JP3226625B2 - Hydraulic control device for hydraulically operated transmission - Google Patents

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 hydraulically operated transmission.

[0002]

2. Description of the Related Art Generally, an automobile is provided with a transmission for changing the output torque of an engine in accordance with the driving state.
As such a transmission, a multi-stage transmission such as a transmission gear mechanism comprising a planetary gear system, or a continuously variable transmission such as a belt-type continuously variable transmission has been conventionally used. 2. Description of the Related Art A transmission that performs an operation, that is, a hydraulically operated transmission is frequently used. In such a hydraulically operated transmission, a hydraulic mechanism for supplying hydraulic pressure to each hydraulic device is provided.

[0003] For example, a hydraulically operated belt-type continuously variable transmission
In the following (hereinafter referred to as CVT), a primary pulley (drive pulley) and a secondary pulley (driven pulley), each of which can change a pulley diameter, and a belt wound around both pulleys are provided. By changing the diameter of the pulley, it is possible to set an arbitrary speed ratio. Then, a hydraulic piston is provided for each of the pulleys to change the diameter of the pulley, and a hydraulic mechanism is provided to supply hydraulic pressure to both the hydraulic pistons.

[0004] The line pressure of a hydraulic mechanism for operating such a hydraulically operated transmission must be set at least according to the input torque to the hydraulically operated 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 may occur. Problems such as occur.

However, it is very difficult to directly detect the input torque to the hydraulically operated transmission. Therefore, such a conventional hydraulic mechanism usually has almost no input torque to the hydraulically operated transmission. There has been proposed a hydraulically operated 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 (for example, Japanese Patent Application Laid-Open No. Sho 62-5324).
No. 5). Even when a torque converter is interposed 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 lock-up region of the torque converter. If the line pressure is set based on the gear ratio and the speed ratio, basically an appropriate line pressure can be obtained.

[0006]

The hydraulic mechanism of such a hydraulically operated transmission is inevitably accompanied by a delay in hydraulic response. Therefore, during a transition, a change in the line pressure may be delayed with respect to a change in the engine torque depending on the operation state. For this reason, during acceleration, the line pressure becomes lower than the required line pressure, and there is a problem that the shift operation is slowed down. In particular, when the transmission is a CVT, there is a problem that belt slippage occurs.
To cope with this, the hydraulic mechanism of the conventional hydraulically operated transmission sets a relatively high safety factor with respect to the line pressure, but this increases pump loss and reduces fuel efficiency. Will be invited.

SUMMARY OF THE INVENTION The present invention has been made to solve the above-mentioned conventional problems. The present invention sets a line pressure of a hydraulic mechanism based on an engine torque and a gear ratio while reducing pump loss to improve fuel economy. It is an object of the present invention to provide a hydraulic control device for a hydraulically operated transmission that can increase the pressure.

[0008]

In order to achieve the above object, as shown in FIG. 1, a first aspect of the present invention relates to a hydraulically operated transmission provided with a transmission mechanism B operated by a hydraulic mechanism A. In the control device, a line pressure target value setting means D for setting a line pressure target value of the hydraulic mechanism A in accordance with the output torque of the engine C, and a line pressure for controlling the line pressure so as to follow the line pressure target value. Control means E;
A hydraulically actuated shift basically comprising: a line pressure target value correcting means F for correcting the line pressure target value to a higher pressure side as the engine speed increases in accordance with the engine speed. Provide hydraulic control device for machine.

In the hydraulic control apparatus for a hydraulically operated transmission according to the first invention, the line pressure target value correcting means F
In a rotation region in which the response of the hydraulic mechanism A to a change in the engine load is slower than the response of the engine C to a change in the engine load, the line pressure target value is set when the output torque of the engine C is maximum. A further feature is that correction is made to the high pressure side within the range up to the oil pressure.

According to a second aspect of the present invention, in a hydraulic control device for a hydraulically operated transmission provided with a transmission mechanism B operated by a hydraulic mechanism A, a line pressure target value of the hydraulic mechanism A in accordance with an output torque of an engine C is provided. And a line pressure target value setting means D for setting the following.
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 operated transmission.

According to a third aspect of the present invention, in the hydraulic control apparatus for a hydraulically operated transmission according to the first or second aspect, the line pressure target value correcting means F determines whether the responsiveness of the hydraulic mechanism A to a change in engine load is high. In a rotation range faster than the response of the engine C to a change in engine load, the line pressure target value is set to a normal target value independent of the engine speed. Provided is a hydraulic control device for a transmission.

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

[0013]

DESCRIPTION OF THE PREFERRED EMBODIMENTS Embodiments of the present invention will be specifically described below.
As shown in FIG. 2, the power train PT for an automobile is:
Four-cylinder engine CE including first to fourth cylinders # 1 to # 4
And a hydraulically operated transmission CT. Here, the engine CE applies the engine torque to the crankshaft 1
(Engine output shaft) to output to the transmission CT. Further, the transmission CT changes 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 thereby change the transmission output shaft. 3 is output. The transmission output shaft 3
Is transmitted to driving wheels (not shown) via the reduction gear mechanism 4 and the differential device 5.

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

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

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

The forward / reverse switching mechanism 9 is a planetary gear system. The forward / reverse 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
, A plurality of pinion gears 23 meshing with the ring gear 21 and the sun gear 22, and rotating these pinion gears 23
A carrier 24 is provided to support (rotate). A forward clutch 26 is provided between the torque input member 20 and the carrier 24, and a reverse brake 2 is provided between the carrier 24 and the transmission case 25.
7 are provided. Here, the forward clutch 26
And the reverse brake 27 are turned on (fastened) when hydraulic pressure is supplied from a hydraulic mechanism FS described later, respectively.
It is turned off (released) when the hydraulic pressure is released.

In the forward / reverse switching mechanism 9, when the forward clutch 26 and the reverse brake 27 are both off, a 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
2 function as a fixed gear train that meshes in this order. At this time, since the sun gear 22 rotates in the opposite direction to the ring gear 21, the intermediate shaft 8 is
And the drive wheels are driven in the reverse direction.
In this case, the gear is shifted at a speed ratio determined by the number of teeth of the ring gear 21 and the number of teeth of the sun gear 22. There is no case where both the forward clutch 26 and the reverse brake 27 are turned on.

The CVT 10 includes a primary pulley 31 (a driving pulley) that rotates integrally with the intermediate shaft 8, a secondary pulley 32 (a 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 for transmitting torque between the V-belt and the V-belt. Hereinafter, for convenience, the engine side (the right side in FIG. 2) as 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 connected to the intermediate shaft 8
And a first movable conical plate 35 disposed at the rear side of the first fixed conical plate 34 so as to be opposed thereto and capable of moving in the front-rear direction. ing. A primary oil chamber 36 for controlling 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 forward, and the holding position of the V-belt 33 changes to the outer peripheral side, The effective pulley diameter of the primary pulley 31 increases. Conversely, when the hydraulic pressure is released, the hydraulic oil in the primary oil chamber 36 is drained and the primary pulley 31
Effective pulley diameter becomes smaller. That is, the effective pulley diameter of the primary pulley 31 can be freely changed by supplying and discharging hydraulic pressure or hydraulic oil to and from the primary oil chamber 36.

The secondary pulley 32 also has basically the same configuration 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 to face this
And 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
From S, a hydraulic pressure corresponding to the gear ratio to be set (hereinafter referred to as a primary hydraulic pressure) is applied to the primary oil chamber 36. On the other hand, the secondary oil chamber 39 basically has V
A hydraulic pressure sufficient to appropriately hold the tension of the belt 33, that is, a minimum hydraulic pressure capable of transmitting a driving force without causing a belt slip (hereinafter, referred to as a secondary hydraulic pressure) is applied. That is, in the CVT 10,
The gear ratio is determined by the primary hydraulic pressure, and the belt tension is determined by the secondary hydraulic pressure. As described later, since the line pressure of the hydraulic mechanism FS is introduced into the secondary oil chamber 39, the secondary oil pressure is substantially the same as the line pressure.

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

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

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

In this hydraulic mechanism FS, the hydraulic oil discharged from the oil pump 19 is first adjusted to a predetermined line pressure by a line pressure adjusting valve 41, and the line L1
(Hydraulic passage) to the secondary oil chamber 39 and to the clutch valve 46 through the line L2. The clutch valve 46 adjusts the oil pressure in the line L2 to a predetermined pressure by the clutch duty solenoid 54, and supplies the adjusted oil pressure to the manual valve 47 and the lockup valve 49 through 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 speed ratio control valve 43, the speed 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 correction valve 45, and the hydraulic correction valve 45 is opened and closed according to this hydraulic pressure. Hydraulic pressure is line pressure adjustment valve 4
It is introduced as one pilot pressure so that a desired line pressure can be obtained. Note that the hydraulic correction valve 45
, 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 a first duty solenoid 51,
3 is supplied to the primary oil chamber 36 via the fixed gear ratio valve 44. The gear ratio fixed valve 44 is controlled by a first on / off solenoid 52. When the first on / off solenoid 52 is in an on state, the line L7 connected to the primary oil chamber 36 communicates with the line L6, and when it is in an off state. The communication is interrupted. 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 regardless of the operation of the gear ratio control valve 43, and thereby the gear ratio is fixed. It has become.

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

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

The lock-up valve 49 is controlled by a second on / off solenoid 55.
A line L16 connected to the front hydraulic pressure 17f communicates with the relief valve 48 via a 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, a 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 microcomputer
U is provided. The control unit CU includes a shift position signal (P, R, N, D, 2, 1) detected by a shift position sensor 62 and a primary pulley 31 detected by a primary rotation speed sensor 63.
, 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. The information is input as control information.

The control unit CU controls the transmission CT comprehensively, including the target line pressure value setting means, the target line pressure control means, the target line pressure value correction means and the operation cycle control means described in the claims. This device outputs a predetermined control signal to each of the solenoids 51 to 55 and the like based on the various types of control information to perform predetermined control. The following describes a line related to the gist of the present application. Only the pressure control will be described.

Hereinafter, a control method of the line pressure control by the control unit CU will be described with reference to FIGS. 2 to 4 as appropriate according to the flowchart shown in FIG. When the control is started, first, in step # 1, various control information such as the throttle opening TVO, the engine speed Ne, the primary speed Np, and the secondary speed Ns are read.
Subsequently, in step # 2, the engine speed Ne is increased to a predetermined value a.
Whether exceeds ₁ are compared and judgment, a predetermined value Ne> a ₁ a is judged to be the engine speed Ne further in Step # 3 If is set to a value larger than the above a ₁ a
It is determined whether or not 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 it is ₁ , after steps # 4 and # 5 are executed, step #
11 step # 14 is executed, Ne> if a ₂ step # 6, the steps # 11 to step # 14 after the # 7 is performed is executed, step # if a ₁ <Ne ≦ a ₂ After steps # 8, # 9, and # 10 are executed, steps # 11 to # 14 are executed.

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

For example, as shown in FIG. 10, during acceleration when the throttle opening increases to a predetermined value in time τs, a change in line pressure of the hydraulic mechanism FS with respect to a change in throttle opening is accompanied by a response delay of τa. On the other hand, a change in engine torque is accompanied by a response delay of τb, for example, as shown in FIG. Here, the hydraulic response delay time τa
Is almost independent of the engine speed as shown by the straight line J ₁ (calculation frequency 40 Hz) in FIG. 12, but the engine torque response delay time τb is increased as shown by the straight line J ₂ in FIG. It becomes shorter with. Where τa
= The τb engine speed are a _1. When the cycle of the line pressure control by the control unit CU is shortened (frequency is increased), for example, as shown by a straight line J ₁ ′ (for example, a calculation frequency of 80 Hz) in FIG. Therefore, the engine speed a ₁ ′ in which τa = τb shifts to the high rotation side.

[0038] In the present embodiment, since Ne ≦ a hydraulic response delay time is ₁ area τa is shorter than the engine torque response delay time .tau.b, for example throttle opening as shown in FIG. 13 (K ₁₎ is changed Even during acceleration, the change in line pressure (K ₂ ) does not lag behind the change in engine torque (K ₃ ). Therefore, in the case of Ne ≦ a _1, and then to set the target line pressure according to the actual engine torque (i.e. the input torque to the CVT 10). That is, a substantially necessary minimum line pressure is set, and the pump loss is reduced.

[0039] Since the hydraulic pressure response delay time τa is longer than the engine torque response delay time τb in contrast Ne> a ₁ region, for example, the throttle opening as shown in FIG. 14
At the time of acceleration when (K ₁ ') changes, the line pressure changes
(K ₂ ′) cannot follow the change in the engine torque (K ₃ ′). For this reason, the hatched portion in FIG. 14 indicates the line pressure corresponding to the actual engine torque, that is, the required line. Only a line pressure lower than the pressure is obtained, and belt slip occurs.

[0040] Therefore, in the case of Ne> a _1, is basically the actual larger than the engine torque virtual line圧演Arabic of engine torque (hereinafter, referred to as line圧演Arabic torque it) to set the The target line pressure is set based on the line pressure calculation torque so that the target line pressure 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 becomes a problem, the line pressure target value is added for safety. In this case, as shown in FIG. 15, at the time of acceleration in which the throttle opening (K ₁ ") changes, the line pressure (K ₂ ") is increased from the beginning, and therefore, as in the example shown in FIG. No trouble occurs at all, and no belt slip occurs.

[0041] However, even in a region where a Ne> a _1, since the engine speed Ne is the difference in the hydraulic pressure response delay time τa and the engine torque response delay time τb increases as time _{high, 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, and in the region where a ₁ <Ne ≦ a ₂ , the line pressure calculation torque is gradually increased in accordance with the increase in the engine speed Ne (see FIG. 7).

The specific calculation method of the line pressure calculation torque is as follows. That is, in the case of Ne ≦ a _1, in step # 4, for example, by using a map having the characteristics shown in FIG. 6, the engine torque Tr corresponding to the throttle opening TVO and the engine speed Ne is calculated, step In # 5, this engine torque Tr is used as the line pressure calculation torque Tin.

[0043] Ne> For a _2, a throttle opening degree T by using a map having the characteristics shown in FIG. 6 at step # 6
The engine torque Tmax corresponding to the case where VO is 100% (curve G ₁ ) is calculated, and this engine torque Tmax is used as the line pressure calculation torque Tin in step # 7.

If a ₁ <Ne ≦ a ₂ , first, step # 8
Then, the actual engine torque Tr is calculated in the same manner as in step # 4. Subsequently, in step # 9, the engine torque Tmax corresponding to TVO = 100% is calculated in the same manner as in step # 6. And then step #
In step 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 ) Equation 1 Line pressure calculation torque Tin set in this manner FIG. 7 shows the characteristics of the engine with respect to the engine speed Ne, for example.

After the line pressure calculation torque Tin is calculated in accordance with the engine speed Ne in this way, in any case, first in step # 11, a map having characteristics as 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
Is multiplied by a predetermined safety factor Ks, then a duty ratio corresponding to the target line pressure Ps is calculated in step # 13, and then the first duty is calculated in step # 14 according to the duty ratio. Solenoid 5
3 is driven and controlled so that the line pressure follows the target line pressure.

When such line pressure control is performed,
The target line pressure, Wachihobo actually formed by the line pressure characteristics for the engine rotational speed Ne is, for example, as a polygonal line H ₁ in FIG. Linear H ₂ in FIG. 9 is a line pressure is actually needed (required line pressure), the difference d of an H ₁ and H ₂ is that the line pressure, Wachi safety factor to be added for the purpose of safety. As is clear from FIG. 9, according to the line pressure control, the additional line pressure d, that is, the safety factor, is set to a necessary minimum according to the change in the engine speed Ne, that is, the response of the engine torque. Therefore, since the line pressure is not unnecessarily increased, the pump loss is reduced and the fuel efficiency is improved.

As described above, according to the present embodiment, the necessary line pressure can be reliably ensured without unnecessarily increasing the line pressure (safety factor). The calculation cycle of the line pressure control is set shorter as the engine speed is higher.
If the engine frequency is increased, the engine torque response delay time and the hydraulic response delay time both become shorter, so it is necessary to increase the safety factor according to the increase in the engine speed. And the safety factor can be set smaller.

[0049]

In general, when the engine speed is high, the response delay of the engine torque to the change in the engine load (change in the throttle opening) becomes short, so that the change in the line pressure is later than the change in the engine torque. The required line pressure may not be obtained. However, according to the first aspect, the higher the engine speed is, the more the line pressure target value is corrected to the higher pressure side, so that the required line pressure is ensured regardless of the engine speed. Also, when the engine speed is low, the line pressure target value is set low, so that the line pressure is not unnecessarily increased, the pump loss is reduced, and the fuel efficiency is improved. That is, the safety factor for setting the line pressure is set according to the response of the engine torque, and the necessary line can be secured without setting the line pressure unnecessarily high.

According to the first invention, in a region where the hydraulic response is slower than the engine torque response, the line pressure target value is shifted to the high pressure side within a range up to the line pressure corresponding to the maximum engine torque. Since the correction is made, the required line pressure is reliably ensured at the time of high rotation.

According to the second aspect, the higher the engine speed, the shorter the calculation cycle is set, and the faster the hydraulic response. Therefore, when the engine speed is high, the engine torque response delay time and the oil pressure response delay time are both short, so that it is not necessary to increase the safety factor in accordance with the increase in the engine speed. Will be able to

According to the third aspect, basically, the same operation and effect as those of the first or second aspect can be obtained. Furthermore, in the region where the hydraulic response is faster than the engine torque response,
Since the line pressure target value is set to the normal value, the line pressure is not set unnecessarily high, and the pump loss is further reduced.

According to the fourth aspect, basically, the same operation and effect as those of the third aspect can be obtained. Further, 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 the drawings]

FIG. 1 is a block diagram showing a configuration of first to fourth inventions corresponding to claims 1 to 4;

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 illustrating a control method of line pressure control.

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

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

FIG. 8 is a diagram illustrating characteristics of a target line pressure with respect to a gear ratio and a torque for calculating a line pressure.

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

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

FIG. 11 is a diagram showing engine torque response to a change in throttle opening.

FIG. 12 is a diagram showing characteristics of a hydraulic 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 response and an engine torque response at low rotation in conventional line pressure control.

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

FIG. 15 is a diagram illustrating a line pressure control according to the present invention;
FIG. 4 is a view similar to FIG.

[Explanation of symbols]

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

────────────────────────────────────────────────── ─── Continued on the front page (72) Inventor Yuji Mori 3-1, Shinchi, Fuchu-cho, Aki-gun, Hiroshima Mazda Co., Ltd. (56) References JP-A-62-9054 (JP, A) (58) Survey Field (Int.Cl. ⁷ , DB name) F16H 59/00-61/12 F16H 61/16-61/24 F16H 63/40-63/48 F16H 9/00

Claims (4)

(57) [Claims] 1. A line pressure target value setting for setting a line pressure target value of a hydraulic mechanism in accordance with an output torque of an engine in a hydraulic control device of a hydraulically operated transmission provided with a transmission mechanism operated by the hydraulic mechanism. Means, a line pressure control means for controlling the line pressure so as to follow the line pressure target value, and correcting the line pressure target value to a higher pressure side as the engine speed increases according to the engine speed. And a line pressure target value correcting means for controlling a line pressure in a rotation region in which the response of the hydraulic mechanism to a change in the engine load is slower than the response of the engine to a change in the engine load. An oil characterized in that the target value is corrected to a high pressure side within a range up to a hydraulic pressure set when the output torque of the engine is maximum. Hydraulic control device for pressure operated transmission. 2. A line pressure target value setting for setting a line pressure target value of a hydraulic mechanism in accordance with an output torque of an engine in a hydraulic control device of a hydraulically operated transmission provided with a transmission mechanism operated by the hydraulic mechanism. Means, a line pressure control means for controlling the line pressure so as to follow the line pressure target value, and the calculation cycle of the line pressure target value setting means and the line pressure control means is shortened as the engine speed increases. A hydraulic control apparatus for a hydraulically operated transmission, comprising: 3. A hydraulic control apparatus for a hydraulically operated transmission as claimed in claim 1 or 2, the line pressure target value correcting means, the responsiveness of the hydraulic mechanism with respect to a change in engine load, to changes in engine load A hydraulic control device for a hydraulically operated transmission, wherein a target line pressure value is a normal target value that does not depend on the engine speed in a rotation range faster than the response of the engine. 4. The hydraulic control device for a hydraulically operated transmission according to claim 3 , wherein the transmission mechanism is a belt-type continuously variable transmission, and the line pressure target value setting means is configured to control a hydraulic pressure with respect to a change in engine load. In a rotation region where the responsiveness of the mechanism is faster than the responsiveness of the engine to a change in the engine load, the target line pressure 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 operated transmission, characterized in that: