JP2002267007A - Hydraulic control device of automatic transmission - Google Patents

Abstract

PROBLEM TO BE SOLVED: To secure the running ability of a vehicle stably by guaranteeing the supply of the base pressure to a solenoid valve installed for supplying the oil pressure to a hydraulic servo of a frictional engagement element for attaining shift ranges. SOLUTION: This hydraulic control device of the automatic transmission is equipped with a control valve installed in a supply passage of the hydraulic servo for operating the frictional engagement element and controlling the oil pressure supplied to the hydraulic servo, the solenoid valve to impress a signal pressure for the control operation on the control valve, and a pressure control valve 54 installed between the line pressure oil passage and the solenoid valve, for adjusting the line pressure and supplying the base pressure for output of the signal pressure to the solenoid valve. Parallel with the control valve, a pressure control circuit is installed which is equipped with a pressure control valve 55 for communicating the line pressure oil passage L1 with the solenoid valve (connected with oil passages L5 and L6). This guarantees the supply of the base pressure to the solenoid valve by one control valve in case of failure of the other valve.

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 an automatic transmission mounted on a vehicle, and more particularly to a technique for guaranteeing the operation of the hydraulic circuit.

[0002]

2. Description of the Related Art In recent years, automatic transmissions for vehicles include a hydraulic circuit that controls clutches and brakes (hereinafter, these are collectively referred to as friction engagement elements) for achieving each shift speed. For each hydraulic servo that operates each frictional engagement element, a dedicated control valve for hydraulic supply control and each solenoid valve (linear solenoid valve) for individually applying a signal pressure for control operation to the control valve Or, a duty solenoid valve) is provided and independently controlled to improve controllability.

[0003] Each of the solenoid valves in such a hydraulic circuit is a normally-open type valve which enters a control operation of reducing an output signal pressure from a fully opened state by application of an electric signal, and is applied with an electric signal applied for the control operation thereof. Even when the signal fails, the output of the signal pressure to the control valve is maintained, and the control valve is actuated to supply hydraulic pressure to the hydraulic servo, thereby ensuring a prima facie running capability of the vehicle. However, even if the supply of the base pressure to the solenoid valve is a normally-open type valve, the output of the signal pressure is naturally impossible. When the output of the solenoid valve becomes impossible as described above, if the control valve that operates by applying a signal pressure from each solenoid valve is a normally-return valve of a spring return type, the solenoid valve that has been supplied with the base pressure is disconnected. Since the supply of hydraulic pressure to the hydraulic servo concerned cannot be performed, there is a concern that the running ability of the vehicle cannot be secured due to the inability to control the corresponding friction engagement element.

[0004]

As described above, the reason why the supply of the base pressure to the solenoid valve is interrupted is that the solenoid valve is disposed between the line pressure oil passage of the hydraulic circuit and the solenoid valve to supply the base pressure. There is a failure due to a stick of a modulator valve as a pressure regulating valve inserted for adjusting the gain of the valve. In general, the modulator valve is a spring-return type normally open type pressure reducing valve, so it is difficult to cause a stick in a closed state, but even if it is extremely rare, if a stick in a closed state occurs, As described above, a situation occurs in which the running ability of the vehicle cannot be secured.

Accordingly, the present invention guarantees the supply of a base pressure to a solenoid valve related to the supply of hydraulic pressure to a hydraulic servo of a friction engagement element that achieves a gear position, thereby stabilizing the traveling performance of a vehicle. It is a general object to provide a hydraulic control device for an automatic transmission which can ensure the above.

[0006]

SUMMARY OF THE INVENTION In order to achieve the above object, the present invention provides a control valve for controlling a hydraulic pressure supplied to a hydraulic servo, which is inserted into a supply oil passage of a hydraulic servo for operating a friction engagement element. A solenoid valve for applying a signal pressure for control operation to the control valve, and a solenoid valve interposed between the line pressure oil passage and the solenoid valve for regulating the line pressure and outputting a signal pressure to the solenoid valve. In a hydraulic control device for an automatic transmission having a pressure regulating valve for supplying a base pressure, a main pressure regulating circuit for connecting a line pressure oil passage and a solenoid valve in parallel with the pressure regulating valve is provided. And

In the above configuration, it is effective that the pressure regulating circuit includes a second pressure regulating valve for regulating the line pressure and supplying a base pressure for outputting a signal pressure to the solenoid valve.

Alternatively, in the above configuration, the pressure regulating circuit may have a configuration including an orifice for reducing the line pressure.

