JPH06109114A - Hydraulic control device of hydraulically operated transmission - Google Patents

Abstract

PURPOSE:To provide a hydraulic control device of a hydraulically operated transmission capable of surely securing the required line pressure without unnecessarily increasing the line pressure. CONSTITUTION:When the engine speed does not exceed the specified value a1, the target line pressure corresponding to the actual engine torque is set by a control unit CU, and the required minimum line pressure is set, the pump loss is reduced, and the fuel consumption is improved. When the engine speed exceeds the specified value a2 (>a1), the target line pressure corresponding to the maximum throttle opening is set, and the line pressure is prevented from being reduced to the required line pressure by the delay in responsiveness of the hydraulic pressure, and the belt slip of CVTI10 is surely prevented. When the engine speed ranges from a1 to a2, the target line pressure is increased according to the engine speed, and the required line pressure is secured without damaging the fuel consumption.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a hydraulic control system for a hydraulically operated transmission.

[0002]

2. Description of the Related Art Generally, an automobile is provided with a transmission that changes the output torque of an engine according to the driving condition.
As such a transmission, a multi-stage transmission such as a transmission gear mechanism including a planetary gear system or a continuously variable transmission such as a belt type continuously variable transmission has been conventionally used. Transmission-type transmissions, that is, hydraulically-operated transmissions are often used. Then, in such a hydraulically operated transmission, a hydraulic mechanism for supplying hydraulic pressure to each hydraulic device is provided.

For example, a hydraulically actuated belt type continuously variable transmission
(Hereinafter, this is referred to as CVT) is usually provided with 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. The gear ratio can be set to an arbitrary value by changing the pulley diameter. A hydraulic piston is provided for each of the pulleys in order to change the pulley diameter, and a hydraulic mechanism is provided for supplying hydraulic pressure to both hydraulic pistons.

The line pressure of the hydraulic mechanism for operating the hydraulically actuated transmission must be set at least according to the input torque to the hydraulically actuated transmission. That is, if the line pressure is too high with respect to the input torque, problems such as a decrease in fuel efficiency due to an increase in power loss and a shift shock occur. Conversely, if the line pressure is too low, the shift operation becomes slow. Problems such as occur.

However, since it is difficult to directly detect the input torque to the hydraulically actuated transmission, it is usually difficult to detect the input torque to the hydraulically actuated transmission in the conventional hydraulic mechanism. A hydraulically actuated transmission in which a proportional engine torque is used and a line pressure is set based on, for example, the engine torque and a gear ratio has been proposed (for example, JP-A-62-5324).
(See Japanese Patent Publication No. 5). Even if a torque converter is installed between the engine and the hydraulically operated transmission, the input torque to the hydraulically operated transmission is almost proportional to the engine torque in the lockup region of the torque converter. By setting the line pressure based on the gear ratio and the gear ratio, basically, an appropriate line pressure can be obtained.

[0006]

The hydraulic mechanism of such a hydraulically actuated transmission is inevitably accompanied by a hydraulic response delay. Therefore, during a transition, the change in line pressure may be delayed with respect to the change in engine torque depending on the operating state. For this reason, there is a problem that the line pressure becomes lower than the required line pressure at the time of acceleration, and the speed change operation becomes slower. Particularly, when the transmission is a CVT, there is a problem that a belt slip occurs.
In order to deal with this, in the hydraulic mechanism of the conventional hydraulically operated transmission, a relatively high safety factor is set with respect to the line pressure. However, doing so increases pump loss and reduces fuel efficiency. Will be invited.

The present invention has been made in order to solve the above-mentioned conventional problems, and reduces the pump loss while setting the line pressure of the hydraulic mechanism based on the engine torque and the gear ratio, thereby reducing fuel consumption. It is an object of the present invention to provide a hydraulic control device for a hydraulically actuated transmission that can improve the hydraulic pressure.

[0008]

In order to achieve the above object, as shown in FIG. 1, a first aspect of the present invention relates to a hydraulic pressure of a hydraulically actuated transmission provided with a speed change mechanism B operated by a hydraulic mechanism A. In the control device, the line pressure target value setting means D for setting the line pressure target value of the hydraulic mechanism A according to the output torque of the engine C, and the line pressure for controlling the line pressure so as to follow the line pressure target value. Control means E,
According to the engine speed, the higher the engine speed, the more the line pressure target value correction means F that corrects the line pressure target value to the high pressure side is provided. A hydraulic control device is provided.

