JPH01269751A - Hydraulic control device of automatic transmission - Google Patents

Abstract

PURPOSE:To ease the speed change shock and improve the durability of a frictionally engaging element by variably controlling the pressure of an oil pump according to the engine load to set a line pressure, and variably controlling the line pressure to a desired pressure at the time of speed change. CONSTITUTION:In speed change, a first pressure regulating means F receives the output of a line pressure correcting means H and operates to maintain the line pressure at a desired high pressure in spite of an engine load E. Thus, the driving torque of an oil pump D is increased, and the engine torque applied to a frictionally engaging element A is reduced so much. Simultaneously with maintaining the line pressure at high pressure in the speed change, a second pressure regulating means G operates to set the fluid pressure supplied to the frictionally engaging element A in a pressure appropriate for operation of the frictionally engaging element A. Hence, the engaging torque charged to the frictionally engaging element A at the time of speed change is reduced, so that the speed change shock is reduced and also the durability of the frictionally engaging element itself is improved.

Description

DETAILED DESCRIPTION OF THE INVENTION (Field of Industrial Application) The present invention relates to a hydraulic control device for an automatic transmission.

(Prior Art) In an automatic transmission that obtains a plurality of gears by engaging and releasing frictional engagement elements using fluid pressure,
Shift shock occurs when changing gears. As a method of alleviating such switching shock, it is known to reduce the engine torque by, for example, changing the ignition timing or reducing the supplied fuel, thereby reducing the torque borne by the frictional engagement elements. There is.

However, this method of reducing engine torque by controlling the combustion state of the engine is undesirable because it complicates engine control and may increase costs or reduce engine reliability.

Conventionally, in hydraulic control devices for automatic transmissions, there has been a technology in which line pressure is controlled by duty-controlling the control signal pressure of a regulator valve (for example, Japanese Patent Laid-Open No. 54-2
349 (see Japanese Patent Publication No. 132-187), technology to control the duty of clutch engagement pressure (for example, Japanese Patent Publication No. 132-187).
No. 80) are known, but no technology has been proposed that relates such a hydraulic control method to a shift shock of an automatic transmission.

(Object of the Invention) The first and second inventions of the present application have been made in view of such conventional problems, and have been made in view of the above-mentioned conventional problems.
In an automatic transmission in which gears are changed by release, the engine torque applied to the frictional engagement element is reduced by relatively simple control, thereby alleviating the shift shock of the automatic transmission and improving the durability of the frictional engagement element. The purpose of this invention is to provide a hydraulic control device that improves the operating characteristics of the frictional engagement element by increasing the responsiveness and stability of the fluid pressure supplied to the frictional engagement element. .

(Means for Achieving the Object) In order to achieve such an object, in the first aspect of the invention, as shown in FIG. 1, the frictional engagement element A is engaged and released by fluid pressure. a transmission gear mechanism B whose power transmission path is switched to multiple stages; an oil pump D driven by the engine C to discharge fluid; and variable control of the pressure of the fluid discharged by the oil pump D according to the engine load E. a first pressure adjusting means F for setting the line pressure;
a second pressure regulating means G that variably controls the line pressure to a desired pressure during gear shifting and supplies it to the frictional engagement element A; and a line pressure that maintains the line pressure at a desired high pressure regardless of the engine load E during gear shifting. The present invention is characterized in that it includes a correction means H.

In a second aspect of the present invention, in the first aspect, the first pressure regulating means F and the second pressure regulating means G are each adjusted so that the pressure regulating level is variably controlled by the control signal pressure. It is characterized in that it is composed of a pressure valve I and a duty solenoid valve J that generates the control signal pressure.

(Function) With this configuration, in the first invention, (1) during gear shifting, the first line pressure correction means receives the output of the line pressure correction means;
(2) The pressure regulating means operates and the line pressure is maintained at a desired high pressure regardless of the engine load, and the driving torque of the oil pump increases, and the engine torque applied to the frictional engagement element decreases accordingly. Simultaneously with maintaining the line pressure at a high level during gear shifting, the second pressure regulating means is operated to set the fluid pressure supplied to the frictional engagement element to a pressure appropriate for the operation of the frictional engagement element.
The following effects can be obtained.

In addition, in the second invention, is the above-mentioned claim 1.
Similar to the invention described above, the line pressure is maintained at a high pressure regardless of the engine load during gear shifting, thereby increasing the driving torque of the oil pump and reducing the engine torque applied to the frictional engagement element. Generally, as in the second aspect of the present invention, the second pressure regulating means is composed of a duty solenoid valve and a pressure regulating valve, and the line pressure is adjusted by the second pressure regulating means and supplied to the frictional engagement element. When the fluid pressure is controlled, the responsiveness and stability of the hydraulic pressure supplied to the frictional engagement elements improve as the line pressure increases. When the line pressure is maintained at a high pressure by the means, an effect is obtained in which the responsiveness and stability of the fluid pressure supplied to the frictional engagement element are improved all the time.

(Effects of the Invention) Therefore, in the first aspect of the invention, (1) the engine torque borne by the frictional engagement element during gearshifting is reduced, which reduces shift shock and improves the durability of the frictional engagement element itself; (2) Since the fluid pressure supplied to the frictional engagement element during gear shifting is set to the appropriate pressure required for operation regardless of the increase in line pressure, its operating characteristics will not be impaired. Effects such as , etc. can be obtained.

In addition, in the invention of claim 2, the above claim 1
The first effect (1) in the invention is obtained, and
During gear changes, the fluid pressure supplied to the frictional engagement elements is highly responsive and stable, resulting in high-level operating characteristics such as responsive and stable operation of the frictional engagement elements. effective.

(Example) Hereinafter, a hydraulic control device for an automatic transmission to which a first aspect of the invention is applied will be described with reference to FIGS. 2 to 8.

Basic configuration of transmission FIG. 2 is a skeleton diagram of an automatic transmission in which a hydraulic control device according to an embodiment of the invention of claim 1 is incorporated, and in the figure, reference numeral 1 indicates an engine output shaft, and 2 indicates a torque converter. , 3 are speed change gear mechanisms, which are arranged coaxially in sequence. Torque converter 2 includes a pump 4, a turbine 5, and a stator 6. The pump 4 is fixed to the engine output shaft 1, and the stator 6 is
It rotates via a one-way clutch 7 on a fixed shaft 9 that is integrated with the case 8 of the speed change gear mechanism 3. The one-way clutch 7 allows the stator 6 to rotate in the same direction as the pump 4, but does not allow the stator 6 to rotate in the opposite direction.

The speed change gear mechanism 3 includes a solid shaft 10 whose base end is fixed to the engine output shaft 1 and extends through the center of the speed change gear mechanism 3. This solid shaft 10 is for driving an oil pump D arranged on a side wall of the speed change gear mechanism 3, and its tip is connected to the oil pump D.

Furthermore, this solid shaft! A hollow turbine shaft 11 whose base end is connected to the turbine 5 of the torque converter 2 is provided on the outside of the torque converter 2 .

Further, a Ravigneau-type planetary gear unit 12 is provided on the turbine shaft ll. The planetary gear unit 12 includes a small-diameter sun gear 13, a large-diameter sun gear 14 disposed on the side of the small-diameter sun gear 13 far from the engine, a long pinion gear 15, a front pinion gear 16, and a ring gear 17. .

A forward clutch 18 and a coast clutch 19 are located on the side of the planetary gear unit 12 that is far from the engine.
are arranged in parallel. This forward clutch 1
Reference numeral 8 denotes a forward traveling clutch, which connects and disconnects power transmission between the small diameter sun gear 13 and the turbine shaft 11 via the first one-way clutch 20. On the other hand, the coast clutch 19 is connected in parallel with the forward clutch 18 and connects and disconnects power transmission between the small diameter sun gear 13 and the turbine shaft If. A 2-4 brake 21 is arranged radially outward of the coast clutch 19. The 2-4 brake 21 is configured as a band brake, and has a brake drum 21a connected to the large-diameter sun gear 14 and a brake band 21b hooked onto the brake drum 21a. radially outward of the forward clutch 18 and the 2-
A reverse clutch 22 is arranged on the side of the four brakes 21. This reverse clutch 22 is a clutch for backward traveling, and performs intermittent transmission of power between the large-diameter sun gear 14 and the turbine shaft 11 via the brake drum 21a of the 2-4 brake 21.