More specifically, the present invention provides at least one
Hydraulic servo of a first system for operating a friction engagement element capable of achieving one forward travel stage and its supply oil passage, and a hydraulic servo of a second system for operating a friction engagement element capable of achieving another forward travel stage A hydraulic servo and its supply oil passage are provided in parallel, and a control valve for individually controlling the supply of the hydraulic pressure to each hydraulic servo is inserted in each of the supply oil passages. In a hydraulic control device for an automatic transmission in which a dedicated solenoid valve for applying a signal pressure for the control operation is applied, a signal pressure is applied to control valves of a line pressure oil passage and a supply oil passage of the first system. A first pressure regulating valve is interposed between the solenoid valve and a solenoid valve that applies a signal pressure to a control valve of the supply oil passage of the second system. Characterized in that the pressure regulating valve is inserted.

[0010]

According to the structure of the first aspect,
The supply of the base pressure to the solenoid valve by the pressure regulating circuit parallel to the pressure regulating valve is ensured even when the pressure regulating valve fails. Therefore, by maintaining the signal pressure output of the solenoid valve by supplying the base pressure from the parallel pressure regulation circuit, the hydraulic servo can be operated by the control valve, and the traveling of the vehicle is secured by the engagement of the corresponding friction engagement element. Is done.

Next, according to the second aspect of the invention, when one of the first and second pressure regulating valves fails, the supply of the base pressure to the solenoid valve by the other pressure regulating valve is maintained. Therefore, according to this configuration, it is possible to guarantee the supply of the base pressure to the solenoid valve when one of the pressure regulating valves fails without affecting the signal pressure output characteristics of the solenoid valve at all.

On the other hand, according to the third aspect of the present invention, an extremely simple circuit configuration having a high reliability due to the absence of the operating portion enables
In the event of a pressure regulator valve failure, the supply of the base pressure to the solenoid valve via the orifice is assured as in the normal case. Therefore, according to this configuration, it is possible to maintain achievement of all shift speeds that can be achieved in the gear train.

Next, according to the configuration of the fourth aspect, even when any of the first and second pressure regulating valves fails, the vehicle can travel by achieving any of the forward traveling stages.

[0014]

Embodiments of the present invention will be described below with reference to the drawings. FIG. 1 shows six forward speeds according to the application of the present invention.
One example of the gear train of the reverse first speed automatic transmission is shown by a skeleton. As shown in the figure, this automatic transmission is for a front engine rear drive (FR) vehicle, and includes a torque converter 2 with a lock-up clutch, a planetary gear transmission 1, and a clutch C-1 for controlling the same. To C-3, brakes B-1, B-3 and one-way clutch F
-2.

The planetary gear transmission 1 includes a Ravigneaux-type planetary gear set G and a deceleration planetary gear G1 for inputting deceleration rotation to the planetary gear set G. The planetary gear set G includes a small diameter sun gear S3, a large diameter sun gear S2, a long pinion P2 meshing with each other and meshing with the large diameter sun gear S2, and a short pinion P3 meshing with the small diameter sun gear S3.
The carrier C3 supports the pair of pinions, and the ring gear R3 meshes with the long pinion P2. Further, the planetary gear G1 for deceleration is composed of a simple planetary gear consisting of a sun gear S1, a pinion P1 meshing with the sun gear S1, a carrier C1 supporting the sun gear S1, and a ring gear R1 meshing with the pinion P1.

The small-diameter sun gear S3 of the planetary gear set G is provided with a clutch C-1 (hereinafter referred to as a C1 clutch).
To the carrier C1 of the reduction planetary gear G1, and the large-diameter sun gear S2 is connected to the clutch C-3 (hereinafter C3).
The clutch B) is connected to the same carrier C1 of the reduction planetary gear G1 by the brake B-1.
(Hereinafter, referred to as B1 brake), the carrier C3 can be engaged with the clutch C-2 (hereinafter, referred to as C1 brake).
Two clutches) are connected to the input shaft 11 and can be locked to the case 10 by a brake B-2 (hereinafter referred to as B2 brake), and the ring gear R3 is connected to the output shaft 19. A one-way clutch F-2 is arranged in parallel with the B3 brake. The reduction planetary gear G1 has its sun gear S1 fixed to the transmission case 10, the ring gear R1 connected to the input shaft 11, the carrier C1 connected to the small-diameter sun gear S3 of the planetary gear set G via a C1 clutch, and C
It is connected to the large-diameter sun gear S2 of the planetary gear set G via three clutches.

The thus configured planetary gear transmission 1
Each of the above clutches and brakes is provided with a hydraulic servo composed of a multi-plate friction member and a piston / cylinder mechanism for engaging and releasing them, as is well known,
Under the control of an electronic control device (not shown) and a hydraulic control device described later, each hydraulic servo is controlled by a hydraulic control device attached to the transmission case 10 based on a vehicle load in a speed range corresponding to a range selected by the driver. The friction engagement member is engaged / disengaged by supplying / discharging the hydraulic pressure to the gears, thereby performing gear shifting.

FIG. 2 graphically illustrates the relationship between the engagement of each clutch, brake and one-way clutch of the gear train illustrated in FIG. 1 and the speeds achieved thereby. The circles in the figure indicate engagement for each clutch and brake, and lock for the one-way clutch.

First gear (1st) in this gear train
Is achieved by automatic engagement of the one-way clutch F-2 corresponding to engagement of the C1 clutch and the B2 brake. In this case, referring to FIG. 1, the rotation reduced from the input shaft 11 via the reduction planetary gear G1 is input to the small-diameter sun gear S3 via the C1 clutch, and the one-way clutch F-
2 takes a reaction force to the carrier C3 locked by the engagement of the gear 2 to reduce the rotation of the ring gear R3 at the maximum gear ratio to the output shaft 1.
9 is output.

Next, the second speed (2nd) is achieved by engagement of the C1 clutch and the B1 brake. In this case, the rotation reduced from the input shaft 11 via the deceleration planetary gear G1 is input to the small-diameter sun gear S3 via the C1 clutch, and takes a reaction force to the large-diameter sun gear S2 locked by engagement of the B1 brake to generate a ring gear. The decelerated rotation of R3 is output to the output shaft 19. The reduction ratio at this time is the first speed (1
st)).

The third speed (3rd) is achieved by simultaneously engaging the C1 clutch and the C3 clutch. In this case, the rotation reduced from the input shaft 11 via the deceleration planetary gear G1 is simultaneously input to the large-diameter sun gear S3 and the small-diameter sun gear S3 via the C1 clutch and the C3 clutch, and the planetary gear set G is directly connected. The rotation of the ring gear R3 at the same speed as the input rotation of the output shaft 1
9 is output.

Further, the fourth speed (4th) is achieved by simultaneous engagement of the C1 clutch and the C2 clutch. In this case, on the one hand, the rotation decelerated from the input shaft 11 via the deceleration planetary gear G1 is input to the sun gear S3 via the C1 clutch, and on the other hand, the non-reduced rotation input from the input shaft 11 via the C2 clutch is input to the carrier C3. Then, the rotation at the intermediate speed between the two input rotations is output to the output shaft 19 as the rotation of the ring gear R <b> 3 slightly reduced with respect to the rotation of the input shaft 11.

Next, the fifth speed (5th) is achieved by simultaneous engagement of the C2 clutch and the C3 clutch. In this case, the rotation decelerated from the input shaft 11 via the deceleration planetary gear G1 is input to the sun gear S2 via the C3 clutch, and the non-decelerated rotation input from the input shaft 11 via the C2 clutch is input to the carrier C3. Then, the rotation of the ring gear R <b> 3 at a speed slightly higher than the rotation of the input shaft 11 is output to the output shaft 19.

The sixth speed (6th) is achieved by engagement of the C2 clutch and the B1 brake. in this case,
Non-reduced rotation is input only to the carrier C3 from the input shaft 11 via the C2 clutch, and further increased rotation of the ring gear R3 which takes a reaction force to the sun gear S2 locked by engagement of the B1 brake is applied to the output shaft 19. Is output.

The reverse (R) means that the C3 clutch and the B3
Achieved by engagement of the brake. In this case, input shaft 1
The rotation reduced from 1 through the reduction planetary gear G1 is input to the sun gear S2 via the C3 clutch, and a large reverse rotation of the gear ratio of the ring gear R3 which takes a reaction force to the carrier C3 locked by engagement of the B3 brake is output. Output to axis 19.

Next, in the gear train shown in FIG.
The configuration of the hydraulic control device for achieving each shift speed shown in the operation chart of FIG. 2 will be described. FIG. 3 shows one half of the circuit configuration of the hydraulic control device, and FIG. 4 shows the other. This hydraulic circuit adjusts the hydraulic pressure sucked up by an oil pump 51 as a hydraulic pressure source and discharged to a line pressure oil passage L1 to a secondary pressure oil passage and a drain oil passage (both not shown) by a primary regulator valve 52. To generate an appropriate line pressure according to the running load of the vehicle, and control the pressure and direction by each valve in the circuit as a base pressure for control to control the hydraulic servo 81 of each friction engagement element.
To 84 are constituted.