According to a second aspect of the present invention, in the hydraulic control device for a hydraulically actuated transmission according to the first aspect, the line pressure target value correction means F is such that the response of the hydraulic mechanism A to changes in the engine load is In the rotation range slower than the response of the engine C to the change of the engine C, the line pressure target value is corrected to the high pressure side up to the hydraulic pressure set when the output torque of the engine C is maximum. A hydraulic control device for a hydraulically actuated transmission is provided.

A third aspect of the present invention is a hydraulic control device for a hydraulically actuated transmission provided with a speed change mechanism B that is operated by the hydraulic mechanism A, and a line pressure target value of the hydraulic mechanism A according to the output torque of the engine C. And a line pressure target value setting means D for setting
A line pressure control means E for controlling the line pressure and a calculation cycle control means G for shortening the calculation cycle of the line pressure target value setting means D and the line pressure control means E as the engine speed increases are provided. A hydraulic control device for a hydraulically actuated transmission characterized by being provided.

A fourth aspect of the present invention is the hydraulic control device for a hydraulically actuated transmission according to any one of the first to third aspects, wherein the line pressure target value correction means F is a hydraulic mechanism for a change in engine load. In a rotation range in which the responsiveness of A is faster than the responsiveness of the engine C with respect to changes in engine load, the line pressure target value is set to a normal target value that does not depend on the engine speed. A hydraulic control device for a hydraulically actuated transmission is provided.

A fifth aspect of the present invention is the hydraulic control device for a hydraulically actuated transmission according to the fourth aspect, wherein the speed change mechanism B is a belt type continuously variable transmission, and the line pressure target value setting means D.
However, in the rotation region where the responsiveness of the hydraulic mechanism A to the change of the engine load is faster than the responsiveness of the engine C to the change of the engine load, the line pressure is determined based on the input torque and the gear ratio of the belt type continuously variable transmission. A hydraulic control device for a hydraulically actuated transmission, characterized in that a target value is set.

[0013]

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 via hydraulic oil and outputs the torque to the turbine shaft 6, and 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 no torque is 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. A primary oil chamber 36 that controls the position of the first movable conical plate 35 in the front-rear direction is provided.
Here, when hydraulic pressure is applied to the primary oil chamber 36, hydraulic oil is supplied into the primary oil chamber 36, the first movable conical plate 35 moves to the front side, and the holding position of the V belt 33 changes to the outer peripheral side, The effective pulley diameter of the primary pulley 31 becomes large. Conversely, when the hydraulic pressure is released, the hydraulic oil in the primary oil chamber 36 is drained and the primary pulley 31
The effective pulley diameter becomes smaller. 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 that of 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 of the secondary pulley 32 (hereinafter referred to as the primary rotation speed), the rotation speed of the secondary pulley 32 detected by the secondary rotation speed sensor 64 (hereinafter referred to as the 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 includes a line pressure target value setting means, a line pressure control means, a line pressure target value correcting means, and a calculation cycle control means, and comprehensively controls the transmission CT. The device is configured to output a predetermined control signal to each of the solenoids 51 to 55 and the like on the basis of the above various control information to perform a predetermined control. However, in the following, a line related to the gist of the present application will be described. Only the pressure control will be described.

The control method of the line pressure control by the control unit CU will be described below according to the flow chart shown in FIG. 5 and with reference to FIGS. When the control is started, first in step # 1, various control information such as throttle opening TVO, engine speed Ne, primary speed Np, secondary speed Ns, etc. is read.
Then, in step # 2, the engine speed Ne is set to a predetermined value a.
_If it is determined that Ne> a ₁ , it is determined whether Ne> a ₁ is exceeded, and in step # 3, the engine speed Ne is set to a value greater than a ₁ above. a
It is determined whether it exceeds ₂ . That is, in the step # 2, # 3, the engine rotational speed Ne, or is a ₁ or less, or it is large and a ₂ less than a _1, or will be determined whether exceeds a ₂ . And Ne ≦ a
_{If 1} , then steps # 4 and # 5 are executed and then step #
11 to step # 14 are executed, and if Ne> a ₂ , steps # 6 and # 7 are executed, and then steps # 11 to # 14 are executed, and if a ₁ <Ne ≦ a ₂ , step # After steps # 8, # 9, and # 10 are executed, steps # 11 to # 14 are executed.