On the radially outward side of the planetary gear unit 12,
A low reverse brake 23 is arranged to engage and disengage the carrier 12a of the planetary gear unit 12 and the case 3a of the speed change gear mechanism 3.

A second one-way latch 24 is arranged between the 2-4 brake 21 and the low reverse brake 23 in parallel with the low reverse brake 23 to engage and disengage the carrier 12a and the case 3a.

On the ennon side side of the planetary gear unit 12, there is a 3.
-4 clutches are arranged. An output gear 26 connected to a ring gear 17 is arranged on the side of the 3-4 gear 25 on the engine side.
This is a lock-up clutch for directly connecting the torque converter 2 without going through the torque converter 2.

Functions of the Speed Change Gear Mechanism 3 The speed change gear mechanism 3 having the structure described above has itself four forward speeds and one reverse speed, and each friction engagement element, that is, each clutch 1B, 19.21. 25 and each brake 2
By operating 1.23 appropriately, the required gear position can be obtained. In the above configuration, the operational relationship between each gear stage, clutch, and brake is shown in Table 1 below. In addition,
Of each clutch and brake, only the actuator for the 2-4 brake 21 has two oil chambers, one on the apply side and one on the release side, as described later, to supply hydraulic pressure to the apply side and release hydraulic pressure on the release side. 2- only when
The 4 brake 2I is engaged, and the 2-4 brake 21 is released in other hydraulic pressure supply modes. The remaining clutch and brake actuators each have only one oil chamber, and are engaged when hydraulic pressure is supplied to this oil chamber, and are opened when hydraulic pressure in this oil chamber is released. be done.

Hydraulic Circuit Next, the hydraulic circuit for the automatic transmission shown in FIG. 1 will be explained with reference to FIG.

■Manual valve In Fig. 3, reference numeral 41 is a manual valve, and as is known, P and R can be controlled manually.

Six range positions are possible: N, D, and 2.1. This manual valve 31 includes a.

It has boats c, e, f, and g, and the oil pressure generated by the oil pump D is regulated by a pressure regulating valve 33 connected to the oil passage lO1, and is supplied to boat g as line pressure. be done. The range position of this manual valve 31 and other boats a and C are communicated with boat g to which line pressure is supplied.

The relationship between e and f is as follows.

P range: No boats communicated R range: Boat r only N range 2 No boats communicated D range: Boats a and C 2 range: Boats a and C Runge: Boats a and C ■Torque converter duty solenoid valve above A part of the hydraulic pressure generated by the oil pump D passes through the oil passage 102 to the solenoid reducing valve 34.
The pressure is reduced to a predetermined pressure (for example, 4.0 Kg/am") at the solenoid reducing valve 34.Furthermore, the hydraulic pressure exerted by the solenoid reducing valve 34 is transmitted to four duty units, which will be described later. The pressure is regulated by solenoid valves 35 to 38, which include main pressure regulating valves 33, 3-4, pressure regulating valve 39, and reverse pressure regulating valve 40, respectively.
, is supplied to the servo pressure regulating valve 41 as a control signal pressure for pressure regulation level control.

That is, the hydraulic pressure regulated by the first duty solenoid valve 35 is supplied to the back pressure side of the main pressure regulating valve 33 via the oil passage 103 as its control signal pressure. The main pressure regulating valve 33 then generates a line pressure corresponding to the magnitude of this control signal pressure. In this case, the line pressure is normally controlled to rise or fall by changing the duty ratio of the first duty solenoid valve 35 in response to changes in the engine load (throttle valve opening) (i.e., the sixth duty solenoid valve 35). In particular, in this embodiment, the engine is controlled based on the characteristic diagram shown by the solid line curve σ1 in the figure. The pressure is maintained at a desired high level regardless of the load (for example, the line pressure characteristics are controlled based on the characteristic diagram shown with broken lines in FIG. 6). That is, in this embodiment, the main pressure regulating valve 33 and the first duty solenoid valve 35 constitute the first pressure regulating means in the first and second aspects of the invention.

The hydraulic pressure regulated by the second duty solenoid valve 36 is transferred to the 3-4 pressure regulating valve 39 via the oil passage 104.
The pressure is also supplied to the reverse pressure regulating valve 40 via an oil passage 104a branching from the oil passage 104 as a control signal pressure for setting a pressure regulation level. Of the reverse clutch 22 and the 3-4 clutch 25 whose supply oil pressure is adjusted by these two pressure regulating valves 39 and 40, the 3-4 clutch 2
Reference numeral 5 is related to the shift operation in the forward gear, and this second duty solenoid valve 36 and the 3-4R pressure valve 3
9 corresponds to the first pressure regulating means in the first and second aspects of the invention.

The oil pressure regulated by the third duty solenoid valve 37 is applied to a servo pressure regulating valve 41 provided in an oil passage communicating with the release side oil chamber 28a of the 2-4 brake actuator 28 via an oil passage 105. It is supplied to the back side of the spool as a control signal pressure for pressure regulation level adjustment. Further, this regulated hydraulic pressure on the release side is supplied to the back side of the spool of the coast control valve 43 as a control signal pressure of the coast control valve 43 via the oil passage 107. Further, the regulated oil pressure on the release side is also used as a forward clutch pressure that is switched by the forward control valve 44 via an oil passage 107a branched from the oil passage 107. Note that in this embodiment, the third duty solenoid valve 37
and the servo pressure regulating valve 41 constitute the first pressure regulating means according to the first and second claims.

The regulated pressure adjusted by the fourth duty solenoid valve 38 is supplied to the lock-up control valve 42 via the oil passage 108 as its control signal pressure.

The duty ratio of each of these duty solenoid valves 35 to 38 is controlled to a predetermined duty ratio determined in advance in accordance with the operating time so that each clutch or brake can be engaged with as little shock as possible.

In addition, Table 2 below shows each gear shifting operation and the 2-4 brake 21.
The correlation between this and the operating pressure control of the 3-4 clutch 25 is shown.

The O mark in Table 2 indicates that duty control of the operating pressure is performed.

■Clutch and brake actuators Two of the clutch and brake actuators for shifting gears
-4 Except for the brake actuator 28, the other actuators are of a type that is simply engaged when hydraulic pressure is supplied, so the actuators are shown in Fig. 3 using the same symbols as the clutches and brakes. There is.

2-4 Brake actuator 28 is a piston 28
It has a release side oil chamber 28a and an apply side oil chamber 28b partitioned by c. And this piston 2
8c is integrally provided with a piston rod 28d connected to the band 21b of the 2-4 brake 21.

This piston 28c is moved to the third position by a spring 28e.
It is biased downward in the figure. The 2-4 brake actuator 28 operates only when the condition that line pressure is supplied to the apply side oil chamber 28b and the oil pressure of the release side oil chamber 28a is released is satisfied.
Brake 21 is engaged. In other words, even if line pressure is supplied to the apply side oil chamber 28b, when line pressure is supplied to the release side oil chamber 28a, the 2-4 brake 21 is open and the release side oil chamber 28b is supplied with line pressure. By adjusting the oil pressure of 28a with the servo pressure regulating valve 41,
-4 The engagement force of the brake 21 is adjusted.

The forward clutch 18 is connected to the boat a of the manual valve 31 via an oil passage 109, the forward control valve 44, and an oil passage 110.

The coast clutch 19 (actuator) is connected to the boat C via the oil passage ill, the coast control valve 43, oil passages 112 and 113, the coast exhaust valve 45, and the oil passage 114.

Additionally, the switching valve 4 is connected to the coast control valve 43.
6, line pressure can also be supplied through the oil passage 115 extending from the boat e.

3-4 clutch 25 (actuator) is connected to oil passage 1
16, Boat C via the above 3-4 pressure regulating valve 39, oil passage 117. 2-3 shift valve 47, oil passage 118, l14
It is connected to

The low reverse brake (actuator) 23 is connected to the boat r via an oil passage 119 and a switching valve 48. Further, the switching valve 48 is connected to the boat e via the oil passage 120.1-2, the lift valve 49, the oil passage 121, and the low redening valve 50.

The reverse clutch 22 (actuator) is connected to the boat f via an oil passage 122, the reverse pressure regulating valve 40, and an oil passage 123.

The apply side oil chamber 28b of the 2-4 brake actuator 28 is connected to the boat a via an oil passage 124, an l-2 shift valve 49, and an oil passage 125. The oil chamber 28a on the release side includes the oil passage 106, the servo pressure regulating valve 41. It is connected to boat a via oil passages 126 and 125.

Next, to explain the actual control of this hydraulic control device, as mentioned above, this hydraulic control device adjusts the line pressure using the main pressure regulating valve 33 in accordance with the engine load during non-shift driving, but when shifting, the line pressure is adjusted by the engine load. Regardless of the size of the load, the line pressure is raised to a predetermined high pressure and held until the end of the shift, and at the end of the shift the line pressure is adjusted again according to the engine load. (See Figure 6). In this way, by maintaining the line pressure at a high pressure during gear shifting, the oil pump is driven by the engine, and its driving torque changes proportionally to the discharge pressure, so the line pressure The proportion of the pump driving torque in the engine torque increases by the amount of increase in , and the torque borne by each frictional engagement element during gear shifting decreases accordingly. As a result,
Switching shock (shift shock) in each frictional engagement element
This has the advantage that the damage is reduced and the durability of these materials is improved accordingly.

This line pressure adjustment control will be explained using an example of shifting from 1st to 2nd speed with reference to FIGS. 3 to 6. In the control flowchart of FIG. 4, first, after initialization (step S1) , it is determined whether or not there is a need for gear shifting (step S).

That is, the current turbine rotation speed (T
sp) and the current throttle valve opening (TVO), and based on the map, it is determined whether it is necessary to shift from 1st gear to 2nd gear.

Here, if the shift command from 1st gear to 2nd gear has not been issued, the flag F indicating shifting is O (that is, not shifting) (step S4), so in this case, the throttle valve is opened. Set line pressure based on degree TVO (step S
S). Specifically, the line pressure is set by duty control of the first duty solenoid valve 35 based on the characteristic diagram shown in FIG.

On the other hand, Step S! When a shift command is issued from 1st gear to 2nd gear, the line pressure is related to the throttle valve opening (engine load) during the shift period from the time the shift command is output until the end of the shift, as shown in Figure 5. An operation is performed to maintain the pressure at a high pressure (as shown by the characteristic diagram U in FIG. 6).

That is, first, in step S6, the theoretical rotational speed '1tHsp of the turbine at the end of the shift is determined by calculation. Next, by controlling the duty of the first duty solenoid valve 35, the pressure regulating valve 33 is operated to set the line pressure to MAX (step S7), and the 2-4 brake 21, which is involved in the shift from 1st gear to 2nd gear, is activated. The release pressure of the actuator 28 is gradually lowered at a predetermined speed by duty control of the third duty solenoid valve 37 (step S). Next, set flag F to 1 (step S,
), and maintain the line pressure at a high pressure until the turbine rotation speed Tsp falls below the theoretical turbine rotation speed TTI (Sp) at the end of the shift (until the shift is completed) (steps S, , S
,,S,,S,,S7). When the gear shift is completed, the flag F is set to O (step S) to prepare for the next gear shift command. FIG. 5 shows the correlation between the engine speed, line pressure, and wheel drive torque (output shaft torque of the automatic transmission) during this shift. In the wheel drive torque characteristic diagram in Fig. 5, the shaded area shows the decrease in wheel drive torque due to maintaining the line pressure at a high pressure and increasing the oil pump drive torque. .

Furthermore, if the line pressure is maintained at a high pressure during gear shifting as in this embodiment, the wheel drive torque will fluctuate f before and after shifting.
fi ml + n I are each smaller than the conventional variation amount m1゜n, and the shift shock is alleviated accordingly.

In addition, since the above explanation takes the example of shifting from 1st gear to 2nd gear,
Only the release pressure of the 2-4 brake actuator 28 is duty-controlled, but when shifting from 2nd to 3rd speed, the clutch pressure of the 3-4 clutch 25 is also duty-controlled (see Table 2 above). .

On the other hand, as in this embodiment, the release pressure of the 2-4 brake actuator 28 and the clutch pressure of the 3-4 clutch 25 are controlled by duty solenoid valves 37 and 36, respectively.
When the pressure is controlled by the pressure regulating valve 39 and 41, the line pressure is maintained at a higher pressure than normal during gear shifting, so that the 2-4 brake 21 Also, the operating characteristics of the 3-4 clutch 25 are improved.

Figures 7 and 8 show the 2-4 brake actuator 2.
There are two cases in which the 2-4 brake 21 is released by lowering the release pressure of No. 8 (increasing the duty ratio of the third duty solenoid valve 37), and another case in which the release pressure is increased (increasing the duty ratio of the third duty solenoid valve 37). The graph shows the results of experiments on the change in release pressure when the 2-4 brake 21 is engaged (lowering the ratio), both at high in-pressure and at low in-in pressure.

Observing these two figures, first of all, in Fig. 7, the delay time from the start of duty control of the third duty solenoid valve 37 until the release pressure actually changes and becomes stable is longer than when the in-pressure is high ( Similarly, in FIG. 8, the delay time tb of the release pressure when the in-pressure is high is shorter than the delay time (tag) when the in-pressure is low (delay time Lay).

It can be seen that the delay time tb of the release pressure is shorter than the release pressure delay time tb when the lining pressure is low. This means that by maintaining the line pressure at high pressure during gear shifting, the responsiveness and stability of the release pressure will increase, and this is due to the fact that
-4 This means that the operational response and stability of the brake 21 are improved. The effect of improving operating characteristics is 2.
The same is true for the 3-4 clutch 25 whose clutch pressure, like the -4 brake 21, is controlled by the second duty solenoid valve 36 and the 3-4F] pressure valve 39.

That is, in this embodiment, the release pressure of the 2-4 brake 21 and the clutch pressure of the 3-4 clutch 25 are adjusted by a pressure regulating means consisting of a combination of a duty solenoid valve and a pressure regulating valve. The operating characteristics are automatically improved in connection with maintaining the line pressure at a high level during gear shifting.

Note that the present invention is not limited to the above configuration (for example, a variable orifice may be used instead of the duty solenoid valve 37 for hydraulic control of the 2-4 brake 21,
It goes without saying that the present invention can also be applied to a brake system in which the apply pressure of the 2-4 brake 2I is applied using an accumulator, as disclosed in Japanese Patent Application Laid-Open No. 62-35153.

[Brief explanation of the drawing]

FIG. 1 is a diagram corresponding to claims of the first and second claims, FIG. 2 is a skeleton diagram of an automatic transmission equipped with a hydraulic control device according to an embodiment of the first invention, and FIG. is the hydraulic circuit diagram for the automatic transmission shown in Figure 2, and Figure 4 is the hydraulic circuit diagram for the automatic transmission shown in Figure 3.
Control flowchart diagram of the hydraulic control device shown in Fig. 5.
The figure is a time chart of engine speed etc. during gear shifting, Figure 6 is a line pressure characteristic diagram, and Figures 7 and 8 are 2-
FIG. 4 is a change characteristic diagram of release pressure of four brakes. l... Engine output shaft 2... Torque converter 3... Speed change gear mechanism 4... Pump 5... Turbine 6. φ... Stator 21... ... 2-4 brake 22 ... Reverse clutch 23 ... Low reverse brake 25 ... 3-4 clutch 27 ... Lock-up clutch 28 ... 2-4 brake actuator 31・・・
...Manual valve 33...Pressure regulating valve 35-38...Duty solenoid valve 39-4
1...Pressure regulating valve time

Claims (1)

[Scope of Claims] 1. A speed change gear mechanism in which a power transmission path is switched to multiple stages by engaging and releasing frictional engagement elements using fluid pressure, and an oil pump that is driven by an engine to discharge fluid; a first pressure adjusting means for variably controlling the pressure of the fluid discharged by the oil pump according to the engine load to set line pressure; What is claimed is: 1. A hydraulic control device for an automatic transmission, comprising: a second pressure adjusting means for supplying a pressure to the engine; and a line pressure correcting means for maintaining the line pressure at a desired high pressure regardless of engine load during gear shifting. 2. The first pressure regulating means and the second pressure regulating means are characterized in that they include a pressure regulating valve whose pressure regulation level is variably controlled by a control signal pressure, and a duty solenoid valve that generates the control signal pressure. A hydraulic control device for an automatic transmission according to claim 1.