of FIG. The whole corresponds to the fluid pressure increasing means, and the line pressure adjusting solenoid valve 270 corresponds to the fluid pressure adjusting means.

Next, a second embodiment in which the control device for a continuously variable transmission according to the present invention is applied to an actual vehicle will be described with reference to FIGS. 11 and 12. The basic vehicle structure of the second embodiment is described in Japanese Patent Application Laid-Open No. 61-105353 previously proposed by the present applicant, including a continuously variable hydraulic control circuit, a microcomputer as a controller, etc., which will be described later. The control device of the continuously variable transmission described above is the same as or almost the same, and it is explained that the same parts are the same in the description parts of each structure, and then the detailed description will be given with reference to the publication. Sometimes omitted.

First, the power transmission mechanism of the continuously variable transmission of the present embodiment is the same as or almost the same as the power transmission mechanism of the continuously variable transmission of FIG. 2 corresponding to the first embodiment, and at the same time, this is Since it is the same as or almost the same as the one described in Japanese Patent Laid-Open No. 61-105353, detailed description thereof will be omitted here. The hydraulic control system for the continuously variable transmission of this embodiment is the same as or substantially the same as the hydraulic control system for the continuously variable transmission of FIG. 3 corresponding to the first embodiment. Since it is the same as or almost the same as that described in Japanese Patent Laid-Open No. 105353, detailed description will be omitted here. The line pressure adjusting solenoid valve 270 and the lockup solenoid valve 118 additionally described in the first embodiment are similar to or substantially the same as those in the first embodiment, and particularly the line pressure adjusting solenoid valve. The operation of 270 is the same as or substantially the same as that shown in FIG. 4 and FIG. 8 including the normal line pressure control by the calculation processing of the gear ratio control described later, and therefore the detailed description thereof will be omitted.

Further, the microcomputer constituting a part of the controller corresponding to the shift control device of the present embodiment is the same as or substantially the same as the electronic control device (microcomputer) 300 of FIG. 5 corresponding to the first embodiment. At the same time, since this is the same as or substantially the same as that described in Japanese Patent Laid-Open No. 61-105353, detailed description thereof will be omitted here.

The sensors 402 and 40 additionally described above
4, 406, the waveform shapers 412, 414, and the A / D converter 416 are indispensable requirements for directly using each detected value via them in this embodiment. Further, the gear ratio control of the normal continuously variable transmission executed by the microcomputer 300 is executed in the same or almost the same manner as in the first embodiment according to the reference calculation process shown in the flowchart of FIG. JP-A-61-105
Since it is the same as or almost the same as that described in Japanese Patent No. 353, the detailed description is omitted here. The above 2
The range shift pattern search is the same as in the first embodiment. Therefore, the gear ratio control by the gear shifting pattern of each range except the R range is performed by the first gear ratio control shown in FIG.
Since it is the same as or almost the same as the embodiment, its detailed description is omitted. Further, the correlation between the gear ratio related to the normal gear ratio control and the line pressure in the oil passage 132, and the action of the throttle valve 114 including the line pressure adjusting solenoid valve 270 are equivalent to those shown in FIG. The detailed explanation is omitted because they are almost the same.

Next, in a vehicle equipped with the continuously variable transmission and its speed change control device as described above, the wheel speed (particularly the drive wheel) generated especially on a low μ road surface such as an icy snow road surface or a wet tile road surface. The problem of speed) and the basic principle of this embodiment for solving the problem will be briefly described. Based on the above-mentioned basic principle, in the present embodiment, for example, when the braking of the wheel is executed by depressing the brake pedal, the line pressure of the oil passage 132 to the drive pulley cylinder chamber 20 and the driven pulley cylinder chamber 32 is increased. To increase the belt clamping force of each pulley, and thereby to prevent slippage between the pulley and the belt against the belt rotational movement (blocking) force due to the inertia torque of the continuously variable transmission caused by the deceleration of the drive wheels. Although the suppression is prevented, it is energy that increases the fluid pressure supplied to the pulley at any braking, that is, increases the belt clamping force despite the fact that the belt rotational movement (blocking) force is not increased. It can be said that it is a loss. Therefore,
The deceleration of the drive wheel is detected, and the supply line pressure to the pulley is increased only when the drive wheel deceleration is larger than a predetermined value set in advance. This is equivalent to determining that the tire is locked or has a tendency to lock due to sudden braking on a low μ road surface, and energy can be reliably prevented while preventing slippage of the belt that occurs during braking on the low μ road surface. The loss can be reduced.

Here, if this control is executed every predetermined sampling time ΔT by the calculation processing of the timer interrupt, etc., for example, each is read from the left driving wheel speed sensor 402 and the right driving wheel speed sensor 404. Wheel speed detection value V
Obtained by dividing a deviation between _Wj the previous processing wheel speeds previous value V _Wj0 during this sampling time ΔT is, the deceleration alpha _Wj next to each wheel, which is given by equation (1) below.

Α _Wj = (V _Wj0 −V _Wj ) / ΔT (1) Since the wheel deceleration calculated by the above formula 1 includes the wheel acceleration to the front of the vehicle, both of them are used here. It is _assumed that the wheel acceleration / deceleration α _Wj is calculated by including the following, and the wheel acceleration toward the front of the vehicle is represented by a positive value and the wheel deceleration toward the rear of the vehicle is represented by a negative value. Therefore, the expression that the wheel deceleration is greater than a certain predetermined value means here that the wheel acceleration / deceleration α _Wj is a certain predetermined value α.
_Described as less than _W0 (negative value). However, although each wheel acceleration / deceleration α _Wj may be compared with a predetermined value α _W0 , it is better to compare a smaller wheel acceleration / deceleration α _Wj (larger as a deceleration) with a predetermined value α _W0. It is desirable in the sense that it is judged during road surface braking. Therefore, in the actual calculation process, each wheel of the acceleration alpha _Wj, whichever is smaller minimum acceleration alpha _W by select low wheel acceleration of
Calculated as, the minimum acceleration alpha _W the predetermined value alpha _W0
Decided to compare with. The predetermined value α _W0 is low μ
When a large braking force is applied to the wheel on the road surface, and as a result, the wheel speed is lower than the wheel speed that satisfies the target slip ratio, that is, when the wheel tends to lock, the deceleration of the wheel, that is, It is set to a value below which the negative wheel acceleration / deceleration is set.On a normal high μ road surface, the value is set to such a value that the negative wheel acceleration / deceleration does not fall below this predetermined value even when sudden braking is performed. There is.

Further, since the inertia torque for sliding the belt, that is, the rotational movement (blocking) force given to the belt depends on the magnitude of the deceleration of the driving wheel speed, When the predetermined increase amount of the supply line pressure is set according to the magnitude of the deceleration of the drive wheel speed, the energy loss can be further reduced while the clamping force of the belt by the pulley is increased as much as necessary. Further, since the deceleration of the driving wheel speed depends on the deviation between the road surface rotational driving force and the braking force, specifically, this braking force is detected by detecting the braking fluid pressure of the braking system. Is detected and the predetermined increase amount of the supply line pressure to the pulley is set according to the magnitude of the braking force, the energy loss is increased while the belt clamping force by the pulley is increased as necessary. It can be made even smaller. Here, the absolute value | α _W | of the minimum wheel acceleration / deceleration is used as the deceleration of the drive wheel speed, and the master cylinder pressure P is used as the braking force, and the increase amount ΔP of the line pressure is calculated according to the following two equations. _L
Is calculated and set.

ΔP _L = K ₁ · | α _W | + K ₂ · P (2) where K ₁ and K ₂ are preset positive weighting control coefficients, and K ₁ + K ₂ = It is assumed that k ₁ (k ₁ is a constant) is satisfied. Then, this line pressure increase amount ΔP _L
The duty ratio Dt _SOL to the solenoid valve solenoid 274 for line pressure adjustment that achieves
_The control signal S _SOL forming _SOL is set by searching the line pressure characteristic diagram shown in FIG. 4 or FIG. 8 on a map and output to the line pressure adjusting solenoid valve solenoid 274. At this time, there is a concern that the duty ratio Dt _SOL calculated and set from the above equations 1 and 2 will be smaller than “0%”, but this is conversely within the range of the magnitude of the deceleration of the drive wheel that is assumed. Is used to set how much the line pressure must be increased, and the duty ratio D to the line pressure adjusting solenoid valve solenoid 274 is set up to the set maximum value of the line pressure.
It would be avoided by allowing change control only in t _SOL .

FIG. 11 shows a calculation process for changing and controlling the supply line pressure to the drive pulley 20 and the driven pulley 32 in an actual vehicle based on the above-mentioned principle of the invention. This arithmetic processing is executed by a timer interrupt at every predetermined time (ΔT) in the microcomputer 30 which is the control device of the continuously variable transmission, and the control signal S _SOL calculated and set here is the line pressure adjusting solenoid valve. is output to the solenoid 274, the line pressure regulating solenoid valve solenoid 274 is the control signal S _SOL configured to duty ratio Dt _SOL
The movement amount of the plunger 274a is changed and controlled accordingly.
Here, when the P range or the N range is selected as the shift position, it is supposed that the vehicle is not originally traveling for sudden braking, and hunting with the control signal by the arithmetic processing of FIG. 6 is prevented. Therefore, the line pressure is not adjusted and controlled (the problem is when the N range is selected during braking, but it is not recommended to select the N range during running including braking). ). Further, the control flag F during the arithmetic processing is in the set state of "1" to indicate that the line pressure increase control is being performed, and in the reset state of "0" to indicate that the normal line pressure control is being performed. In addition, each time the control flag or the timer is set or reset in the figure, the storage update to the RAM 315 is executed at the same time.

In the arithmetic processing of FIG. 11, first, in step S11, the respective wheel speed detection values (also simply referred to as wheel speed) V _Wj are read from the left drive wheel speed sensor 402 and the right drive wheel speed sensor 404, respectively. . Next, in step S12,
The shift position is read based on the signal from the shift position switch 304.

Next, in step S13, it is determined whether the shift position read in step S12 is in the P range or N range, and the selected shift position is in the P range or N range. If there is, the process proceeds to step S14, and if not, the process proceeds to step S15. In step S15, it is determined whether or not the output signal of the brake sensor 307 is in the OFF state, and the output signal of the brake sensor 307 is O.
If it is in the FF state, the process proceeds to step S16, and if not, the process proceeds to step S17.

In step S17, it is determined whether or not the control flag F is in the set state of "1". If the control flag F is in the set state of "1", step S17 is executed.
If not, the process proceeds to step S19. In step S19, the latest wheel speed previous value V _Wj0 stored in the RAM 315 of the microcomputer 300 is read, and then the _{process proceeds} to step S20.

In the step S20, the step S
Wheel speed V read in 11_WjAnd read in step S19
Previous wheel speed V_Wj0Using and, according to the above equation 1,
Each wheel acceleration / deceleration α_WjAnd then move to step S21
To go. In step S21, in step S20
Acceleration / deceleration α of each wheel calculated in _WjOf these, small wheels
Deceleration α_WjSelect with the select low
Speed α_WjIs the minimum wheel acceleration / deceleration α_WSet as
Then, the process proceeds to step S22.

In the step S22, the step S
Minimum wheel acceleration alpha _W calculated set at 21, determines whether the absolute value is large and is greater than the predetermined value alpha _W0 set in advance a negative value, the minimum wheel acceleration alpha _W is predetermined If it is larger than the value α _{W0, the process} proceeds to step S16, and if not, the process proceeds to step S23.

In step 23, the control flag F is set to "1", and then the process proceeds to step S24. In step S24, the master cylinder pressure P is read from the master cylinder pressure sensor 406, and then the process proceeds to step S25. In the step S25, the absolute value of the minimum wheel acceleration / deceleration | α _W | calculated in the step S21 and the master cylinder pressure P read in the step S24 are used to calculate the line pressure increase amount ΔP _L according to the equation (2). Then, the process proceeds to step 26.

In the step S26, the step S
The duty ratio Dt of the solenoid valve solenoid 274 for line pressure adjustment that achieves the line pressure increase amount ΔP _L calculated in step 25.
_After calculating _SOL (%), the process proceeds to step S27.
In step S27, the duty ratio Dt _SOL (%) of the solenoid valve solenoid 274 for line pressure adjustment calculated in step S27 is updated and stored in the RAM 315, and then the process proceeds to step S18.

In step S18, the latest duty ratio Dt _SOL (%) of the solenoid valve solenoid 274 for line pressure adjustment, which is updated and stored in the RAM 315, is read, and then the process proceeds to step S28. On the other hand, the step S1
In step 6, the control flag F is reset to "0" and then the process proceeds to step S29. In step S29, the duty ratio Dt _{SOL of the} line pressure adjusting solenoid valve solenoid 274 is set.
Is set to 100% and then the process proceeds to step S28.

In the step S28, the step S
18 or a control signal S _SOL corresponding to each duty ratio Dt _SOL calculated and set in step S29 is formed, and this control signal S _SOL is used for the line pressure adjusting solenoid valve solenoid 27.
After outputting to 4, the process proceeds to step S30. On the other hand, in step S14, the control flag F is reset to "0" and then the process proceeds to step S30.

Then, in step S30, each wheel speed V _Wj read in step S11 is updated and stored in the RAM ₃₁₅ as each wheel speed previous value V _Wj0 , and then the process returns to the main program. Next, the operation of the arithmetic processing of FIG. 11 will be described. Now, on a high μ road surface such as a dry asphalt road surface or concrete road surface where a sufficient friction coefficient state is maintained with the tire, the accelerator pedal is depressed and the vehicle normally runs at a constant speed or acceleration. I assume that the state. It is assumed that the shift position is maintained in the D range suitable for normal traveling. In this state, the brake pedal is not depressed, and as a result, each wheel speed detection value V _Wj is read in step S11 at every predetermined sampling time when the arithmetic processing of FIG. 11 is executed, and steps S12 and S13 are executed. The process proceeds from step S15 to step S16, the control flag F is reset to "0" at step S16, and the procedure goes to step S29 to set the duty ratio Dt _SOL of the line pressure adjusting solenoid valve solenoid 274 to 100%. In step S28, this 100% duty ratio Dt _SOL
The control signal S _SOL set to the above is formed and output to the line pressure adjusting solenoid valve solenoid 274, and in step S30, the wheel speeds V _Wj are _set to the previous wheel speed values V _Wj0.
Is updated and stored in the RAM 315.

[0070] Thus, the control signal S _SOL enter the line pressure regulating solenoid valve 270 of the set duty ratio Dt _SOL to the 100%, the pilot pressure working oil in the oil passage 209 as shown in FIG. 4 Since it is discharged, the oil passage 14
The throttle pressure of 0 is only the output hydraulic pressure of the negative pressure diaphragm 198 in the figure, and this throttle pressure is added to the line pressure of the oil passage 132 and supplied to the drive pulley cylinder chamber 20 and the driven pulley cylinder chamber 32. Therefore, in each of these pulleys, the belt clamping force necessary and sufficient to transmit the rotational driving force of the engine at the gear ratio is maintained.

Next, from this state, it is assumed that the accelerator pedal is released and the brake pedal is depressed large and fast without changing the shift position on the high μ road surface, and so-called sudden braking and sudden braking are performed. On such a high μ road surface, even if a large braking force is applied to each wheel by sudden braking, the target wheel speed at which the wheel speed satisfies the target slip ratio is increased due to the large friction coefficient between the road surface and the tire. It is unlikely that it will fall below. Therefore, at each sampling time when the calculation process of FIG.
Read _Wj , go through steps S12 and S13, and then step S
Since the brake sensor signal is ON in step S15, the process proceeds to step S17. Since the control flag F is still in the reset state of "0", the process proceeds to step S19. Each calculated in step S19 and step S20. of the wheel acceleration alpha _Wj, the minimum wheel acceleration alpha _W selected in the step S21 becomes a considerably large one negative value of the absolute value, the minimum wheel acceleration alpha _W is in a high μ road surface Since the absolute value does not fall below the predetermined value α _W0 , which is a large and negative value in the sudden braking, the process proceeds from step S22 to step S16 to maintain the control flag F in the reset state of "0", and then the duty ratio Dt _SOL the line pressure adjusting solenoid valve solenoid 274 is set to 100% in step S29, in step S28, set to the 100% duty ratio Dt _SOL The determined control signal S _SOL is _generated and output to the line pressure adjusting solenoid valve solenoid 274, and further, in step S30, each wheel speed V _Wj is updated and stored in the RAM ₃₁₅ as each wheel speed previous value V _Wj0. The flow repeats as long as the brake pedal is depressed.

Therefore, in the line pressure adjusting solenoid valve 270 to which the control signal S _SOL having the duty ratio Dt _SOL set to 100% is input, the pilot pressure hydraulic oil in the oil passage 209 is supplied as shown in FIG. Since the output oil pressure of the negative pressure diaphragm 198 is reduced because the gas is discharged and the accelerator pedal is released, the oil passage 14
The throttle pressure of 0 is relatively small, and the line pressure of the oil passage 132 to which the throttle pressure is applied is also small. Therefore, the pulleys are driven by the drive pulley cylinder chamber 20 and the driven pulley cylinder chamber 32 to which the small line pressure is supplied. It is considered that the belt gripping force of is reduced. However, during sudden braking on this high μ road surface, the wheel speeds of the drive wheels do not show a tendency to lock or lock significantly below the target wheel speed, so the absolute value of the wheel acceleration / deceleration | α _Wj | Because the inertia torque of the continuously variable transmission is increased to an extremely large value, the rotational movement (blocking) force of the belt is not increased by this, and slippage between the belt and the pulley does not occur. There is no need to increase the clamping force. From such a meaning, when braking on a high μ road surface,
Even if it is during sudden braking, not performing unnecessary increase control of the line pressure reduces energy loss.

On the other hand, on a low μ road surface such as the above-mentioned snow and snow road surface or a wet tile road surface, since the friction coefficient state between the tire and the road surface is small, the wheels show a tendency to lock during braking including the above-mentioned sudden braking. The wheel speed easily falls below the target wheel speed that satisfies the target slip ratio. On such a low μ road surface, it is considered that the actual wheel speed is very difficult to increase with respect to the road surface, that is, the input of the vehicle body speed to rotate the wheels as described above. ,
When the wheel speed starts to fall below the target wheel speed, the wheel speed continues to decelerate and locks after a relatively short time unless at least the accelerator pedal is depressed to positively increase the wheel speed by the rotational driving force of the engine. Alternatively, it is considered that the state immediately before locking is reached. Therefore, this low μ
At the time of braking on the road surface, the wheel speed detection value V _Wj is read in step S11 in the first arithmetic processing of FIG. 8 in which the wheel speed starts to rapidly _decrease , and then steps 12 and S1.
After 3, the process moves from step S15 to step S17. In this step S17, since the control flag F is still in the reset state of "0", the process proceeds to step S19. Then, this step S19 and step S20
Each wheel acceleration / deceleration rate α _Wj calculated in step 1 will be a considerably large negative value. Therefore, the minimum wheel acceleration / deceleration α calculated and set in step S21 of the arithmetic processing of FIG.
_W also has a considerably large negative absolute value. Then, when the minimum wheel acceleration / deceleration α _W falls below the predetermined value α _W0, which has a large absolute value and is set to a negative value, in step S22 of the arithmetic processing of FIG. 11, the process _proceeds to step S23. In step S23, the control flag F is set to "1", and then in step S24, the master cylinder pressure P is set.
Is read. Next, at step S25, the line pressure increase amount ΔP _L is calculated according to the above equation 2 using the absolute value of the minimum wheel acceleration / deceleration | α _W | and the master cylinder pressure P.
Then, the duty ratio Dt _{SOL of the} solenoid valve solenoid 274 for line pressure adjustment that achieves this line pressure increase amount ΔP _L
Is calculated in step S26, and the duty ratio Dt
_SOL is updated and stored in the RAM 315 in step S27. Then, the latest duty ratio Dt _SOL is read in at step S18, in step S28, the output to the said line pressure adjusting solenoid valve solenoid 274 to form a control signal S _SOL in accordance with the duty ratio Dt _SOL Further, in step S30, each wheel speed V _Wj is updated and stored in the RAM ₃₁₅ as each wheel speed previous value V _Wj0 .

Then, after the same braking, as shown in FIG.
For each sampling time at which the arithmetic processing of 1 is executed, the process proceeds from step S15 to step S17. Since the control flag F is in the set state of "1" at this step S17, the process proceeds to step S18, and this step S18. The latest duty ratio Dt _SOL updated and stored at this time is read in, and in step S28, the duty ratio Dt _SOL concerned is read.
Forming a control signal S _{SO L} corresponding to output toward the line pressure regulating solenoid valve solenoid 274, further the RAM315 update store each wheel speed V _Wj as the wheel speeds previous value V _Wj0 in step S30 The flow is repeated.

During this period, the throttle valve 114 shown in FIG. 4 has a throttle pressure set as described above, although the output hydraulic pressure of the negative pressure diaphragm 198 is small because the accelerator pedal is released. the control signal S _SOL enter the line pressure regulating solenoid valve 270 ratio Dt _SOL, since the emissions of the pilot pressure hydraulic fluid of the oil passage 209 is reduced in proportion to the duty ratio Dt _SOL, the oil passage 140 The throttle pressure increases in inverse proportion to the decrease in the hydraulic oil discharge amount, and as a result, the line pressure increase amount Δ calculated in step S25 of the calculation process of FIG.
Increase by P _L. Therefore, the line pressure of the oil passage 132 to which the throttle pressure thus increased is also increased, and the belt of each pulley is sandwiched by the drive pulley cylinder chamber 20 and the driven pulley cylinder chamber 32 to which the large line pressure is supplied. Power increases. As a result, when braking on this low μ road surface, if the wheel speed of the driving wheel is much lower than the target wheel speed and tends to lock, the absolute value of the wheel acceleration / deceleration | α _Wj | The inertial torque of the transmission is increased, which increases the rotational movement (prevention) force of the belt to cause slippage between the belt and the pulley, but the slippage occurs due to the increase in the belt clamping force of each pulley. Can be prevented. Also, the rotational movement (blocking) force of this belt _{depends on} the absolute value of the wheel acceleration / deceleration of the drive wheel | α _Wj | (substantially negative acceleration, that is, only the deceleration). Since the absolute value | α _Wj | of the road surface rotation driving force and the braking force, that is, the deviation between the master cylinder pressure P,
The line pressure increase amount ΔP _L calculated by the formula gives a necessary and sufficient belt holding force to each pulley, and the energy loss is small compared to always increasing the line pressure to the maximum value. Become.

On the other hand, when the depression of the brake pedal is released, the brake sensor signal is turned off, so that the process proceeds from step S15 to step S16. At step S16, the control flag F is reset to "0", and then, In step S29, the line pressure adjusting solenoid valve solenoid 2
The duty ratio Dt _SOL 74 were set to 100%, in step S28, the output to the said line pressure adjusting solenoid valve solenoid 274 to form a control signal S _SOL in accordance with the duty ratio Dt _SOL, further steps In S30, the wheel speeds V _Wj are _set as the previous wheel speed values V _Wj0 in the RAM 3
15 is updated and stored. Therefore, when the depression of the brake pedal is released, the line pressure returns to the normal line pressure.

Then, the operation during the line pressure increase control
On a low μ road surface such as ice and snow road surface or wet tile road surface,
Control device for continuously variable transmission disclosed in Japanese Patent Laid-Open No. 4-254405
In situations where the gear ratio is fixedly controlled during sudden braking,
The time chart of FIG. 12 will be considered. Same figure
Then, at the time t while traveling on the low μ road surface,₁₁Rapid break in
Start sudden braking by depressing the pedal, then drive
Wheel speed V_WjRapidly decelerates at time t₁₄Target wheel speed
V^* _WjWill soon fall below time t₁₆In a completely locked state
It is supposed to be in a fallen state. At this time, the vehicle speed V is
Almost decelerates when the grip between the vehicle and the road surface is extremely small.
Therefore, the time t₁₆Then only the car body is gliding
It is transitioning to a state. Of this simulated drive wheel
Wheel speed is low μ Wheel speed on road surface V_WjLFIs expressed as
Said anti-skid control device to facilitate understanding
Pseudo vehicle speed adopted in the vehicle, that is, maximum wheel speed V_WHiAdopted
It was In addition, a normal control routine based on the arithmetic processing shown in FIG.
In pressure is P_LnThis embodiment, that is, the calculation of FIG.
Control line pressure by processing P_LExpressed as In addition, the drive
Wheel speed V of driving wheel_WjLORotation of the belt affected by deceleration of
Necessary to clamp the belt against the movement (blocking) force.
The pulley required to generate the required clamping force at each pulley.
Converted to the required line pressure P_LOExpressed as
The shift position is the D range which is the normal driving range.
2 or L range which is J or engine braking range
It shall be selected. By the way, dried asph
Wheel speed on high μ roads such as alt roads and concrete roads
V_WjHFIs the target wheel speed V^* _WjNever falls below
Considering that the deceleration slope of the actual vehicle speed V will be even greater.
available.

In the arithmetic processing of FIG. 11, the brake pedal is
The brake sensor signal is turned on when the pedal is depressed.
However, it is calculated in steps S11, S19 to S21.
Minimum wheel acceleration / deceleration α_WAt time t₁₂Start with a predetermined value α_W0
Became smaller than. Therefore, in step S25,
Absolute value of this minimum wheel acceleration / deceleration | α_W｜ and Master Siri
Line pressure increase amount ΔP obtained from the bonder pressure P_LIs calculated
In step 26, the line pressure increase amount ΔP_LReach
Duty ratio Dt_SOLIs calculated in the step
In S27, this duty ratio Dt_SOLIs updated to RAM315
Newly memorized. Then, the time t₁₂After that, step S
18, at S28, the duty ratio Dt_{SO L}According to
Signal S_SOLIs a solenoid valve for adjusting the line pressure 270
It is output to the terminal 274. Therefore, this time t₁₂Since
After that, the drive pulley cylinder chamber 20 and the driven pulley cylinder
Line pressure P supplied to chamber 32_LIs the negative shown in FIG.
Excluding the control oil pressure of the pressure diaphragm 298, the line pressure
Increase ΔP_LOnly increase and then the brake pedal depression
Until the line is released, the line pressure P_LMaintained at. one
On the other hand, as mentioned above, the stepless change caused by the magnitude of wheel deceleration
The inertia torque to the speed machine is the wheel speed V_WjLFIs the target wheel speed V^*
_WjIn the range of or slightly smaller than
It increases according to the magnitude of the wheel deceleration, and then
Wheel speed V_WjLFBecomes smaller and decreases gradually. Therefore,
Depending on the fluctuation of the inertia torque, the pulley-belt
Required line pressure P necessary to prevent slippage_LOFigure 1
As shown in 0, time t₁₁After that, the number of wheels gradually increased
Time t to completely lock₁₆Gradually decrease over time
become. Here, in the normal case where the arithmetic processing of FIG. 9 is not executed
Line pressure P_LnThen, the required line pressure P that increases and decreases in this way
_LOAt time t₁₃From time t ₁₅Time to the relevant normal line
Pressure P_LnWill exceed. That is, this time t₁₃~ T₁₅
The rotational movement (blocking) force to the belt is
It means that it exceeds the belt clamping force of
This may cause slippage between the pulley and belt.
Be done. On the other hand, it is necessary to perform the arithmetic processing of FIG.
In pressure P_LThen, the required line pressure P_LOIs this line pressure
P_LOf the pulley and the belt
It is thought that the belt will not occur, thus improving the durability of the belt.
You can go up. At this time, the arithmetic processing of FIG.
Line pressure P increased by_LIs the maximum line pressure P_LMAXYo
Less than the time t₁₂At time t₁₁Slower than
The maximum line pressure at the same time as the brake pedal is depressed.
Line pressure P_LMAXCompared to line pressure control which increases to
The energy loss is reduced by the area difference between the two shown in FIG.
It will be possible.

In the above embodiment, the operation of this embodiment has been described only for the vehicle not equipped with the anti-skid control device. However, even in the vehicle equipped with the anti-skid control device, the same problem as described above occurs. This may occur, and can be solved almost in the same way by executing the arithmetic processing of FIG. 11 in combination with the operation signal of the anti-skid control device. Further, in this case, if the slip ratio is set as the determination requirement in accordance with the determination of the wheel acceleration / deceleration, the configuration becomes more reliable.

From the above, it is considered that the present embodiment implements the control device for a continuously variable transmission according to all claims 2, 4 and 5 of the present invention, and step S1 of the arithmetic processing of FIG.
5 and the brake sensor 307 correspond to the braking detecting means of the present invention, and hereinafter, similarly, step S of the arithmetic processing of FIG.
11, S19 to S21 and each wheel speed sensor 402, 40
Reference numeral 4 corresponds to a wheel deceleration detecting means, which corresponds to step S24 of the arithmetic processing of FIG. 11 and the master cylinder pressure sensor 406.
Corresponds to the working fluid pressure detecting means, the whole arithmetic processing of FIG. 11 corresponds to the fluid pressure increasing means, and the line pressure adjusting solenoid valve 270 corresponds to the fluid pressure adjusting means.

The above-mentioned embodiments are based on the control device for a continuously variable transmission described in Japanese Patent Application Laid-Open No. 61-105353 previously proposed by the present applicant. Needless to say, it can be widely applied to belt type continuously variable transmissions other than this. Further, in each of the above-mentioned embodiments, only the gear ratio control controller constructed by a microcomputer has been described in detail. However, the invention is not limited to this, and may be configured by combining an electronic circuit such as an arithmetic circuit. Needless to say.

[0082]

As described above, according to the control apparatus for a continuously variable transmission of the present invention, when the braking detecting means detects the execution of braking of the wheel by the braking system, the fluid pressure increasing means causes the fluid to flow. The signal output to the pressure adjusting means increases the fluid pressure supplied between the pulley pieces by a predetermined increase amount, and the clamping force of the pulley with respect to the belt increases, so that the road surface rotational driving force and the braking force are increased. It is possible to prevent the slip between the belt and the pulley from being suppressed against the rotational movement (prevention) force of the belt caused by the magnitude of the deceleration of the drive wheels generated from the deviation.

Further, when the deceleration of the wheel rotation detected by the wheel deceleration detecting means is equal to or more than a predetermined value set in advance, the gripping force increase control of the pulley is performed, so that the sharpness on the low μ road surface is controlled. It is possible to determine the magnitude of the deceleration of the driving wheels that is generated from the difference between the road surface rotational driving force and the braking force, which is considered to occur when the tire locks or tends to lock due to braking, and only that much Energy loss can be reduced.

Further, since the predetermined increase amount of the fluid pressure supplied between the pulley pieces is set according to the wheel rotation deceleration detection value detected by the wheel deceleration detecting means, it is necessary. The energy loss can be further reduced while applying a sufficient belt holding force to the pulley.
Further, since the predetermined increase amount of the fluid pressure supplied between the pulley pieces is set according to the braking fluid pressure detection value detected by the braking fluid pressure detecting means, it is necessary and sufficient. Energy loss can be further reduced while applying a belt holding force to the pulley.

[Brief description of drawings]

FIG. 1 is a basic configuration diagram of a control device for a continuously variable transmission according to the present invention.

FIG. 2 is a configuration diagram showing an example of a power transmission mechanism of a continuously variable transmission.

FIG. 3 is a configuration diagram showing an example of a hydraulic control device for a continuously variable transmission.

FIG. 4 is a characteristic explanatory diagram of throttle pressure by the throttle valve of FIG.

FIG. 5 is a configuration diagram showing an example of a controller corresponding to a gear ratio control device for a continuously variable transmission.

6 is a flowchart showing an example of a calculation process of a gear ratio control of a normal continuously variable transmission executed by the controller of FIG.

7 is an explanatory diagram of a shift pattern by the calculation process of FIG.

8 is a characteristic explanatory diagram of a supply line pressure to each pulley by the calculation processing of FIG.

FIG. 9 is a flowchart of a calculation process for performing a line pressure increase control showing a first embodiment of a control device for a continuously variable transmission according to the present invention.

FIG. 10 is a time chart for explaining the operation of line pressure increase control by the arithmetic processing of FIG.

FIG. 11 is a flowchart of a calculation process for performing line pressure increase control showing a second embodiment of the control device for a continuously variable transmission according to the present invention.

12 is a time chart for explaining the operation of line pressure increase control by the arithmetic processing of FIG.

[Explanation of symbols]

16 is a drive pulley (primary pulley) 18 is a fixed conical plate (fixed pulley piece) 20 is a drive pulley cylinder chamber 22 is a movable conical plate (movable pulley piece) 24 is a V belt 26 is a driven pulley (secondary pulley) 29 is stepless Transmission mechanism (continuously variable transmission) 30 is a fixed conical plate (fixed pulley piece) 32 is a driven pulley cylinder chamber 34 is a movable conical plate (movable pulley piece) 110 is a step motor 118 is a lock-up solenoid valve 224 is a solenoid 270 Line pressure adjusting solenoid valve (fluid pressure adjusting means) 274 is a solenoid 300 is a microcomputer 302 is a vehicle speed sensor (vehicle speed detecting means) 303 is a throttle opening sensor 304 is a shift position switch 402 is a left drive wheel speed sensor (wheel rotation reduction) Speed detection means) 404 is a right drive wheel speed sensor (wheel rotation deceleration detection means) 06 the master cylinder pressure sensor (brake fluid pressure detection means)

Claims (5)

[Claims] 1. A fluid capable of changing and controlling the clamping force of a belt clamped between the paired pulley pieces by increasing / decreasing the fluid pressure supplied between each pair of relatively movable pulley pieces. A control device for a continuously variable transmission including pressure adjusting means, comprising: braking detecting means for detecting execution of braking of a wheel by a braking system; and when the braking detecting means detects execution of braking. , A fluid pressure increasing means for outputting a signal for increasing the fluid pressure supplied between the pulley pieces by a predetermined increment toward the fluid pressure adjusting means, the control device for the continuously variable transmission. 2. A fluid capable of changing and controlling the holding force of a belt held between the pair of pulley pieces by increasing / decreasing the fluid pressure supplied between each pair of relatively movable pulley pieces. A control device for a continuously variable transmission equipped with a pressure adjusting means, the braking detecting means for detecting execution of braking of a wheel by a braking system, and the wheel deceleration detecting means for detecting deceleration of wheel rotation. And when the braking detection means detects the execution of braking and the wheel rotation deceleration detection value detected by the wheel deceleration detection means is equal to or greater than a predetermined wheel rotation deceleration value, it is supplied between the pulley pieces. A control device for a continuously variable transmission, comprising: a fluid pressure increasing means for outputting a signal for increasing the fluid pressure by a predetermined increment toward the fluid pressure adjusting means. 3. A wheel deceleration detecting means for detecting a wheel rotation deceleration, wherein the fluid pressure increasing means is responsive to the wheel rotation deceleration detection value detected by the wheel deceleration detecting means. The control device for a continuously variable transmission according to claim 1, wherein a predetermined increase amount of the fluid pressure supplied to one side is set. 4. The fluid pressure increasing means, according to the wheel rotation deceleration detection value detected by the wheel deceleration detecting means,
The control device for the continuously variable transmission according to claim 2, wherein a predetermined increase amount of the fluid pressure supplied between the pulley pieces is set. 5. A braking fluid pressure detecting means for detecting a fluid pressure for generating a braking force of a braking system, wherein the fluid pressure increasing means detects the braking fluid detected by the braking fluid pressure detecting means. 5. The control device for a continuously variable transmission according to claim 1, wherein a predetermined increase amount of the fluid pressure supplied between the pulley pieces is set according to a pressure detection value.