JPH10196777A - Hydraulic control device for automatic transmission - Google Patents

Abstract

(57) Abstract: An object of the present invention is to provide a hydraulic control device for an automatic transmission capable of reliably and accurately controlling a hydraulic pressure for engaging and disengaging a friction engagement device of the automatic transmission. SOLUTION: In a hydraulic control device of an automatic transmission for controlling a hydraulic pressure of a friction engagement device engaged at the time of shifting of an automatic transmission 20 connected to an engine 200, an output control state for detecting a state of the output control device. Detecting means 202; basic oil pressure calculating means 203 for calculating a basic oil pressure based on the state of the output control device; engine output estimating means 204 for estimating the engine output based on the engine load; and based on the estimated engine output. A correction value calculating means 205 for calculating a correction value for correcting the basic oil pressure, and a hydraulic pressure calculating means 206 for calculating the oil pressure of the friction engagement device based on at least the basic oil pressure among the basic oil pressure or the correction value are provided. .

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 controlling a hydraulic pressure for engaging and releasing a friction engagement device of an automatic transmission.

[0002]

2. Description of the Related Art An automatic transmission for a vehicle is configured such that a frictional engagement device such as a clutch or a brake is engaged or released by hydraulic pressure to set a gear ratio according to a running state of the vehicle. I have. The torque capacity or engagement pressure of the friction engagement device is set to be slightly higher than the minimum capacity or pressure that can withstand the torque applied to the friction engagement device, for the purpose of reducing the size and weight and improving fuel efficiency. . Usually, this is a torque capacity that maximizes a line pressure set based on an engine load such as a throttle opening.

On the other hand, a frictional engagement device is used as a torque transmitting means in order to absorb inertial energy due to slippage during a shift transition and to eliminate a shock accompanying rotation fluctuation. Therefore, when engaging / disengaging the friction engagement device, it is necessary to smoothly change the engagement pressure. However, if the friction engagement device slips excessively,
It takes a long time to shift gears, resulting in inferior shifting responsiveness, and intensified wear of the friction engagement device, resulting in poor durability.

[0004] Therefore, conventionally, by using an accumulator, the change in the engagement pressure of the frictional engagement device is made smooth, and by continuously changing the back pressure, the characteristics of the accumulator are variously changed, and the speed change is performed. Further improvements are being made to shock. The back pressure of the accumulator, in other words, the accumulator control oil pressure is controlled by a linear solenoid valve.

Conventionally, as means for controlling the accumulator control oil pressure, one of the following two oil pressure control means has been applied. The first hydraulic pressure control means controls the accumulator control hydraulic pressure based on the operation state of the output control device of the engine. Specifically, an accelerator pedal opening or a throttle opening as an output control device operated by the driver's intention is detected by a sensor, and the duty ratio of the linear solenoid valve is controlled based on the detected data, It controls the accumulator control hydraulic pressure.

The second oil pressure control means estimates the engine output by detecting the engine load, for example, the engine speed, the throttle opening, the intake air amount, the fuel injection amount, etc., using various sensors, and based on the estimated value. Thus, the duty ratio of the linear solenoid valve is controlled to control the accumulator control oil pressure.

[0007]

However, each of the two conventional hydraulic control means has the following problems. That is, according to the first control means,
There is a time delay before changes in the accelerator opening and throttle opening actually appear as engine output. For this reason, when the accumulator control oil pressure is controlled based on the accelerator opening and the throttle opening, the torque to be transmitted does not match the engagement pressure of the friction engagement device, which causes a shift shock and a response delay. There was a possibility.

According to the second control means, since the engine output is estimated based on the engine load and the hydraulic pressure of the friction engagement device is controlled based on the engine output, the accumulator control hydraulic pressure is controlled with high precision. Is possible. However, when the engine output cannot be estimated due to a failure of various sensors, or when an error occurs in data of various sensors, there is a possibility that the control accuracy of the accumulator control hydraulic pressure is reduced or the control becomes impossible. Was.

Further, there is a case where control for automatically lowering the engine output is performed in response to a request from the automatic transmission. According to the second control means, the engine output is estimated even in the case of such a torque down control. Therefore, another hydraulic control is performed based on the estimation result. There was a problem of worsening.

The present invention has been made in view of the above circumstances, and it is possible to reliably and precisely control a hydraulic pressure for engaging and releasing a friction engagement device of an automatic transmission. It is an object of the present invention to provide a hydraulic control device for an automatic transmission.

[0011]

SUMMARY OF THE INVENTION In order to achieve the above object, the present invention, as shown in FIG. 1, employs a frictional engagement which is engaged during shifting of an automatic transmission 201 connected to an engine 200. An output control state detecting means for detecting a state of an output control device operated to control an output of the engine; and a state of the output control device. Basic oil pressure calculating means 203 for calculating the basic oil pressure based on
An engine output estimating means 204 for estimating an engine output based on the load of the engine, a correction value calculating means 205 for calculating a correction value for correcting the basic hydraulic pressure based on the estimated engine output, And a hydraulic pressure calculating means 206 for calculating a hydraulic pressure of the friction engagement device based on at least a basic hydraulic pressure of the basic hydraulic pressure or the correction value.

The output control device includes an accelerator opening and a throttle opening. The engine load includes an engine speed, an intake air amount, a throttle opening, a fuel injection amount, and the like.

According to the present invention, since the engine output is detected based on the state of the output control device operated by the driver, the output control state of the engine can be reliably detected. Further, the basic hydraulic pressure is calculated based on the detected output control state of the engine.

On the other hand, an engine output is estimated based on the engine load, and a correction value is calculated based on the estimated engine output. Then, since the hydraulic pressure is calculated by correcting the basic hydraulic pressure with the correction value, the hydraulic pressure of the friction engagement device can be controlled reliably and with high accuracy, and the shift shock suppression function and operation responsiveness are improved. Is improved.

Further, when a failure occurs in the engine output estimating means and it becomes impossible to estimate the engine output, the oil pressure of the friction engagement device is calculated based on the basic oil pressure. Therefore, even when the engine output estimating means fails, the engagement pressure of the friction engagement device can be controlled to an appropriate state.

[0016]

Next, the present invention will be described more specifically with reference to the drawings. FIG. 2 is a block diagram showing the overall control system of the engine (drive power source) E and the automatic transmission A which are the objects of the present invention. In FIG. 2, an engine E to which an automatic transmission A is connected is configured to electrically control its output.
An electronic throttle valve 13 is provided in the intake pipe 12.

On the other hand, the depression amount of the accelerator pedal 15 for controlling the output of the engine E, that is, the accelerator opening is detected by a sensor (not shown), and the detection signal is input to an engine electronic control unit (E-ECU) 17. Have been. The engine electronic control unit 17 includes a central processing unit (CPU) and a storage device (RAM, RO
M) and an input / output interface. The engine electronic control device 17 includes engine (E / G) rotation speed Ne,
Intake air amount Q, intake air temperature, throttle opening, vehicle speed,
Various signals such as engine water temperature, signals from brake switches and idle switches are input. The opening of the electronic throttle valve 13 is controlled based on these data, and the fuel injection amount and the ignition timing of the engine E are controlled.

In the automatic transmission A, the hydraulic pressure control device 18 controls the speed change and the lock-up clutch, the line pressure, or the engagement pressure of a predetermined friction engagement device. The hydraulic control device 18 is configured to be electrically controlled, and has first to third shift solenoid valves S1 to S3 for executing a shift, mainly for controlling an engine braking state. The fourth solenoid valve S4,
A linear solenoid valve SLT for mainly controlling the line pressure, a linear solenoid valve SLN for mainly controlling the accumulator back pressure, and a linear solenoid valve SLU mainly for controlling the engagement pressure of the lock-up clutch are provided.

An electronic control unit (T-ECU) 19 for an automatic transmission for outputting a signal to these solenoid valves to control a shift, a line pressure or an accumulator back pressure.
Is provided. This electronic control unit for automatic transmission 19
Is a central processing unit (CPU) and a storage device (RA
M, ROM) and an input / output interface.

The electronic control unit 19 for an automatic transmission includes:
Accelerator opening, throttle opening, vehicle speed, engine water temperature, signal from brake switch, shift position, signal from pattern select switch, signal from overdrive switch, signal from overdrive switch, idle switch, clutch C0 A signal from an NC0 sensor for detecting a rotation speed, an oil temperature of an automatic transmission, a signal from a manual shift switch, and the like are input.

The electronic control unit 19 for the automatic transmission and the electronic control unit 17 for the engine are connected to each other so as to be able to perform data communication, and the electronic control unit 17 for the engine is connected to the electronic control unit 19 for the automatic transmission. Signals such as the amount of intake air per revolution (Q / Ne) are transmitted to the electronic control unit 19 for the automatic transmission and the electronic control unit 17 for the engine. And a signal instructing a gear position are transmitted.

That is, in the electronic control unit 19 for the automatic transmission, based on the input vehicle speed V, the throttle opening θ and the previously stored shift map, the shift determination, the ON / OFF of the lock-up clutch, the line pressure and the friction are performed. Judgment of the pressure adjustment level of the engagement pressure of the engagement device, etc., outputs an instruction signal to a predetermined solenoid valve based on the judgment result, and further performs judgment of failure and control based on the judgment result of failure. ing. In the electronic control unit 19 for the automatic transmission, the shift output and the start of the shift of the automatic transmission A, in other words, the start of the inertia phase are determined by the input rotation speed N.
The calculation is performed by a known method based on C0, the output rotation speed N0, and the gear ratio.

The engine electronic control unit 17 controls the fuel injection amount, the ignition timing, the opening of the electronic throttle valve 13 and the like based on the input data. Reduce injection volume,
Alternatively, the output torque is temporarily reduced by changing the ignition timing or reducing the opening of the electronic throttle valve 13.

FIG. 3 is a diagram showing an example of the gear train of the automatic transmission A. In the configuration shown here, five forward speeds and one reverse speed are set. That is, the automatic transmission A shown here is
, A sub transmission unit 21 and a main transmission unit 22.

The torque converter 20 has a lock-up clutch 23. The lock-up clutch 23 is a member (hub) in which a front cover 25 integrated with a pump impeller 24 and a turbine runner 26 are integrally mounted. ) 27. Engine E
(Not shown) of the front cover 25
And an input shaft 28 connected to the turbine runner 26 is connected to a carrier 30 of an overdrive planetary gear mechanism 29 constituting the subtransmission portion 21.

The carrier 3 in this planetary gear mechanism 29
A multi-plate clutch C0 and a one-way clutch F0 are provided between the first gear 0 and the sun gear 31. The one-way clutch F0 is engaged when the sun gear 31 rotates forward relative to the carrier 30 (rotation in the rotation direction of the input shaft 28). A multi-disc brake B0 for selectively stopping the rotation of the sun gear 31 is provided.
The ring gear 3 which is an output element of the subtransmission portion 21
2 is connected to an intermediate shaft 33 which is an input element of the main transmission unit 22. Further, an NC0 sensor 34 for detecting the rotation speed of the multi-plate clutch C0, that is, the input rotation speed, is provided.

Therefore, in the sub-transmission portion 21, the entire planetary gear mechanism 29 rotates integrally with the multi-plate clutch C0 or the one-way clutch F0 in the engaged state, so that the intermediate shaft 33 has the same speed as the input shaft 28. At low speed. When the brake B0 is engaged and the rotation of the sun gear 31 is stopped, the speed of the ring gear 32 is increased with respect to the input shaft 28 and the ring gear 32 is rotated forward, so that a high gear is established.

On the other hand, the main transmission section 22 includes three sets of planetary gear mechanisms 40, 50, and 60, and their rotating elements are connected as follows. That is, the sun gear 41 of the first planetary gear mechanism 40 and the sun gear 51 of the second planetary gear mechanism 50 are integrally connected to each other, and the ring gear 43 of the first planetary gear mechanism 40 and the carrier 52 of the second planetary gear mechanism 50 are connected to each other. The three members of the third planetary gear mechanism 60 and the carrier 62 are connected, and the output shaft 65 is connected to the carrier 62. The rotation speed of the output shaft 65 is detected by the vehicle speed sensor 67. Further, the second planetary gear mechanism 5
The zero ring gear 53 is connected to the sun gear 61 of the third planetary gear mechanism 60.

In the gear train of the main transmission section 22, a reverse gear and four forward gears can be set, and a clutch and a brake for this are provided as follows. First, the clutch will be described. The ring gear 53 of the second planetary gear mechanism 50 and the third gear
A first clutch C1 is provided between the sun gear 61 of the planetary gear mechanism 60 and the intermediate shaft 33, and the sun gear 41 of the first planetary gear mechanism 40 and the sun gear 51 of the second planetary gear mechanism 50 and the intermediate shaft are connected to each other. A second clutch C2 is provided between the clutch 33 and the second clutch C2.

Next, the brake will be described. The first brake B1 is a band brake, and the sun gears 41 and 5 of the first planetary gear mechanism 40 and the second planetary gear mechanism 50.
1 is arranged to stop rotation. A first one-way clutch F1 and a second brake B2, which is a multi-plate brake, are arranged in series between the sun gears 41 and 51 (that is, the common sun gear shaft) and the casing 66. One-way clutch F1 has sun gears 41 and 5
1 is engaged when it is about to rotate in the reverse direction (rotation in the direction opposite to the rotation direction of the input shaft 28).

The third brake B 3, which is a multi-plate brake, is provided between the carrier 42 of the first planetary gear mechanism 40 and the casing 66. As a brake for stopping the rotation of the ring gear 63 of the third planetary gear mechanism 60, a fourth brake B4, which is a multi-plate brake, and a second one-way clutch F2.
Are arranged in parallel with the casing 66. The second one-way clutch F2 is adapted to be engaged when the ring gear 63 is about to rotate in the reverse direction.

In the above-mentioned automatic transmission A, five forward speeds and one reverse speed can be set by engaging and disengaging the clutches and brakes as shown in the operation chart of FIG. In FIG. 4, the mark ○ indicates the engaged state, the mark 時 に indicates the engaged state during engine braking, the mark △ indicates either engaged or released, and the blank indicates the released state.

As shown in the operation chart of FIG. 4, the shift between the second speed and the third speed is performed by the second brake B2 and the third speed.
The clutch-to-clutch shift changes both the engagement and release states with the brake B3. In order to smoothly perform this shift, the hydraulic control device 18 described above incorporates a hydraulic circuit shown in FIG.

In FIG. 5, reference numeral 70 denotes a 1-2 shift valve, reference numeral 71 denotes a 2-3 shift valve, and reference numeral 72 denotes a 3-4 shift valve. The communication state of each port of these shift valves 70, 71, 72 at each shift speed is determined by the respective shift valves 70, 71, 7
2 as shown below. The numbers indicate the respective gears. A third brake B3 is connected via an oil passage 75 to a brake port 74 communicating with the input port 73 at the first speed and the second speed among the ports of the 2-3 shift valve 71.

An orifice 76 is interposed in the oil passage, and a damper valve 77 is connected between the orifice 76 and the third brake B3. This damper valve 77 absorbs a small amount of hydraulic pressure when the line pressure is rapidly supplied to the third brake B3 to perform a buffering action.

Reference numeral 78 denotes a B-3 control valve which directly controls the engagement pressure of the third brake B3 by the B-3 control valve 78. That is, the B-3 control valve 78 includes a spool 79, a plunger 80, and a spring 81 interposed therebetween, and an oil passage 75 is connected to an input port 82 opened and closed by the spool 79. An output port 83 selectively connected to the input port 82 is connected to the third brake B3. Further, the output port 83 is connected to a feedback port 84 formed on the distal end side of the spool 79.

On the other hand, among the ports 85 of the 2-3 shift valve 71, a port 86 for outputting the D range pressure at the third or higher speed is connected to an oil passage 87. Are communicated through. The control port 88 formed on the end of the plunger 80 includes:
The lock-up clutch linear solenoid valve SLU is connected.

Therefore, the B-3 control valve 78
The pressure adjustment level is set by the elastic force of the spring 81 and the hydraulic pressure supplied to the port 85, and the elastic force of the spring 81 increases as the signal pressure supplied to the control port 88 increases. I have.

In FIG. 5, reference numeral 89 denotes a 2-3 timing valve. The 2-3 timing valve 89 is a spool 9 having a small land and two large lands.
0 and a first plunger 91, and a first plunger 9 with a spring 92 and a spool 90 interposed therebetween.
1 and a second plunger 93 arranged on the opposite side. An oil passage 95 is connected to a port 94 at an intermediate portion of the 2-3 timing valve 89, and the oil passage 95
Of the ports of the 2-3 shift valve 71, it is connected to a port 96 which is communicated with the brake port 74 at the third or higher speed.

Further, the oil passage 95 is branched on the way, and a port 97 is opened between the small land and the large land.
Connected through an orifice. The port 98 selectively communicated with the port 94 at the intermediate portion is the oil passage 9
9 is connected to the solenoid relay valve 100. A lock-up clutch linear solenoid valve SLU is connected to a port opened to the end of the first plunger 91, and a second brake B2 is connected to the port opened to the end of the second plunger 93 via an orifice. Connected.

The oil passage 87 is for supplying and discharging hydraulic pressure to and from the second brake B2, and a small-diameter orifice 101 and an orifice 102 with a check ball are interposed midway. The oil passage 103 branched from the oil passage 87 has a large-diameter orifice 10 having a check ball which opens when the pressure is released from the second brake B2.
The oil passage 103 is connected to an orifice control valve 105 described below.

The orifice control valve 105 is a valve for controlling the exhaust pressure speed from the second brake B2. The orifice control valve 105 has a port 107 formed at an intermediate portion so as to be opened and closed by a spool 106.
The oil passage 103 is connected to a port 108 formed below the port 107 in the figure. A port 109 formed above the port 107 to which the second brake B2 is connected in the figure is a port selectively communicated with a drain port. The port 111 of the B-3 control valve 78 is connected. The port 111 is a port selectively connected to the output port 83 to which the third brake B3 is connected.

A control port 112 formed at the end of the port of the orifice control valve 105 opposite to the spring that presses the spool 106 is connected to a port 114 of the 3-4 shift valve 72 via an oil passage 113. ing. The port 114 outputs the signal pressure of the third solenoid valve S3 at the third speed or lower and the fourth speed.
This port outputs the signal pressure of the fourth solenoid valve S4 at a speed higher than the speed. Further, an oil passage 115 branched from the oil passage 95 is connected to the orifice control valve 105, and the oil passage 115 is selectively connected to a drain port.

The port 11 for outputting the D range pressure at the second or lower speed in the 2-3 shift valve 71.
6 is connected to an oil passage 11 through a port 117 which is opened at a place where the spring 92 is disposed in the 2-3 timing valve 89.
8 are connected. 3-4 shift valve 72
The port 119 communicated with the oil passage 87 at the third or lower speed is connected to the solenoid relay valve 100 via the oil passage 120.

In FIG. 5, reference numeral 121 denotes an accumulator for the second brake B2, and an accumulator control pressure PACC regulated according to the oil pressure output from the linear solenoid valve SLN is supplied to the back pressure chamber. ing. The accumulator control pressure PACC is configured to increase as the output pressure of the linear solenoid valve SLN decreases. Therefore, the transient hydraulic pressure for engagement / disengagement of the second brake B2 changes at a higher pressure as the signal pressure of the linear solenoid valve SLN is lower.

Reference numeral 122 denotes a C-0 exhaust valve, and reference numeral 123 denotes an accumulator for the clutch C0. C-0 exhaust valve 1
Reference numeral 22 designates an operation for engaging the clutch C0 to apply the engine brake only in the second speed in the second speed range.

Therefore, according to the hydraulic circuit described above,
If the port 111 of the B-3 control valve 78 communicates with the drain, the engagement pressure of the third brake B3 can be directly regulated by the B-3 control valve 78, and the pressure regulation level is controlled by a linear solenoid valve. SLU
Can be changed by If the spool 106 of the orifice control valve 105 is located at the position shown in the left half of the figure, the second brake B2 is connected to the oil passage 103 through the orifice control valve 105. It is possible to relieve pressure and thus control the drain speed from the second brake B2.

Here, the correspondence between the configuration of this embodiment and the configuration of claim 1 will be described. The accelerator pedal 15 and the electronic throttle valve 13 correspond to the output control device of claim 1, and each brake and clutch Corresponds to the friction engagement device of the first aspect.

The correspondence between the hydraulic control device for an automatic transmission having the above-described hardware configuration and the functional means described in claim 1 will be described with reference to FIG. In FIG. 1, reference numerals different from those in FIGS. 2 and 3 are used for convenience.

That is, an output control state detecting means 201 for detecting the opening degree of the accelerator pedal 15 or the opening degree θ of the electronic throttle valve 13 operated for controlling the output of the engine 200, the accelerator opening degree and the throttle opening degree θ
Hydraulic pressure calculating means 2 for calculating the basic hydraulic pressure DT based on
03 and an engine output estimating means 204 for estimating an engine output based on an engine load, for example, an engine speed Ne, an accelerator opening θ, an intake air amount, a fuel injection amount, and the like.
Correction value calculating means 205 for calculating a correction value for correcting the basic oil pressure DT based on the estimated engine output.
And a hydraulic pressure calculating means 206 for calculating an actual hydraulic pressure based on at least the basic hydraulic pressure DT of the basic hydraulic pressure DT or the correction value.

In this embodiment, the shift state determining means 207 determines the shift state of the automatic transmission A, that is, the shift output and the start of the shift, and the calculation processing of the correction value calculating means 205 is performed based on the determination result. Will be

In this embodiment, the line pressure is controlled by the duty ratio of the linear solenoid valve SLT. This line pressure is the original pressure of the automatic transmission A, and the line pressure is controlled by a predetermined frictional engagement device. The pressure becomes the control oil pressure. For example, although not specifically shown, the line pressure may be supplied as a control oil pressure to a relay valve that controls the supply and discharge of the oil pressure to and from the third brake B3.

Hereinafter, control of the duty ratio of the linear solenoid valve SLT will be specifically described. First, the content of control of the duty ratio of the linear solenoid valve SLT differs between when the automatic transmission A is upshifted and when it is downshifted. When the automatic transmission A is upshifted, the duty ratio tDSLT of the linear solenoid valve SLT is controlled by the following equation.

TDSLT = DT + DTN + DTR + DTE (1) When the automatic transmission A is downshifted, the duty ratio tDSLT of the linear solenoid valve SLT is controlled by the following equation.

TDSLT = DT + DTN + DTR + DTE + DTH (2) In the above formulas (1) and (2), DT is a basic oil pressure (base value) DT, DTN which is set according to the throttle opening θ or the accelerator opening. , DTR is the correction amount of the basic oil pressure DT due to the ignition timing retard of the engine E, and DTE is the ratio of the engine speed Ne to the input speed NC0, in other words, the torque in the torque converter 20. The correction amount of the basic hydraulic pressure DT based on the ratio, and DTH indicates the correction amount of the basic hydraulic pressure DT based on the depression speed of the accelerator pedal 15.

The correction amount DTN of the basic oil pressure DT corresponding to the engine torque is obtained by interpolating the oil pressure control map with the value tDTEN calculated by the following equation from the shift output of the automatic transmission A to the start of the shift. The automatic transmission A
After the start of the shift, the correction amount DTN is held. A flag WIDL indicating that the idle switch has been turned on.
If = 1 holds, the output of the engine E becomes unstable, so that control is performed to set the correction amount DTN = 0.

TDTEN = tTEN−TE0 (3) In the above equation (3), tTEN is an engine torque map and a map interpolation value corresponding to the basic oil pressure DT.
0 is a basic engine torque.

The correction amount DTR of the basic oil pressure DT by the retard control of the engine E is calculated by interpolating the oil pressure control map with the value tDTER calculated by the following equation from the shift output of the automatic transmission A to the start of the shift. Desired. Note that the correction amount DTR is held after the automatic transmission A starts shifting.
When the flag WIDL = 1, which indicates that the idle switch has been turned on, is satisfied, the output of the engine E becomes unstable, so that control is performed to set the correction amount DTR = 0.

TDTER = TE0−TE (4) In the above equation (4), TE0 is the basic engine torque, and TE is the engine torque after the retard correction.

The correction amount DTE of the basic oil pressure DT based on the torque ratio is obtained by interpolating the oil pressure control map with the value tDTEE calculated by the following equation from the shift output of the automatic transmission A to the start of the shift. Note that the correction amount DTE is held after the automatic transmission A starts shifting. Also, a flag WIDL = 1 indicating that the idle switch has been turned on
Holds, the output of the engine E becomes unstable, so that control is performed to set the correction amount DTE = 0.

TDTEE = TRQHI / TRQHI0 (5) In the above equation (5), TRQHI0 is a torque ratio corresponding to the basic oil pressure DT, a map interpolation value, and TRQHI is a current torque ratio.

Further, the correction amount DTH of the basic oil pressure DT based on the depression speed of the accelerator pedal 15 is determined by the automatic transmission A
During the period from the shift output to the start of the shift, the hydraulic control map is obtained by interpolating the hydraulic control map with the value DELTT calculated by the following equation.

DELTAT = TT−TTACT (6) In the above equation (6), TT is calculated by pre-reading the torque input to the automatic transmission operation A after a predetermined time from the estimated value of the output torque of the engine E. TTACT is an estimated value of the actual output torque of the engine E calculated after a predetermined time.

The basic hydraulic pressure D corresponding to the engine torque
The correction amount DTN of T, the correction amount DTR of the basic oil pressure DT by retarding the ignition timing of the engine E, the correction amount DTE of the basic oil pressure DT by the torque ratio, and the correction amount DTH of the basic oil pressure DT by the depression speed of the accelerator pedal 15 are: The electronic throttle valve 13 by the electronic control unit 19 for the automatic transmission.
Control request (TTASFTUP> 0) or a request to close the electronic throttle valve 13 (TTASFT <TA).
When the (MAX) control is performed, the value of the control previously performed is held until the control is completed. At the start of control, each correction value is calculated again.

In the hydraulic control device for an automatic transmission having the above-described hardware configuration, a sensor for detecting data for estimating the output torque of the engine, for example, an engine speed sensor, an intake air amount sensor, a throttle opening sensor If the flag XFTRQ = 1 indicating a failure, such as a failure, is established, there is a possibility that an error occurs in the detection accuracy of the estimated engine output or the engine output value cannot be estimated. , The duty ratio of the linear solenoid valve SLT is controlled.

Further, according to the hydraulic control apparatus for an automatic transmission of this embodiment, for 100 msec from the start of the control, the tDSLT is subjected to a smoothing process represented by the following equation to obtain the final output value of the linear solenoid valve SLT. The DSLT is determined. For this reason, the fluctuation of the signal pressure of the linear solenoid valve SLT is averaged and changes gradually.
After 100 msec has elapsed from the start of the control, the above-mentioned tDSLT is used as a final output value.

DSLT _i = {tDSLT + DSLT _i-1 * (α-1)} / α (7) In the above equation (7), DSLT1 _i-1 is the last output value of the previous time, and α is a smoothing constant. In this embodiment, α = 1 is used when the oil temperature of the automatic transmission fluid is lower than a predetermined set temperature, and a value larger than 1 is used when the oil temperature of the automatic transmission fluid is higher than the set temperature. This is because the viscosity of the automatic transmission fluid is taken into account.

If the flag XFTHO = 1 indicating that the oil temperature sensor has failed, the above-mentioned smoothing process is not performed. This is because the viscosity of the automatic transmission fluid, which is the operating oil of the automatic transmission A, cannot be detected.

As described above, according to this embodiment, the output state of the engine E is detected based on the accelerator opening and the throttle opening θ, and the basic hydraulic pressure D is determined based on the output state.
T is calculated. On the other hand, the engine output is estimated based on the engine speed Ne, the throttle opening θ, the intake air amount, and the fuel injection amount, and the correction values DTN, DTR, DTE, DTH of the basic hydraulic pressure DT are determined based on the estimated engine output. Is calculated. Then, the basic hydraulic pressure DT becomes the correction value DTN, D
The output value tDSLT of the linear solenoid valve SLT is calculated after being corrected by TR, DTE, and DTH.

According to the present invention, as shown in FIG. 1, the state of the accelerator opening and the throttle opening θ is detected by the output control state detecting means 202.
The output control state of 0 can be reliably detected. Further, the basic oil pressure DT is calculated based on the detected output control state of the engine 200.

On the other hand, the engine output is estimated by the engine output estimating means based on the engine load, for example, the engine speed Ne, the throttle opening θ, the intake air amount, the fuel injection amount, and the like. A correction value is calculated based on the engine output.

The final output value tDSLT is calculated based on the basic oil pressure DT and the correction value, and the final output value tDLT is calculated.
The line pressure is controlled by the signal of the SLT. Therefore, it is possible to control the oil pressure of the friction engagement device engaged during the shift transition period of the automatic transmission A reliably and with high accuracy, and the shift shock suppressing function and the operation responsiveness are improved.

Further, when a failure occurs in the engine output estimating means 204 and it becomes impossible to estimate the engine output, the oil pressure of the friction engagement device is calculated based on the basic oil pressure DT. Therefore, even when the engine output estimating means 204 fails, the engagement pressure of the friction engagement device can be controlled to an appropriate state.

Here, the characteristic configurations of the present invention disclosed based on the above specific examples are listed as follows.

That is, when the engine output cannot be estimated by the engine output estimating means due to failure of various sensors, the correction value is set to zero. That is, the hydraulic pressure is determined only by the basic hydraulic pressure.

When the engine torque is reduced by a request from the automatic transmission, the reduced torque is not included in the estimated value of the engine output estimating means.

Further, a map of the engine torque corresponding to the basic hydraulic pressure is set, and a correction value of the basic hydraulic pressure is determined based on a difference between the torque value of the map and the current output estimated value.

Further, when the hydraulic pressure is used as the shift hydraulic pressure, when the start of the shift of the automatic transmission is detected, the correction value based on the engine output estimated value is not updated. This is to prevent the output estimated value from fluctuating due to a change in the engine speed when the shift is started, and to prevent the oil pressure from fluctuating.

Further, as an estimated engine output value,
Using the predicted torque estimated value and the actual torque estimated value, a correction value is determined based on the relationship between the two, for example, magnitude, difference, and ratio.

[0080]

As described above, according to the present invention,
Since the engine output is detected based on the state of the output control device operated by the driver, the output control state of the engine can be reliably detected. Further, the basic hydraulic pressure is calculated based on the detected output control state of the engine. On the other hand, an engine output is estimated based on the engine load, and a correction value is calculated based on the estimated engine output. Then, since the hydraulic pressure is calculated by correcting the basic hydraulic pressure with the correction value, the hydraulic pressure of the friction engagement device can be controlled reliably and with high accuracy, and the shift shock suppression function and operation responsiveness are improved. Is improved.

Further, when a failure occurs in the engine output estimating means and it becomes impossible to estimate the engine output, the oil pressure of the friction engagement device is calculated based on the basic oil pressure. Therefore, even when the engine output estimating means fails, the engagement pressure of the friction engagement device can be controlled to an appropriate state.

[Brief description of the drawings]

FIG. 1 is a block diagram showing the present invention by functional means.

FIG. 2 is a block diagram for explaining an overall control system of an engine and an automatic transmission to which the present invention is applied.

FIG. 3 is a skeleton diagram illustrating an example of a gear train in the automatic transmission according to the present invention.

4 is an operation chart showing an engaged / disengaged state of a friction engagement device for setting each shift speed of the automatic transmission shown in FIG. 3;

FIG. 5 is a diagram showing a part of a hydraulic circuit of the hydraulic control device shown in FIG. 2;

[Explanation of symbols]

 A, 201 Automatic transmission E, 200 Engine 17 Electronic control device for engine 18 Hydraulic control device 19 Electronic control device for automatic transmission 15 Accelerator pedal 13 Electronic throttle valve 202 Output control state detecting means 203 Basic hydraulic calculating means 204 Engine output estimation Means 205 Correction value calculating means 206 Hydraulic pressure calculating means

 ──────────────────────────────────────────────────続 き Continued on the front page (72) Inventor Noriaki Asahara 1 Toyota Town, Toyota City, Aichi Prefecture Inside Toyota Motor Corporation (72) Inventor Keihan Fukumura 1 Toyota Town, Toyota City, Aichi Prefecture Toyota Motor Corporation Inside

Claims (1)

[Claims] 1. A hydraulic control device for an automatic transmission for controlling a hydraulic pressure of a friction engagement device engaged during a gear shift of an automatic transmission connected to an engine, wherein the hydraulic control device is operated to control an output of the engine. Output control state detecting means for detecting the state of the output control device, basic hydraulic pressure calculating means for calculating a basic hydraulic pressure based on the state of the output control apparatus, and engine output estimating for estimating engine output based on the load of the engine Means, a correction value calculating means for calculating a correction value for correcting the basic oil pressure based on the estimated engine output, and the friction engagement based on the basic oil pressure or at least the basic oil pressure among the correction values. A hydraulic pressure control device for an automatic transmission, comprising: hydraulic pressure calculation means for calculating the hydraulic pressure of the device.