Hereinafter, the relationship between each valve constituting the circuit and the connection of the oil passage will be described. First, the primary regulator valve 52 is constituted by a spring-loaded spool and a pressure regulating valve having a plunger abutting on a spring-loaded spool end. The primary regulator valve 52 has an input port connected to the line pressure oil passage L1,
An output port that communicates with the secondary pressure oil passage and a drain port that communicates with the oil pump on the suction side via a drain oil passage are provided. A direct feedback pressure of the line pressure is applied to the spool for controlling the degree of communication between these ports via the orifice in opposition to the spring force, and the throttle solenoid valve 53 outputs the throttle pressure in a direction superimposed on the spring force. Pressure is applied as a signal pressure. At the appropriate line pressure, the primary regulator valve 52 reduces the degree of communication to the drain port to supply the excess pressure mainly to the secondary pressure oil passage, and when the applied signal pressure increases, the degree of communication to the drain port is reduced. The drain amount is increased, and the line pressure in the line pressure oil passage L1 is maintained at a predetermined value.

On the other hand, the line pressure oil passage L1 supplies first and second solenoid modulator valves 54 and 55 for supplying a solenoid modulator pressure as a base pressure for generating solenoid signal pressure to each of the solenoid valves 71 to 74.
To each other via a strainer, while C3
The solenoid valve (SLC3) as a control valve for the clutch and the B1 brake is connected to input ports of the solenoid valve (SLB1).

Here, the first and second solenoid modulator valves 54 and 55 and each of the solenoid valves 71 to 7
4 will be described. In this hydraulic control device, the first system hydraulic servos 81 and 83 for operating the C1 clutch alone or the C1 clutch and the C3 clutch as friction engagement elements capable of achieving at least one forward traveling stage, and the supply oil passage L4 thereof , L8 and hydraulic servos 82, 84 of the second system for operating the C2 clutch and the B1 brake as friction engagement elements capable of achieving other forward travel stages and their supply oil passages L7, L9 are arranged in parallel. Spool valve portions 71A-71 of solenoid valves are provided as control valves for individually controlling the supply of hydraulic pressure to each hydraulic servo in each supply oil passage.
74A (to be described in detail later) is inserted, and dedicated solenoid valve portions 71B to 74B (to be described in detail later) for applying a signal pressure for their control operation are arranged in each spool valve portion. Between the line pressure oil passage L1 and the solenoid valve portion that applies signal pressure to the spool valve portion of the supply oil passage of the first system.
Solenoid modulator valve 5
4 is interposed between the line pressure oil passage L1 and a solenoid valve portion for applying a signal pressure to a spool valve portion of the supply oil passage of the second system, the second solenoid modulator serving as a second pressure regulating valve. A valve 55 is inserted.

As is well known, the manual valve 56 is a spool valve having seven positions that can be switched by operating a shift lever by a vehicle driver. That is, the operation of the spool closes the input port connected to the line pressure oil passage L1, the "P" position,
An "R" position for communicating with the range output port and draining the other output ports, an "N" position for closing the input port to all output ports (the figure shows this "N" position), and an input "D", which connects the port to the D range output port, drains the R range output port, and closes the second D range output port
The “4” and “3” positions and the input port communicate with both the D range output port and the second D range output port,
It has a "2" position to drain the R range output port. The D range output port of this valve is
Via the range oil passage L2, it is connected to the input ports of the solenoid valves SLC1, SLC2 of the C1 clutch and the C2 clutch. Further, the R range output port is connected to a reverse signal pressure port communicating with the plunger end side pressure receiving portion of the primary regulator valve 52 via the R range oil passage L3.

A supply oil passage L4 for the C1 clutch hydraulic servo 81 is connected to a D range oil passage L2, and a C1 solenoid valve (SLC1) 71 for adjusting the pressure based on a signal from an electronic control unit is provided on the supply oil passage L4. It is configured to be provided. The C1 solenoid valve 71 has a spool valve portion 71A as a three-port control valve for controlling the degree of communication between an input / output port and a drain port by a spring-loaded spool, and a solenoid pressure at the end of the spool opposite to the spring load. It is constituted by a combination of a linear solenoid valve portion 71B as a three-port type solenoid valve to which a solenoid load and a spring load to be applied are applied opposite to each other. The linear solenoid valve portion 71B has its input port connected to the output port of the first solenoid modulator valve 54 via the modulator pressure oil passage L5, and has its output port connected to the spool valve portion 71B.
A signal pressure port. The spool port 71A has an input port connected to the D range oil passage L2,
The output port is connected to the C1 clutch hydraulic servo 81,
A feedback port communicating with the spool end on the spring load side is connected to an oil passage downstream of the output port via an orifice.

A supply oil passage L7 for the C2 clutch hydraulic servo 82 is connected to a D range oil passage L2, and a C2 solenoid valve (SLC2) on the supply oil passage L7 is operated to perform pressure adjustment based on a signal from an electronic control unit. 72. The C2 solenoid valve 72 also has a three-port spool valve portion 72A as a control valve for controlling the degree of communication between the input / output port and the drain port by a spring-loaded spool, and a solenoid pressure at the end of the spool opposite to the spring load. It is constituted by a combination of a three-port linear solenoid valve portion 72B as a solenoid valve to which a solenoid load and a spring load to be applied are applied in opposition to each other. The input port of the linear solenoid valve portion 72B is connected to the output port of the second solenoid modulator valve 55 via the second modulator pressure oil passage L6, and the output port is connected to the signal pressure port of the spool valve portion 72A. It is connected. The spool valve portion 72A has an input port connected to the D range oil passage L2, an output port connected to the C2 clutch hydraulic servo 82, and a feedback port communicating with the spool end on the spring load side downstream of the output port via an orifice. Connected to oil passage.

A supply oil passage L8 for the C3 clutch hydraulic servo 83 is connected to a line pressure oil passage L1, and a C3 solenoid valve (SLC3) on the supply oil passage L8 is operated to regulate the pressure based on a signal from an electronic control unit. 73 is provided. C3 solenoid valve 73 in this case
Also, a three-port type spool valve portion 73A as a control valve for controlling the degree of communication between an input / output port and a drain port by a spring-loaded spool, and a solenoid for applying a solenoid pressure to the end of the spool opposite to the spring-loaded side. It is composed of a combination of a three-port linear solenoid valve portion 73B as a solenoid valve to which a load and a spring load are applied in opposition. The input port of the linear solenoid valve portion 73B is connected to the output port of the first solenoid modulator valve 54 via the first modulator pressure oil passage L5, and the output port is connected to the signal pressure port of the spool valve portion 73A. It is connected. The spool valve portion 73A has an input port connected to the line pressure oil passage L1, an output port connected to the C3 clutch hydraulic servo 83, and a feedback port communicating with the spool end on the spring load side downstream of the output port via an orifice. Connected to oil passage.

A supply oil passage L9 for the B1 brake hydraulic servo 84 is connected to a line pressure oil passage L1, and a B1 solenoid valve (SLB1) on the supply oil passage L9 is operated to regulate the pressure based on a signal from an electronic control unit. 74. B1 solenoid valve 74 in this case
Also, a three-port type spool valve portion 74A as a control valve for controlling the degree of communication between an input / output port and a drain port by a spring-loaded spool, and a solenoid for applying a solenoid pressure to the end of the spool opposite to the spring-loaded side. It is constituted by a combination of a three-port type linear solenoid valve portion 74B as a solenoid valve to which a load and a spring load are applied oppositely. The input port of the linear solenoid valve portion 74B is connected to the output port of the second solenoid modulator valve 54 via the second modulator pressure oil passage L6, and the output port is connected to the signal pressure port of the spool valve portion. Have been. The spool valve 74A is connected to the input port via the line pressure oil passage L.
1 and the output port is B1 brake hydraulic servo 8
4 and a feedback port communicating with the spool end on the spring load side is connected to an oil passage downstream of the output port via an orifice.

It should be noted that the B2 brake hydraulic servo and the supply oil passage for the B2 brake are separately supplied hydraulic pressure without passing through the solenoid modulator valve according to the present invention, so that the illustration is omitted.

In the hydraulic circuit having the above configuration, at the "N" position of the manual valve 56, the input port connected to the line pressure oil passage L1 is closed by a land, and all output ports are drained. Double solenoid modulator valve 5 connected to pressure oil passage L1
The modulator pressure adjusted at 4,55 is the first and second modulator pressure.
Are supplied to the modulator pressure oil passages L5 and L6, however, since the solenoid valves 71 to 74 are turned on and closed, supply of the apply pressure from these solenoid valves 71 to 74 is not performed. This communication relationship is the same for the "P" position of the manual valve 56, although the spool position is different. In these positions, hydraulic pressure is supplied to the input ports of the spool valves 73A and 74A of the solenoid valves 73 and 74 connected to the line pressure oil passage L1.

When the manual valve 56 is switched to the "D" position, the line pressure is also output to the D range oil passage L2, so that the input ports of the spool valve portions 71A to 74A of all the solenoid valves 71 to 74 are connected. Is supplied with line pressure. That is, the oil pressure of the D range oil passage L2 is supplied to the input ports of the C1 solenoid valve 71 and the C2 solenoid valve 72, and the oil pressure of the line pressure oil passage L1 is supplied to the input ports of the C3 solenoid valve 73 and the B1 solenoid valve 74. Become like

Next, the normal valve operation will be described. When the signal to the C1 solenoid valve 71 is turned off to achieve the first speed, the line pressure of the D range oil passage L2 supplied to the C1 solenoid valve 71 is regulated by the valve 71 to be the apply pressure. , C1 clutch hydraulic servo 81. As a result, the C1 clutch is engaged, and the first clutch is cooperated with the one-way clutch F-2.
Speed is achieved.

The second speed is achieved by turning off the signal to the C1 solenoid valve 71 and turning off the signal to the B1 solenoid valve 74. In this state, in addition to the supply state of the apply pressure to the C1 clutch hydraulic servo 81, the B1 solenoid valve 74 enters a pressure adjustment state, and the adjusted apply pressure is supplied to the B1 brake hydraulic servo 84. In this way, the second speed is achieved by engaging the C1 clutch and supporting the B1 brake reaction force.

The third speed is achieved by turning off the signal to the C1 solenoid valve 71 and turning off the signal to the C3 solenoid valve 73. In this case, the supply pressure supply state to the C1 clutch hydraulic servo 81 is kept as it is, the C3 solenoid valve 73 enters a pressure adjustment state, and the apply pressure is supplied to the C3 clutch hydraulic servo 83. Thus, the third speed is achieved by the simultaneous engagement of the C1 clutch and the C3 clutch.

The fourth speed is achieved by turning off the signal to the C1 solenoid valve 71 and turning off the signal to the C2 solenoid valve 72. In this state, the supply pressure supply state to the C1 clutch hydraulic servo 81 remains unchanged,
2 The solenoid valve 72 is in the applied pressure regulating state,
The apply pressure is supplied to the C2 clutch hydraulic servo 82. Thus, the fourth speed is achieved by the simultaneous engagement of the C1 clutch and the C2 clutch.

The fifth speed is achieved by turning off the signal to the C2 solenoid valve 72 and turning off the signal to the C3 solenoid valve 73. In this state, the apply pressure to the C2 clutch hydraulic servo 82 is similarly applied to the same valve as in the fourth speed, and the apply pressure to the C3 clutch hydraulic servo 83 is the same as in the third speed. Is similarly applied. Thus, the fifth speed is achieved by the simultaneous engagement of the C2 clutch and the C3 clutch.

The sixth speed is achieved by turning off the signal to the C2 solenoid valve 72 and turning off the signal to the B1 solenoid valve 74. Even in this state, the apply pressure to the B1 brake hydraulic servo 84 is similarly applied to the same valve as in the case of the second speed. Thus, C2 clutch engagement,
The sixth speed is achieved by the B1 brake reaction force support.

Next, the operation of the hydraulic control device which normally enters the above-described hydraulic pressure supply state when the solenoid modulator valve fails will be described. First, when the first solenoid modulator valve 54 sticks in the closed state, the supply of the hydraulic pressure to the first modulator pressure oil passage L5 stops, so that even if the electric signals of the C1 solenoid valve 71 and the C3 solenoid valve 73 are turned off, No signal pressure is output. As a result, the spools of the spool valve portions 71A and 73A of both solenoid valves 71 and 73 both close between the input and output ports due to the spring load. As a result, hydraulic pressure supply to the C1 clutch hydraulic servo 81 and the C3 clutch hydraulic servo 83 becomes impossible. In this state, as can be seen with reference to the engagement chart of FIG.
It is impossible to achieve the first to fourth speeds that involve the engagement of the clutch, and it is also impossible to achieve the fifth speed that involves the engagement of the C3 clutch. However, at this time, the achievable state of the sixth speed by engagement of the C2 clutch and the B1 brake is maintained. In this case, the supply of the apply pressure to the hydraulic servos 82 and 84 of the C2 clutch and the B1 brake is performed by the supply of the solenoid modulator pressure by the second solenoid modulator valve 55 to the C2 solenoid valve 7.
2 and the pressure adjusting operation of the B1 solenoid valve 74.

On the other hand, if the second solenoid modulator valve 55 sticks in the closed state, the supply of the hydraulic pressure to the second modulator pressure oil passage L6 stops, so that the electric signals of the C2 solenoid valve 72 and the B1 solenoid valve 74 are turned off. Also, no signal pressure is output from them. As a result, the spools of the spool valve portions 72A, 74A of both solenoid valves 72, 74 both close between the input and output ports due to the spring load. As a result, C2
Clutch hydraulic servo 82 and B1 brake hydraulic servo 84
Supply of hydraulic pressure to the motor becomes impossible. In this state, as can be seen with reference to the engagement chart of FIG. 2, it is impossible to achieve the fourth to sixth speeds involving the engagement of the C2 clutch, and the second gear involving the engagement of the B1 brake. Achieving speed is also impossible.
However, at this time, the single engagement (B
The engagement of the first brake by the engagement of the three brakes is possible regardless of the failure of the solenoid modulator valve) or the third gear by the engagement of the C1 clutch and the C3 clutch is maintained. In this case, supply of the apply pressure to the hydraulic servos 81 and 83 of the C1 clutch and the C3 clutch is performed by the first
C1 solenoid valve 71 by supply of solenoid modulator pressure by solenoid modulator valve 54 of FIG.
And the pressure adjusting operation of the C3 solenoid valve 73.

Thus, according to the circuit configuration of the first embodiment, the second solenoid modulator valve 55 of the pressure regulating circuit in parallel with the first solenoid modulator valve 54 connects the other solenoid valves 72 and 74 with each other. The base pressure is supplied to the first solenoid modulator valve 54.
Guaranteed on failure. Conversely, even when the second solenoid modulator valve 55 fails, the supply of the base pressure to the other solenoid valves 71 and 73 by the first solenoid modulator valve 54 is guaranteed. Therefore, by maintaining the signal pressure output of the solenoid valve related to achieving one of the forward gears by supplying the base pressure from the two pressure adjusting circuits in parallel, it is possible to operate the hydraulic servo by operating each spool valve unit. The running of the vehicle is ensured by the engagement of the corresponding friction engagement element. In particular, according to this circuit configuration, the solenoid valves 71 to 74
This has the advantage that the supply of the base pressure to the solenoid valve at the time of one solenoid modulator valve failure can be guaranteed without any influence on the signal pressure output characteristic of the solenoid valve.

Variations of the connection relationship between the first and second solenoid modulator valves 54 and 55 and the respective solenoid valves 71 to 74 in this embodiment are shown in FIG. The horizontal headings in this table represent the abbreviations of the hydraulic servos corresponding to the respective solenoid valves, and the circles in the respective columns indicate connection with the first solenoid modulator valve 54,
The mark represents the connection with the second solenoid modulator valve 55. The vertical heading pattern 1 corresponds to the connection of the present embodiment, and the connection relation in this case and the guarantee at the time of failure due to this are as described above.

The next pattern 2 is an example in which the first solenoid modulator valve 54 is connected to only the C1 solenoid valve 71, and all other solenoid valves are connected to the second solenoid modulator valve 55. Even in such a connection, when the first solenoid modulator valve 54 fails, referring to the engagement chart of FIG.
The achievable states of the fifth and sixth speeds, in which the engagement of the C1 clutch is not involved, are maintained. Conversely, when the second solenoid modulator valve 55 fails, the state of the first speed, in which the engagement of the other clutches and the brake is not involved, is maintained. It can be seen that the achievable state is maintained.

In the next pattern 3, the first solenoid modulator valve 54 is connected to the C1 solenoid valve 71 and B1
In this example, the solenoid valve is connected to a solenoid valve 74, and another solenoid valve is connected to a second solenoid modulator valve 55. In the case of this connection, similarly, referring to the engagement chart of FIG. 2, when the first solenoid modulator valve 54 fails, the state in which the fifth speed can be achieved by engagement of the C2 clutch and the C3 clutch is maintained, and When the second solenoid modulator valve 55 fails, the achievable state of the second speed by engagement of the C1 clutch and the B1 brake is maintained.

FIG. 6 shows a second embodiment of the present invention. As shown in FIG. 6, in this embodiment,
A pressure regulating circuit that connects the line pressure oil passage L1 and each of the linear solenoid valve portions 71B to 74B as a solenoid valve in parallel with the solenoid modulator valve 54A as a pressure regulating valve bypasses the solenoid modulator valve 54A to reduce the line pressure. The orifice is provided with a reduced pressure. That is, in this circuit, the same diameter as the orifice 54a of the oil passage which feeds back the solenoid modulator pressure to the spool end of the solenoid modulator valve 54A similar to the first and second solenoid modulator valves 54 and 55 of the first embodiment. The orifice 6 has a circuit configuration inserted in the middle of an oil passage that short-circuits the input side and the output side of the solenoid modulator valve 54A.

In this embodiment, the hydraulic pressure is supplied to the modulator pressure oil passage regardless of the failure of the solenoid modulator valve 54A.
All the solenoid valves 71 to 74 as one system oil passage
Are connected to the input ports of the solenoid valve portions 71B to 74B.

In the case of adopting such a configuration, it is denied that the normal supply of hydraulic pressure from the parallel circuit via the orifice 6 may affect the solenoid signal pressure output characteristics of the solenoid valve portions 71B to 74B during normal times. Although there is no very simple movable circuit, the hydraulic servo is controlled by the signal output operation of each of the linear solenoid valves 71 to 74 at the time of stick failure of the solenoid modulator valve 54A and the operation of the spool valve units 71A to 74A. Supply of the application pressure to 81 to 84 can be guaranteed. Therefore, according to this configuration,
The advantage that the gear train can ensure the vehicle traveling at all the gear stages that can be achieved originally can be obtained.

Although the present invention has been described in detail with reference to the two embodiments, the concept of the present invention is not limited to the illustrated hydraulic circuit, but is applicable to a general hydraulic control circuit. .

[Brief description of the drawings]

FIG. 1 is a skeleton diagram showing a gear train of a six-speed automatic transmission controlled by a hydraulic control device according to a first embodiment of the present invention.

FIG. 2 is a chart showing an operation of a gear train by the hydraulic control device of the first embodiment.

FIG. 3 is a circuit diagram showing one half of a circuit of the hydraulic control device according to the first embodiment;

FIG. 4 is a circuit diagram showing the other half of the circuit of the hydraulic control device according to the first embodiment;

FIG. 5 is a table showing a connection pattern between a pressure regulating valve and a control valve of the hydraulic control device according to the first embodiment.

FIG. 6 is a partial circuit diagram of a hydraulic control device according to a second embodiment.

[Explanation of symbols]

C-1 C1 clutch (friction engagement element) C-2 C2 clutch (friction engagement element) C-3 C3 clutch (friction engagement element) B-1 B1 brake (friction engagement element) 54 First solenoid modulator Valve (first pressure regulating valve) 55 Second solenoid modulator valve (second pressure regulating valve) 6 Orifice 71 C1 solenoid valve 71A Spool valve (control valve) 71B Solenoid valve (solenoid valve) 72 C2 solenoid valve 72A Spool Valve section (control valve) 72B solenoid valve section (solenoid valve) 73 C3 solenoid valve (solenoid valve and control valve) 73A spool valve section (control valve) 73B solenoid valve section (solenoid valve) 74 B1 solenoid valve (solenoid valve and control) Valve) 74A spool valve Control valve) 74B solenoid valve section (solenoid valve) 81 C1 hydraulic servo 82 C2 hydraulic servo 83 C3 hydraulic servo 84 B1 hydraulic servo

Continued on the front page (72) Inventor Masahiro Hayabuchi 10 Takane, Fujii-machi, Anjo, Aichi Prefecture Inside Aisin AW Co., Ltd. (72) Inventor Masaaki Nishida 10 Takane, Fujii-cho, Anjo, Aichi Aisin Ai F-term (reference) in W3M Co., Ltd. 3J552 MA02 MA12 MA26 NA01 NB01 PB06 QA13A QA14A QA36A QA42A QB02 SA52

Claims (4)

[Claims] A control valve for controlling a hydraulic pressure to be supplied to the hydraulic servo, the control valve being interposed in a supply oil passage of a hydraulic servo for operating a friction engagement element; and applying a signal pressure for control operation to the control valve. Hydraulic control of an automatic transmission including a solenoid valve and a pressure regulating valve interposed between the line pressure oil passage and the solenoid valve to regulate the line pressure and supply a base pressure for signal pressure output to the solenoid valve A hydraulic control device for an automatic transmission, comprising: a pressure adjusting circuit that connects a line pressure oil passage and a solenoid valve in parallel with the pressure adjusting valve. 2. The hydraulic control of an automatic transmission according to claim 1, wherein the pressure control circuit includes a second pressure control valve that controls a line pressure and supplies a base pressure for outputting a signal pressure to a solenoid valve. apparatus. 3. The hydraulic control device for an automatic transmission according to claim 1, wherein said pressure regulating circuit includes an orifice for reducing a line pressure. 4. A hydraulic servo of a first system for operating a friction engagement element capable of achieving at least one forward travel stage and its supply oil passage, and a friction engagement element capable of achieving another forward travel stage. A hydraulic servo of the second system to be operated and its supply oil passage are provided in parallel, and a control valve for individually controlling the supply of hydraulic pressure to each hydraulic servo is inserted in each of the supply oil passages. A hydraulic control device for an automatic transmission, in which a dedicated solenoid valve for applying a signal pressure for the control operation to the control valve is disposed, wherein a line pressure oil passage and a supply oil passage of the first system are provided. A first pressure regulating valve is interposed between a solenoid valve for applying a signal pressure to the control valve, and a solenoid valve for applying a signal pressure to a control valve of a line pressure oil passage and a control valve of a supply oil passage of the second system. A second pressure regulating valve is interposed between Hydraulic control system for an automatic transmission that.