JPH06109121A - Hydraulic pressure control device of continuously variable transmission - Google Patents

Abstract

PURPOSE:To provide a hydraulic pressure control device of a belt type continuously variable transmission which is capable of rapidly and surely returning the speed change ratio to LOW without causing any inconveniences such as the shortage of the working oil at various parts of the hydraulic system when the brake is applied for deceleration. CONSTITUTION:The target value of the line pressure of the hydraulic mechanism FS is set to the minimum pressure where no belt slip is caused between a V-belt 33 and pulleys 31, 3 by means of a control unit CU in the normal operation, the pump loss is reduced, and the fuel consumption is improved. On the other hand, the target value of the line pressure is set to the pressure value where the supplyable flow rate is equal to the flow rate to require the speed change by means of the control unit CU when the brake is applied for deceleration, and the speed-changing speed is maximized while the sufficient amount of working oil being supplied to various parts except the line pressure control requiring the working oil, leading to smooth re-acceleration after deceleration braking and smooth re-starting.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a hydraulic control device for a belt type continuously variable transmission.

[0002]

2. Description of the Related Art Generally, an automobile is provided with a transmission that changes the output torque of an engine in accordance with the driving state. , 4 forward and 1 reverse) are often used. However, in the multi-stage transmission, since it is not possible to set a gear ratio other than several preset gear ratios, there is a problem that the gear ratio most suitable for the operating state of the vehicle cannot be obtained. There is a problem that a gear shift shock occurs when changing gears.

Therefore, in recent years, development of a continuously variable transmission capable of setting a gear ratio to an arbitrary value within a predetermined range has been advanced, and as one of such continuously variable transmissions, a belt type continuously variable transmission ( Hereinafter, this is referred to as CVT) has been proposed. In such a CVT, usually, a primary pulley (driving pulley) and a secondary pulley (driven pulley) whose pulley diameters can be changed respectively, and a belt wound around both pulleys are provided. The gear ratio can be set to an arbitrary value by changing.

In such a CVT, a hydraulic piston is provided for each of the primary pulley and the secondary pulley in order to change the pulley diameter (hereinafter, these are referred to as a primary piston and a secondary piston, respectively). A hydraulic mechanism for supplying hydraulic pressure is provided, and the pulley diameter of each pulley is changed by changing the amount of hydraulic oil or hydraulic pressure supplied to each hydraulic piston.

In such a conventional CVT, normally, the gear ratio is controlled by the amount or hydraulic pressure of hydraulic oil supplied to the primary piston (hereinafter referred to as primary pressure), and the belt tension is hydraulic pressure supplied to the secondary piston.
(Hereinafter, this is referred to as secondary pressure). In a normal CVT, the line pressure of the hydraulic mechanism is generally introduced into the secondary piston, so the line pressure is the secondary pressure. Here, the line pressure that affects the belt tension
The (secondary pressure) is set to a minimum pressure within a range that does not cause the belt to slip. This is because if the line pressure is increased more than necessary, power loss will increase and fuel consumption performance will deteriorate.

By the way, in the belt slip, the force attracting the belt and the pulley in the rotational direction (hereinafter referred to as the belt driving force) is the maximum frictional force between the belt and the pulley (hereinafter referred to as the belt maximum force). It is generated when it becomes larger than (friction force). Here, the belt driving force is basically determined by the input torque to the CVT and the gear ratio of the CVT. On the other hand, the maximum belt friction force is almost proportional to the contact surface pressure between the belt and the pulley, that is, the belt tension. Therefore, the minimum belt tension that does not cause belt slip depends on the input torque and the gear ratio.

Therefore, in the CVT, the line pressure
(Secondary pressure) needs to be set according to the input torque and the gear ratio. However, since it is generally difficult to detect the input torque to the CVT, in the conventional CVT, the line pressure is usually the throttle opening (engine load) which is almost proportional to the input torque to the CVT.
And the gear ratio of the CVT.

As described above, the conventional CV in which the line pressure is set based on the throttle opening and the gear ratio.
At T, the throttle opening becomes almost 0 during deceleration, so the line pressure is set to a very low value.
(For example, 6 to 10 kg / cm ² ). Even in this case, as described above, since the line pressure sufficient to prevent the occurrence of belt slip is secured, no particular problem occurs during normal deceleration. However, during deceleration with sudden braking (hereinafter, this is referred to as deceleration braking), the following problems occur.

That is, during deceleration braking, the vehicle speed suddenly decreases or stops, but in this case, in order to smoothly re-accelerate or restart the vehicle, the CVT gear ratio is set to the maximum value (hereinafter, this state is LOW. Need to be set back to). In the present specification, the “gear ratio” means the torque ratio, that is, (output torque) / (input torque). Here, in order to change the gear ratio to the LOW side, it is necessary to drain the hydraulic oil in the primary piston, but this drain speed depends on the primary pressure almost balanced with the line pressure, and the primary pressure is high. The higher the line pressure, the faster the speed. However,
During such deceleration braking, the line pressure is set to be extremely low as described above, so that the drain speed at the primary piston becomes extremely small.

On the other hand, during deceleration braking (during sudden braking), the drive wheels may be locked early (for example, 2 seconds on a normal road surface). Then, when the drive wheels are locked, the rotation of the secondary pulley is stopped accordingly. However, when the rotation of the secondary pulley is stopped in this way, the gear ratio hardly changes thereafter. In this case, since the drain speed at the primary piston is small as described above, the gear ratio is L when the drive wheels are locked.
I haven't returned to OW. Therefore, there is a problem that the acceleration performance is deteriorated at the time of re-acceleration or restart, and further, there is a problem that a so-called belt squeal occurs and the driver feels uncomfortable.

[0011]

In order to improve this,
A CVT has been proposed in which the line pressure is increased during deceleration braking to increase the drain speed of the primary piston (see, for example, JP-A-61-52457).
However, in the conventional CVT in which the line pressure is increased during deceleration braking as disclosed in Japanese Patent Laid-Open No. 61-52457, for example, the amount of increase in line pressure during deceleration braking is the same as that of the entire hydraulic system. It is not set with due consideration of operating conditions. Therefore, the amount of increase in the line pressure during deceleration braking is not always the optimum value, and there is a problem that the gear ratio cannot return to LOW when the amount of increase in pressure is too small, and conversely when the amount of increase in pressure is too large. In spite of the fact that the gear ratio returns to LOW, there arises a problem that the required amount of hydraulic oil in each part of the hydraulic system increases as the line pressure increases, and a part where the hydraulic oil supply amount becomes insufficient occurs.

The present invention has been made in order to solve the above-mentioned problems of the prior art, and the speed ratio can be promptly and reliably brought about during deceleration braking without causing a problem such as insufficient hydraulic oil in each part of the hydraulic system. It is an object of the present invention to provide a hydraulic control device for a belt type continuously variable transmission that can return the engine to LOW.

[0013]

In order to achieve the above object, as shown in FIG. 1, a first aspect of the present invention is a hydraulic mechanism A for continuously changing a gear ratio by changing a pulley diameter by hydraulic pressure.
In a hydraulic control device for a belt type continuously variable transmission B provided with, a line pressure control means C for setting the line pressure of the hydraulic mechanism A according to the operating state of the vehicle, and a vehicle in a predetermined deceleration state. Deceleration state detection means D for detecting whether or not there is
When the deceleration state detecting means D detects that the vehicle is in the predetermined deceleration state, among the hydraulic fluids discharged from the oil pump E, the hydraulic fluid that can be used for line pressure control is selected. A deceleration line pressure correction control means F for increasing the line pressure to a maximum value that can be set under the condition that the flow rate is not less than the hydraulic oil flow rate required for the shifting operation of the continuously variable transmission B is provided. A hydraulic control device for a continuously variable transmission, which is characterized by being provided.

A second aspect of the present invention is the hydraulic control device for a continuously variable transmission according to the first aspect, wherein the flow rate of the hydraulic oil that can be used for controlling the line pressure is that of the continuously variable transmission B. A map G is provided which shows the maximum value of the line pressure that can be set under the condition that the flow rate is not less than the hydraulic oil flow rate required for the operation, and the map G is provided for deceleration line pressure. A hydraulic control device for a continuously variable transmission, characterized in that the correction control means F reads the maximum value of the line pressure to be set from the map G.

[0015]

EXAMPLES Examples of the present invention will be specifically described below.
As shown in FIG. 2, the power train PT for automobiles is
4-cylinder engine CE including first to fourth cylinders # 1 to # 4
And a hydraulically actuated transmission CT. Here, the engine CE outputs the engine torque to the crankshaft 1
Output is made to the transmission CT side via the (engine output shaft). Further, the transmission CT shifts the torque of the transmission input shaft 2 that rotates integrally with the engine output shaft 1 according to the operating state, and reverses the rotation direction when the reverse range is selected to change the transmission output shaft 2. It is designed to output to 3. The transmission output shaft 3
Is transmitted to the drive wheels (not shown) via the reduction gear mechanism 4 and the differential device 5.

The transmission CT includes a torque converter 7 that shifts the torque of the transmission input shaft 2 through hydraulic oil and outputs it to the turbine shaft 6, and the rotation of the turbine shaft 6 when the reverse range is selected. A forward / reverse switching mechanism 9 that reverses and transmits to the intermediate shaft 8 and a belt type continuously variable transmission 10 that continuously changes the torque of the intermediate shaft 8 and outputs it to the transmission output shaft 3 (hereinafter, referred to as CV
(T10) is provided.

The torque converter 7 includes a pump cover 11
A pump 12 connected to a transmission input shaft 2 via a turbine, a turbine 14 connected to a turbine shaft 6 via a connecting member 13 and driven to rotate by hydraulic fluid discharged from the pump 12, and a turbine 14 to a pump 12 And a stator 15 that rectifies the hydraulic oil that flows back to the pump 12 in a direction that assists the rotation of the pump 12. The transmission has a gear ratio corresponding to the speed ratio of the pump 12 and the turbine 14 (turbine rotation speed / pump rotation speed). The torque of the input shaft 2 is changed. Here, the stator 15 is a one-way clutch 16
It is fixed to the transmission case 25 (fixed portion) via the.

Further, the torque converter 7 is provided with a lockup clutch 17 for directly connecting (locking up) the transmission input shaft 2 and the turbine shaft 6 in a predetermined operating region in order to improve fuel efficiency. The lockup clutch 17 is locked up (on) when hydraulic pressure is applied to the rear oil chamber 17r from a hydraulic mechanism FS, which will be described later, and released when hydraulic pressure is applied to the front oil chamber 17f. (Off). An oil pump 19 that is rotatably driven by the pump 12 (pump shell 49) via the connecting shaft 18 is disposed slightly behind (on the left side in FIG. 2) the torque converter 7.

The forward / reverse switching mechanism 9 is a planetary gear system. The forward / backward switching mechanism 9 includes a ring gear 21 connected to the turbine shaft 6 via a torque input member 20 and a sun gear connected to the intermediate shaft 8. 22
And a plurality of pinion gears 23 that mesh with the ring gear 21 and the sun gear 22, and rotate these pinion gears 23.
A carrier 24 that supports (rotatably) is provided. A forward clutch 26 is provided between the torque input member 20 and the carrier 24, and the reverse brake 2 is provided between the carrier 24 and the transmission case 25.
7 is provided. Here, the forward clutch 26
The reverse brake 27 is turned on (fastened) when hydraulic pressure is supplied from a hydraulic mechanism FS, which will be described later,
It is designed to be turned off (released) when the hydraulic pressure is released.

In the forward / reverse switching mechanism 9, when both the forward clutch 26 and the reverse brake 27 are turned off, the neutral state is established and torque is not transmitted from the turbine shaft 6 to the intermediate shaft 8. When only the forward clutch 26 is turned on, the ring gear 21 and the carrier 24 cannot be differentiated from each other, so that the forward / reverse switching mechanism 9 is directly connected and the intermediate shaft 8 is integrated with the turbine shaft 6 in the same direction. It rotates and the drive wheels are driven in the forward direction.

When only the reverse brake 27 is turned on, the carrier 24 is fixed to the transmission case 25, so that the ring gear 21, the pinion gear 23 and the sun gear 2 are provided.
2 and 2 function as a fixed gear train that meshes in this order. At this time, the sun gear 22 rotates in the opposite direction to the ring gear 21, so that the intermediate shaft 8 becomes the turbine shaft 6.
And the drive wheels are driven in the reverse direction.
In this case, the gear is changed at a gear ratio determined by the number of teeth of the ring gear 21 and the number of teeth of the sun gear 22. The forward clutch 26 and the reverse brake 27 are not both turned on.

The CVT 10 includes a primary pulley 31 (driving pulley) that rotates integrally with the intermediate shaft 8, a secondary pulley 32 (driven pulley) that rotates integrally with the transmission output shaft 3, and a primary pulley 31 and a secondary pulley 32.
And a V-belt 33 that transmits torque between and. Note that, hereinafter, for convenience, the engine side (right side in FIG. 2) when viewed in the axial direction of the intermediate shaft 8 is referred to as “front” or “front”, and the opposite side is referred to as “rear” or “rear”.

The primary pulley 31 is the intermediate shaft 8
And a first movable conical plate 35 which is arranged on the rear side of the first fixed conical plate 34 so as to be opposed thereto and is movable in the front-rear direction. ing. Then, a primary oil chamber 36 (hydraulic piston mechanism) that controls the position of the first movable conical plate 35 in the front-rear direction
Is provided. When hydraulic pressure is applied to the primary oil chamber 36, hydraulic oil is supplied into the primary oil chamber 36, and the first movable conical plate 35 moves to the front side to move the V-belt 33.
The holding position of is changed to the outer peripheral side, and the effective pulley diameter of the primary pulley 31 is increased. On the contrary, when the hydraulic pressure is released, the working oil in the primary oil chamber 36 is drained and the effective pulley diameter of the primary pulley 31 is reduced. That is, the effective pulley diameter of the primary pulley 31 can be freely changed by supplying or discharging hydraulic pressure or hydraulic oil to or from the primary oil chamber 36.

The secondary pulley 32 has basically the same structure as the primary pulley 31, and includes a second fixed conical plate 37 fixed to the transmission output shaft 3 and a front side of the second fixed conical plate 37. The second which is arranged so as to face this
It is composed of a movable conical plate 38. A secondary oil chamber 39 is provided to control the position of the second movable conical plate 38 in the front-rear direction.

In the CVT 10, the hydraulic mechanism F
A hydraulic pressure (hereinafter, referred to as a primary hydraulic pressure) corresponding to the gear ratio to be set is applied from S to the primary oil chamber 36. On the other hand, the secondary oil chamber 39 basically has V
A hydraulic pressure sufficient to maintain the tension of the belt 33, that is, a minimum hydraulic pressure capable of transmitting the driving force without causing belt slip (hereinafter, referred to as secondary hydraulic pressure) is applied. That is, in CVT10,
The gear ratio is determined by the primary hydraulic pressure, and the belt tension is determined by the secondary hydraulic pressure. As will be described later, since the line pressure of the hydraulic mechanism FS is introduced into the secondary oil chamber 39, the secondary hydraulic pressure is substantially synonymous with the line pressure.

Specifically, when the primary hydraulic pressure rises, the effective pulley diameter of the primary pulley 31 increases accordingly. For this reason, the tension of the V-belt 33 tries to increase, but the secondary hydraulic pressure is adjusted so as not to increase this tension.
(Line pressure) is adjusted, and the effective pulley diameter of the secondary pulley 32 is reduced. In this way, the primary pulley 31
While the effective pulley diameter of the secondary pulley 3 increases
Since the effective pulley diameter of No. 2 becomes small, the gear ratio of the CVT 10 changes to the speed increasing side (OD side). On the other hand, when the primary oil pressure decreases, the gear ratio of the CVT 10 changes to the deceleration side (LOW side) contrary to the above case.

The hydraulic mechanism F is connected to the transmission CT.
S is provided, and this hydraulic mechanism FS responds to a signal from the control unit CU in accordance with the operating state, according to the operating state, the front oil chamber 17f and the rear oil chamber 17r of the lockup clutch 17, the forward clutch 26 of the forward / reverse switching mechanism 9 and the reverse clutch. The brake 27, the primary oil chamber 36 and the secondary oil chamber 39 of the CVT 10 are supplied with and discharged with operating oil or control oil pressure to perform a predetermined gear shift operation. Here, the line pressure (secondary pressure) of the hydraulic mechanism FS is also controlled by the control unit CU.

The hydraulic mechanism FS will be described below. As shown in FIG. 3, the hydraulic mechanism FS is supplied with operating oil (original pressure) from the oil pump 19. The hydraulic mechanism FS includes a line pressure adjusting valve 41, a pressure reducing valve 42, a gear ratio control valve 43, a gear ratio fixed valve 44,
A hydraulic pressure correction valve 45, a clutch valve 46, a manual valve 47, a relief valve 48, a lockup valve 49, etc. are provided. Here, the gear ratio control valve 4
3 is controlled by the first duty solenoid 51,
The fixed gear ratio valve 44 is controlled by the first on / off solenoid 52, the hydraulic pressure correction valve 45 is controlled by the second duty solenoid 53, and the clutch valve 46.
Is controlled by the clutch duty solenoid 54, and the lockup valve 49 is controlled by the second on / off solenoid 55.

In the hydraulic mechanism FS, the hydraulic oil discharged from the oil pump 19 is first adjusted to a predetermined line pressure by the line pressure adjusting valve 41, and the line L1
The oil is supplied to the secondary oil chamber 39 through the (hydraulic passage) and to the clutch valve 46 through the line L2. The clutch valve 46 adjusts the hydraulic pressure in the line L2 to a predetermined pressure by the clutch duty solenoid 54, and then supplies the adjusted hydraulic pressure to the manual valve 47 and the lockup valve 49 via the line L3. Has become. Pressure reducing valve 42
Reduces the line pressure supplied to the secondary oil chamber 39 to form pilot pressure for the hydraulic pressure correction valve 45, the gear ratio control valve 43, the gear ratio fixed valve 44, and the clutch valve 46.

The pilot pressure for controlling the line pressure is adjusted by controlling the duty ratio of the second duty solenoid 53. That is, the hydraulic pressure controlled by the second duty solenoid 53 is introduced into the pilot chamber of the hydraulic pressure correction valve 45, the hydraulic pressure correction valve 45 is opened / closed according to this hydraulic pressure, and the line L4 in the line L4 formed according to this open / closed state is opened. Hydraulic pressure is line pressure adjustment valve 4
It is introduced as a pilot pressure of 1 to obtain a desired line pressure. The hydraulic pressure correction valve 45
Alternatively, the line pressure adjusting valve 41 may be directly controlled by a duty solenoid or the like.

The gear ratio control valve 43 is controlled by the first duty solenoid 51, and the gear ratio control valve 4
The hydraulic pressure in the line L6 formed by 3 is supplied to the primary oil chamber 36 via the gear ratio fixed valve 44. The fixed gear ratio valve 44 is controlled by the first on / off solenoid 52. When the first on / off solenoid 52 is in the on state, the line L7 connected to the primary oil chamber 36 communicates with the line L6, while in the off state. The communication is cut off. In other words, by turning off the first solenoid 52, the hydraulic pressure applied to the primary oil chamber 36 is fixed to the current value irrespective of the operation of the gear ratio control valve 43, thereby fixing the gear ratio. It has become.

The gear ratio control valve 43 is controlled by the first duty solenoid 51, and when the first duty solenoid 51 is in the ON state, the hydraulic pressure in the primary oil chamber 36 is in order of line L7, line L6 and line L8. And drained through the relief ball 58,
No oil pressure is applied to the primary oil chamber 36. On the other hand, the first
When the duty solenoid 51 is off,
While the line L8 (drain path) is closed, the transmission ratio control valve 43 is opened at an opening ratio according to the duty ratio of the first duty solenoid 51, and the line pressure is changed to the orifice 59.
And is introduced into the primary oil chamber 36 via the line L6. Since the orifice 59 is provided, the oil chamber in the primary oil chamber 36 does not suddenly rise.

The clutch valve 46 is controlled by the clutch duty solenoid 54, and the line pressure adjusted by the clutch duty solenoid 54 is
It is supplied to the manual valve 47 and the lockup control valve 49 via the line L3. This adjusted line pressure is applied to the forward clutch 2 via the line L3, the manual valve 47 and the line L10 in the forward drive state.
6, while the hydraulic pressure in the reverse brake 27 is released through the line L12. On the other hand, in the reverse drive state, the line pressure is set to the line L3 and the line L1 only when the lockup valve 49 is in the non-lockup state.
3 is supplied to the reverse brake 27 via the line L12.

The lockup valve 49 is controlled by the second on / off solenoid 55, and at the time of lockup,
The line L16 connected to the front hydraulic pressure 17f communicates with the relief valve 48 via the relief line L15. On the other hand, when the lockup is released, the line L17 connected to the rear oil chamber 17r communicates with the relief valve 48 via the relief line L15.

Next, the control mechanism of the transmission CT will be described. As shown in FIG. 4, the control mechanism of the transmission CT includes:
Control unit C consisting of a microcomputer
U is provided. The control unit CU includes a shift position signal (P, R, N, D, 2, 1) detected by the shift position sensor 62 and a primary pulley 31 detected by the primary rotation speed sensor 63.
Rotation speed (hereinafter, referred to as primary rotation speed), the rotation speed of the secondary pulley 32 detected by the secondary rotation speed sensor 64 (hereinafter, referred to as secondary rotation speed), and the throttle detected by the throttle opening sensor 65. The opening degree, the engine speed detected by the engine speed sensor 66, the turbine speed detected by the turbine speed sensor 67, the oil temperature detected by the oil temperature sensor 68, the oil pressure detected by the oil pressure sensor 69, and the like. It is designed to be input as control information.

The control unit CU is a comprehensive control device for the transmission CT including the line pressure control means, the deceleration line pressure correction means and the deceleration state detection means described in the claims. A predetermined control signal is output to each of the solenoids 51 to 55 and the like based on various control information to perform a predetermined control. The line pressure control during deceleration braking according to the gist of the present application will be described below. Only explained.

The control method of the line pressure control of the hydraulic mechanism FS by the control unit CU will be described below with reference to the flow chart shown in FIG. When control starts, first in step # 1,
The throttle opening detected by the throttle opening sensor 65, the engine speed detected by the engine speed sensor 66, the brake signal detected by the brake sensor 70, etc. are read as control information. Next, in step # 2, it is compared and determined whether or not deceleration braking is being performed, that is, whether or not the throttle opening is 0 and the brake signal is on (the brake pedal is depressed). . If it is determined that it is not during deceleration braking (NO), step # 5 is executed after the normal-time line pressure target value is calculated in step # 4.

In normal times (other than during deceleration braking),
The line pressure target value, that is, the secondary pressure target value is set to a minimum pressure within a range in which belt slip does not occur between the V-belt 33 and the pulleys 31 and 32. This is because if the line pressure is increased more than necessary, the fuel consumption is reduced due to an increase in pump loss, the durability of the V-belt 33 is reduced due to an increase in belt tension, and noise is further generated. In this case, as described above (prior art),
The line pressure target value basically needs to be set according to the input torque to the CVT 10 and the gear ratio of the CVT 10, but it is difficult to detect the input torque to the CVT 10. The line pressure target value is set according to the throttle opening, which is an index of the input torque, and the gear ratio of the CVT 10. The normal line pressure target value is set with respect to the input torque (throttle opening) and the gear ratio, for example, with the characteristics shown in FIG. In FIG. 6, LOW is the maximum value of the gear ratio that can be set in the CVT 10, and OD is the minimum value.
(Overdrive).

On the other hand, if it is determined in step # 2 that deceleration braking is being performed (YES), in step # 3, the line pressure target value for deceleration braking is calculated, and then step #
5 is executed. As described above (prior art), when the line pressure is set in accordance with the throttle opening degree during the deceleration braking as in the normal state, the primary chamber 3
The drain speed of the hydraulic oil in 6 decreases, and the gear ratio becomes LO.
Since it does not return to W and problems such as deterioration of acceleration or belt squeal occur when re-accelerating or restarting, the line pressure (secondary pressure) is corrected to a higher pressure side than during normal operation during deceleration braking. .

As shown in FIG. 7, the target value of the line pressure for deceleration braking is calculated from the oil pump discharge flow rate G ₁ to the hydraulic oil for lubricating each part, the hydraulic oil required by the torque converter 7, etc. A value obtained by subtracting the flow rate of hydraulic oil used for other than line pressure control (hereinafter referred to as lubricating hydraulic oil), that is, the flow rate G _{2 of} hydraulic oil that can be used for line pressure control (hereinafter, referred to as The supplyable flow rate) is set to the maximum pressure P ₁ which can be set under the condition that the flow rate G _{3 of} hydraulic oil required for the shift operation of the CVT 10 (hereinafter, referred to as the required shift flow rate) is not less than To be done.

More specifically, the oil pump discharge flow rate increases as the engine speed increases, but decreases as the line pressure increases if the engine speed is constant, as shown in FIG. In other words, when the engine speed is constant, the oil pump discharge flow rate decreases correspondingly when trying to obtain a high line pressure. Then, the flow rate of the lubricating hydraulic oil increases as the line pressure increases, as shown in FIG. 9, for example. Therefore, the flow rate of the hydraulic oil that can be supplied decreases as the line pressure increases, as indicated by G ₂ in FIG. 7.

Further, the necessary shift speed of the hydraulic oil during deceleration braking is determined by the line pressure and the gear ratio as shown in FIG. 10. Accordingly, G _{3 in} FIG.
Increase like. The required shift flow rate is determined by the line pressure and the gear ratio because the required shift flow rate is a function of the primary pressure while the primary pressure is the line pressure.
This is because it is a function of (secondary pressure) and the gear ratio.

Here, when the required shift speed of the hydraulic oil is smaller than the supplyable flow rate, the shifting operation is performed without any trouble. However, when the required shift flow rate exceeds the supplyable flow rate, the flow rate of the lubricating hydraulic oil becomes insufficient, Alternatively, a problem occurs such that the gear shifting operation is not smoothly performed.
Therefore, during deceleration braking, it is necessary to maintain the flow rate of hydraulic oil that can be supplied above the flow rate necessary for shifting (for example, the line pressure in FIG. 7 is P ₀ ).

On the other hand, during such deceleration braking,
As the line pressure is increased, the drain speed of the hydraulic oil from the primary oil chamber 36 can be increased, and the gear ratio can be returned to LOW more quickly. Therefore, from this viewpoint, it is preferable to increase the line pressure as much as possible.

Therefore, during deceleration braking, the line pressure at which the gear ratio can be returned to LOW as quickly as possible without causing any trouble in the gear shifting operation, that is, the optimum line pressure during deceleration braking is as shown in FIG. It is the line pressure P ₁ when the supplyable flow rate and the required shift flow rate are equal. Therefore, in this embodiment, the line pressure target value is set to P ₁ during deceleration braking.

More specifically, the control unit C
In U, such an optimum line P ₁ at the time of deceleration braking is stored as a map having the engine speed and the gear ratio of the CVT 10 as parameters, the characteristics of which are shown in FIG. 11, for example. At times,
An optimum line pressure target value is set according to the operating state of the engine using such a map. In this way, since the optimum line pressure P ₁ during deceleration braking is mapped, the control unit C
The control logic for U or line pressure control is simplified.

After the line pressure target value is set in this way, the duty ratio corresponding to the line pressure target value is calculated in step # 5, and this duty ratio is applied to the second duty solenoid 53 to generate the line pressure. Is preferably controlled so as to follow the line pressure target value. In this way, during deceleration braking, it is possible to maximize the speed of change in the gear ratio while ensuring the amount of hydraulic oil supplied to each part that requires hydraulic oil, and to re-accelerate or restart after deceleration braking. Can be smoothed. Then, the process returns to step # 1.

[0048]

According to the first aspect of the present invention, the line pressure (secondary pressure) is reduced during predetermined deceleration (during deceleration braking).
Since the flow rate of hydraulic oil required for shifting operation of the continuously variable transmission is set to the maximum value under the condition that it is less than the flow rate of hydraulic oil that can be used for line pressure control, other than line pressure control It is possible to maximize the speed change operation after supplying sufficient hydraulic oil to the respective parts requiring the hydraulic oil, and smooth re-acceleration or restart after deceleration braking.

According to the second invention, basically, the same operation and effect as those of the first invention can be obtained. Furthermore, since the optimum line pressure during deceleration braking is mapped,
During deceleration braking, the optimum line pressure target value can be set simply by referring to this map, and the control device or control logic is simplified.

[Brief description of drawings]

FIG. 1 is a block diagram showing a configuration of a first or second invention corresponding to claim 1 or claim 2.

FIG. 2 is a system configuration diagram of a power train including a hydraulic control device for a continuously variable transmission according to the present invention.

FIG. 3 is a system configuration diagram of a hydraulic mechanism.

FIG. 4 is a system configuration diagram of a control mechanism including a control unit.

FIG. 5 is a flowchart showing a control method of line pressure control.

FIG. 6 is a diagram showing characteristics of a target line pressure value with respect to an input torque and a gear ratio.

FIG. 7 is a diagram showing characteristics of an oil pump discharge flow rate, a hydraulic oil supplyable flow rate, and a shift required flow rate with respect to a line pressure.

FIG. 8 is a diagram showing characteristics of an oil pump discharge flow rate with respect to line pressure and engine speed.

FIG. 9 is a diagram showing a characteristic of a flow rate of lubricating hydraulic oil with respect to a line pressure.

FIG. 10 is a diagram showing characteristics of an oil inflow speed into a secondary oil chamber with respect to a line pressure and a gear ratio.

FIG. 11 is a diagram showing characteristics of the optimum line pressure with respect to engine speed and gear ratio.

[Explanation of symbols]

 CE ... Engine CT ... Transmission CU ... Control unit FS ... Hydraulic mechanism 10 ... Continuously variable transmission (CVT) 19 ... Oil pump 31 ... Primary pulley 32 ... Secondary pulley 33 ... V belt 36 ... Primary oil chamber 39 ... Secondary oil chamber

Claims (2)

[Claims] 1. A hydraulic control device for a belt-type continuously variable transmission provided with a hydraulic mechanism for continuously changing a gear ratio by changing a pulley diameter by hydraulic pressure. Line pressure control means set according to the operating state, deceleration state detection means for detecting whether or not the vehicle is in a predetermined deceleration state, and deceleration state detection means for determining that the vehicle is in the predetermined deceleration state When detected, the flow rate of the hydraulic oil that can be used for controlling the line pressure in the hydraulic oil discharged from the oil pump is less than or equal to the hydraulic oil flow rate required for the shifting operation of the continuously variable transmission. A hydraulic control device for a continuously variable transmission, comprising: line pressure correction control means during deceleration for increasing the line pressure to a maximum value that can be set under non-restrictive conditions. 2. The hydraulic control device for a continuously variable transmission according to claim 1, wherein a flow rate of hydraulic oil that can be used for controlling the line pressure is required for a gear shifting operation of the continuously variable transmission. There is a map that shows the maximum line pressure value that can be set under conditions that do not fall below the hydraulic oil flow rate, for engine speed and gear ratio, and is set by the line pressure correction control means during deceleration. A hydraulic control device for a continuously variable transmission, wherein the maximum value of the above-mentioned line pressure to be read is read from the map.