Here, a ₁ is in the lower rotation region than this,
The hydraulic response delay time (line pressure response delay time) of the hydraulic mechanism FS with respect to the throttle opening change (that is, engine load change) becomes shorter than the engine torque response delay time with respect to the throttle opening change. On the other hand, the boundary value is set so that the relationship between the two is reversed in the higher rotation range than this.

For example, as shown in FIG. 10, during acceleration such that the throttle opening increases to a predetermined value in time τs, the line pressure change of the hydraulic mechanism FS with respect to the throttle opening change is accompanied by a response delay of τa. On the other hand, the engine torque change is accompanied by a response delay of τb as shown in FIG. 11, for example. Here, the hydraulic response delay time τa
Is almost independent of the engine speed as indicated by the straight line J ₁ (computation frequency 40 Hz) in FIG. 12, but the engine torque response delay time τb is increased by the engine speed as indicated by the straight line J ₂ in FIG. Becomes shorter with. Where τa
The engine speed at which = τb is set to a ₁ . When the cycle of line pressure control by the control unit CU is shortened (frequency becomes high), for example, the hydraulic response delay time τa becomes short as shown by a straight line J ₁ ′ (for example, operation frequency 80 Hz) in FIG. Therefore, the engine speed a ₁ 'where τa = τb shifts to the high speed side.

In this embodiment, since the hydraulic response delay time τa is shorter than the engine torque response delay time τb in the region of Ne ≦ a ₁ , the throttle opening (K ₁ ) changes as shown in FIG. 13, for example. Even during acceleration, the line pressure change (K ₂ ) does not lag behind the engine torque change (K ₃ ). Therefore, when Ne ≦ a ₁ , the target line pressure is set according to the actual engine torque (that is, the input torque to the CVT 10). That is, the minimum required line pressure is set, and the pump loss is reduced.

On the other hand, in the region of Ne> a ₁ , the hydraulic response delay time τa becomes longer than the engine torque response delay time τb, so that, for example, as shown in FIG.
During acceleration when (K ₁ ') changes, the line pressure changes
(K ₂ ') cannot follow the engine torque change (K ₃ '). Therefore, in the shaded portion in FIG. 14, the line pressure corresponding to the actual engine torque, that is, the required line Only a line pressure lower than the pressure is obtained, which causes belt slip.

Therefore, when Ne> a ₁ , basically, a virtual engine torque for line pressure calculation (hereinafter, referred to as line pressure calculation torque) larger than the actual engine torque is set. The line pressure target value is set based on the line pressure calculation torque so that the line pressure target value does not become lower than the line pressure corresponding to the actual engine torque. That is, in the rotation region where the response delay of the line pressure is a problem, the line pressure target value is added for safety. In this case, since the line pressure (K ₂ ") is increased from the beginning during acceleration when the throttle opening (K ₁ ") changes as shown in FIG. 15, the case shown in FIG. No such problems occur, and belt slip does not occur.

However, even in the region where Ne> a ₁ , the higher the engine speed Ne is, the larger the difference between the hydraulic response delay time τa and the engine torque response delay time τb becomes. Therefore, Ne> a ₂ (> a ₁ ), The engine torque corresponding to the maximum throttle opening (maximum engine torque)
Is used as the line pressure calculation torque, while the line pressure calculation torque is gradually increased as the engine speed Ne increases in the region where a ₁ <Ne ≦ a ₂ (see FIG. 7).

The specific calculation method of the line pressure calculation torque is as follows. That is, when Ne ≦ a ₁ , in step # 4, the engine torque Tr corresponding to the throttle opening TVO and the engine speed Ne is calculated using the map having the characteristics shown in FIG. In # 5, the engine torque Tr is used as the line pressure calculation torque Tin.

If Ne> a ₂ , then in step # 6 the throttle opening T is calculated using the map having the characteristics shown in FIG.
The engine torque Tmax corresponding to the case where VO is 100% (curve G ₁ ) is calculated, and this engine torque Tmax is set as the line pressure calculation torque Tin in step # 7.

If a ₁ <Ne ≤ a ₂ , then step # 8
Then, the actual engine torque Tr is calculated in the same manner as in step # 4, and then in step # 9, the engine torque Tmax corresponding to TVO = 100% is calculated in the same manner as in step # 6. After this step #
In 10, the line pressure calculation torque Tin is calculated based on Tr and Tmax calculated in steps # 8 and # 9 by the following equation 1.

[Equation 1] Tin = Tr + (Tmax−Tr) · (Ne−a ₁ ) / (a ₂ −a ₁ ) ………………………………………………………………………………………… 1 1 The characteristic of the above with respect to the engine speed Ne is as shown in FIG. 7, for example.

After the line pressure calculation torque Tin is calculated according to the engine speed Ne in this way, in any case, first, at step # 11, a map having the characteristics shown in FIG. The target line pressure Ps is calculated based on the line pressure calculation torque Tin and the gear ratio R of the CVT 10.

Next, in step # 12, step # 11
The target line pressure Ps calculated in step 1 is multiplied by a predetermined safety factor Ks, the duty ratio corresponding to the target line pressure Ps is subsequently calculated in step # 13, and then the first duty is calculated in step # 14 in accordance with the above duty ratio. Solenoid 5
3 is driven, and the line pressure is controlled so as to follow the target line pressure.

When such line pressure control is performed,
The characteristic of the target line pressure, that is, the line pressure that is actually formed with respect to the engine speed Ne becomes as shown by a polygonal line H ₁ in FIG. In FIG. 9, the straight line H ₂ is the line pressure actually required (required line pressure), and the difference d between H ₁ and H ₂ is the line pressure added for safety, that is, the safety factor. As is apparent from FIG. 9, according to the line pressure control, the additional line pressure d, that is, the safety factor is set to the necessary minimum according to the change in the engine speed Ne, that is, the response of the engine torque. Therefore, the line pressure cannot be increased unnecessarily, so that the pump loss is reduced and the fuel efficiency is improved.

As described above, according to this embodiment, the required line pressure can be reliably secured without unnecessarily increasing the line pressure addition (safety factor). The calculation cycle of the line pressure control is set shorter as the engine speed is higher.
If (calculation frequency is increased), when the engine speed increases, both the engine torque response delay time and the hydraulic pressure response delay time become short, so it is necessary to increase the safety factor as the engine speed increases. It becomes possible to set the safety factor smaller.

[0049]

In general, when the engine speed is high, the response delay of the engine torque to the engine load change (throttle opening change) becomes short, so the line pressure change is delayed from the engine torque change, However, according to the first aspect of the invention, the target line pressure is corrected to a higher pressure side as the engine speed is higher, so that the required line pressure is irrespective of the engine speed. Is secured. Further, since the line pressure target value is set low when the engine speed is low, the line pressure is not unnecessarily increased, pump loss is reduced, and fuel economy is improved. That is, the safety factor for setting the line pressure is set corresponding to the response of the engine torque, and the necessary line can be secured without setting the line pressure unnecessarily high.

According to the second invention, basically, the same operation and effect as those of the first invention can be obtained. Further, in a region where the hydraulic response is slower than the engine torque response, the line pressure target value is set to the line pressure corresponding to the maximum engine torque, so that the required line pressure is reliably ensured at the time of high rotation.

According to the third aspect of the invention, the higher the engine speed, the shorter the calculation cycle is set, and the faster the hydraulic response becomes. Therefore, when the engine speed is high, both the engine torque response delay time and the hydraulic pressure response delay time become short, so there is no need to increase the safety factor as the engine speed increases, and the safety factor should be set smaller. Will be able to.

According to the fourth invention, basically, the same operation and effect as any one of the first to third inventions can be obtained. Further, in the region where the hydraulic pressure response is faster than the engine torque response, the line pressure target value is set to the normal value, so the line pressure is not set unnecessarily high, and the pump loss is further reduced.

According to the fifth invention, basically, the same action and effect as the fourth invention can be obtained. Furthermore, since the speed change mechanism is a belt type continuously variable transmission, such line pressure control can reliably prevent the occurrence of belt slip while improving fuel efficiency.

[Brief description of drawings]

FIG. 1 is a block diagram showing a configuration of first to fifth inventions corresponding to claims 1 to 5.

FIG. 2 is a system configuration diagram of a power train including a hydraulic control device 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.

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

FIG. 6 is a diagram showing characteristics of engine torque with respect to engine speed and throttle opening.

FIG. 7 is a diagram showing characteristics of line pressure calculation torque with respect to engine speed.

FIG. 8 is a diagram showing characteristics of a target line pressure with respect to a gear ratio and a line pressure calculation torque.

FIG. 9 is a diagram showing a characteristic of a target line pressure with respect to an engine speed.

FIG. 10 is a diagram showing a hydraulic response with respect to a change in throttle opening.

FIG. 11 is a diagram showing engine torque responsiveness to changes in throttle opening.

FIG. 12 is a diagram showing characteristics of an oil pressure response delay time and an engine torque response delay time with respect to an engine speed.

FIG. 13 is a diagram showing a hydraulic pressure response and an engine torque response at a low rotation speed in the conventional line pressure control.

FIG. 14 is a diagram showing a hydraulic response and an engine torque response at a high speed in the conventional line pressure control.

FIG. 15 FIG. 1 in line pressure control according to the present invention
It is a figure similar to FIG.

[Explanation of symbols]

 CE ... Engine CT ... Transmission CU ... Control unit FS ... Hydraulic mechanism 10 ... Belt type continuously variable transmission (CVT)

 ─────────────────────────────────────────────────── ─── Continuation of the front page (72) Inventor Yuji Mori 3-1, Shinchi, Fuchu-cho, Aki-gun, Hiroshima Mazda Motor Corporation

Claims (5)

[Claims] 1. A line pressure target value setting for setting a line pressure target value of a hydraulic mechanism according to an output torque of an engine in a hydraulic control device for a hydraulically actuated transmission provided with a speed change mechanism operated by a hydraulic mechanism. Means, a line pressure control means for controlling the line pressure so as to follow the line pressure target value, and the line pressure target value is corrected to a high pressure side as the engine speed increases, depending on the engine speed. And a line pressure target value correcting means for controlling the hydraulic pressure of the hydraulically actuated transmission. 2. The hydraulic control device for a hydraulically actuated transmission according to claim 1, wherein the line pressure target value correction means is such that the responsiveness of the hydraulic mechanism with respect to changes in the engine load and the responsiveness of the engine with respect to changes in the engine load In a rotation range slower than the responsiveness, the line pressure target value is corrected to the high pressure side up to the hydraulic pressure set when the output torque of the engine is maximum. Hydraulic control device. 3. A line pressure target value setting for setting a line pressure target value of the hydraulic mechanism according to an output torque of an engine in a hydraulic control device for a hydraulically actuated transmission provided with a speed change mechanism operated by a hydraulic mechanism. Means, line pressure control means for controlling the line pressure so as to follow the line pressure target value, and the higher the engine speed, the shorter the calculation cycle of the line pressure target value setting means and the line pressure control means. A hydraulic control device for a hydraulically actuated transmission, comprising: 4. The hydraulic control device for a hydraulically actuated transmission according to claim 1, wherein the line pressure target value correction means is a response of the hydraulic mechanism to a change in engine load. However, in the rotational speed range that is faster than the responsiveness of the engine to changes in engine load, the line pressure target value is set to a normal target value that does not depend on the engine speed. Hydraulic control device for transmission. 5. The hydraulic control device for a hydraulically actuated transmission according to claim 4, wherein the speed change mechanism is a belt type continuously variable transmission, and the line pressure target value setting means sets the hydraulic pressure with respect to a change in engine load. In a rotation range where the response of the mechanism is faster than the response of the engine to changes in the engine load, the line pressure target value is set based on the input torque and the gear ratio of the belt type continuously variable transmission. A hydraulic control device for a hydraulically actuated transmission characterized in that: