JPH1137264A - Hydraulic pressure control device of automatic transmission - Google Patents

Abstract

PROBLEM TO BE SOLVED: To control speed change with no speed change shock and delay of speed change by controlling the hydraulic pressure at the speed change time, based on the amount related to an input torque detected before speed change when the speed change is detected. SOLUTION: At the speed change operation time, in a computer 30 for automatic speed change control, an engine torque deviation ΔTE is compared with a standard value A for a prescribed time from this point of time. This standard value A divides ΔTE to plural areas and a correction value is set per respective areas and it is judged whether ΔTE before a prescribed time from a speed change output is bigger than the standard value A. Here, YES is judged, the biggest value α out of the advancely set value is adopted as the correction value of a duty ratio. While, when ΔTE is below the standard value A, it is judged whether ΔTE is bigger than the other standard value B and when YES is judged, a smaller value β than the value α is adopted as the correction value of the duty ratio and the correction amount for reducing the oil pressure of a friction engagement device is minified.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a device for controlling a hydraulic pressure in an automatic transmission for a vehicle, and more particularly to a device for controlling a hydraulic pressure of a friction engagement device at the time of shifting.

[0002]

2. Description of the Related Art In a conventional stepped automatic transmission, a friction engagement device such as a clutch or a brake is engaged or disengaged by hydraulic pressure, thereby changing a torque transmission path and executing gear shifting. I have. The rotation change in that case is
The output torque is smoothly changed by absorbing the slippage of the friction engagement device. Therefore, it is desirable to control the oil pressure of the friction engagement device that is engaged or released at the time of gear shifting to a pressure that changes the torque capacity of the friction engagement device so that a change in rotation occurs smoothly.

[0003] Conventionally, in addition to controlling the line pressure, which is the overall source pressure of the automatic transmission, based on the engine load such as the throttle opening, the output torque of the engine is estimated, and based on the estimated value. Devices for controlling the hydraulic pressure of the friction engagement device have been developed. According to such control,
Since the input torque of the automatic transmission can be reflected in the hydraulic pressure in real time, in principle, the hydraulic pressure is suitable for the input torque or the torque applied to the friction engagement device, and there is no shift shock or shift delay. Speed change control becomes possible.

[0004]

The above-described control for detecting or estimating the output torque of the engine and reflecting the detected output torque on the hydraulic pressure of the automatic transmission basically involves adjusting the hydraulic pressure of the friction engagement device during gear shifting to an appropriate level. It is to make it. However, a shift in the automatic transmission occurs when the accelerator pedal is depressed to increase the engine load, or conversely, when the engine load is reduced. Therefore, when shifting is performed by an automatic transmission, the throttle opening (engine load) often changes continuously or discontinuously. In such a case, it is difficult to accurately detect or estimate the output torque of the engine because the amount of intake air or the amount of fuel supplied to the engine changes. In particular, in the case of the engine, when the throttle opening changes, the intake air amount and the actual fuel supply amount to the cylinder do not change at the same time, but the engine torque changes with a delay and the change Because it can be complicated
It is difficult to accurately detect or estimate the engine torque in such a transient state.

[0005] Therefore, in the conventional device for controlling the hydraulic pressure during shifting based on the detected or estimated engine torque, the detected or estimated value of the engine torque during shifting does not always match the actual engine torque. On the contrary, in many cases, the hydraulic pressure of the automatic transmission becomes an inappropriate pressure with respect to the input torque. As a result, the shift shock deteriorates and, conversely, the shift time becomes longer and the shift feeling is delayed or sluggish. It was likely to happen.

[0006] In addition, the aging of a power source such as an engine is as follows.
Usually appears as a decrease in output torque. Therefore, if such changes are always reflected in the hydraulic pressure of the automatic transmission to optimize the hydraulic pressure, the hydraulic pressure will be lower than the initial set value. However, in a steady state other than during shifting, even if the hydraulic pressure is somewhat high, there is no inconvenience such as a shortage of the torque capacity of the friction engagement device. The control of the hydraulic pressure based on this does not solve the inconvenience in practical use, but rather performs excessive or useless control.

SUMMARY OF THE INVENTION The present invention has been made in view of the above-mentioned circumstances, and a shift control without shift shock or shift delay is realized by more accurately reflecting the input torque during shifting to the hydraulic pressure of the friction engagement device. It is an object of the present invention to provide a hydraulic control device that can perform the control.

[0008]

SUMMARY OF THE INVENTION In order to achieve the above object, a first aspect of the present invention is to set a shift speed by engaging a friction engagement device by hydraulic pressure.
And in a hydraulic control device for an automatic transmission that controls the hydraulic pressure based on the amount related to the input torque, means for detecting an amount related to the input torque, means for detecting a shift,
Means for controlling a hydraulic pressure during the shift based on an amount related to an input torque detected before the shift when the shift is detected.

Therefore, in the device of the present invention, the hydraulic pressure at the time of shifting is controlled based on an amount related to the input torque before the shifting is determined or started. The amount related to the input torque includes a detected value or an estimated value of an output torque of a power source such as an engine, which is a value before shifting in the automatic transmission, and is a torque due to an acceleration request or a deceleration request. Is a value in a stable state before the fluctuation of. Therefore, the input torque in the stable state immediately before the shift is reflected on the hydraulic pressure during the shift, and the hydraulic pressure of the friction engagement device becomes an appropriate pressure with respect to the torque applied thereto. Effectively prevented.

[0010]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS Next, the present invention will be described based on a specific example shown in the drawings. First, the automatic transmission 10 as a target in this specific example will be described. As shown in FIG.
The automatic transmission 10 has a hydraulic control device 20 for controlling a torque converter and shifting control incorporated therein.
Is configured to control engagement / disengagement of a friction engagement device for achieving pressure adjustment and a shift speed of a hydraulic pressure generated by an oil pump and engagement / release of a lock-up clutch in a torque converter. ing. Such control is basically configured to be performed electrically,
For this purpose, the hydraulic control device 20 includes first to third solenoid valves SL1, SL2, S
L3, a fourth solenoid valve SL4 for mainly controlling the engine braking state, and a linear solenoid valve S for mainly controlling the lock-up clutch and the third brake pressure.
An LU, a linear solenoid valve SLT for mainly controlling the line pressure, and a linear solenoid valve SLN for mainly controlling the accumulator back pressure are provided. An automatic transmission control computer 30 for controlling these solenoid valves is provided.

The automatic transmission 10 is connected to an engine E which is a power source in a vehicle-mounted state. An electronic throttle valve 12 driven by an actuator 11 such as a servomotor is disposed in an intake pipe 13 of the engine E. The throttle opening is controlled by driving the actuator 11 according to the amount of depression of an accelerator pedal 14. Is appropriately controlled. An engine control computer 50 for controlling the entire engine E including the actuator 11 is provided.

These computers 30 and 50 are both a central processing unit (CPU) and a storage device (RA).
M, ROM) and an input / output interface, and performs calculations based on various data indicated by reference numeral 40 in FIG.
Is configured to be controlled. In particular,
The engine control computer 50 receives, in addition to the signal of the opening degree (depressed amount) of the accelerator pedal 14, the engine rotation speed, intake air amount, intake air temperature, throttle opening degree, vehicle speed, engine water temperature, and signals from the brake switch. Based on these signals, the actuator 11 is driven based on these signals to control the throttle opening, and to control the fuel injection amount and the ignition timing. The engine control computer 50 also controls the engine E based on the engine speed, throttle opening, intake air amount, and data (map value) stored in advance.
Is calculated and the actual engine torque is estimated.

On the other hand, the automatic transmission control computer 30 includes a throttle opening, a vehicle speed (output shaft rotation speed), a signal from a brake switch, a shift position, a signal from a pattern select switch, a cruise signal, and a C0 sensor S.
A signal from N1, a signal from C2 sensor SN2, a transmission oil temperature, a signal from a manual shift switch, and the like are input, and are calculated together with pre-stored data such as a shift map (shift map). , And is configured to execute shift control and hydraulic control. These computers 30 and 50 are connected so as to mutually transmit data. As one example, a basic engine torque and an estimated engine torque value are transmitted from the engine control computer 50 to the automatic transmission control computer 30.

The gear transmission mechanism in the automatic transmission 10 will be described with reference to FIG. 4. The gear transmission mechanism includes a sub-transmission mechanism D having a front overdrive configuration,
A 5-speed configuration combining a simple connection 3 planetary gear train configuration with a forward 4-speed reverse 1-speed main transmission mechanism M,
This mechanism is connected to a torque converter T with a lock-up clutch L.

The auxiliary transmission mechanism D includes a one-way clutch (OWC) F-0 and a multi-plate clutch C-parallel to the sun gear S0, the carrier C0 and the ring gear R0.
0 and a multi-disc brake B-0 in series therewith. On the other hand, the main transmission mechanism M includes three gear units P1 in a simple connection in which respective transmission elements including sun gears S1 to S3, carriers C1 to C3, and ring gears R1 to R3 are appropriately connected.
, P3 to P3, and the multiple disc clutches C-1, C-2, the band brake B-1, the multiple disc brakes B-2 to B-4, and the one-way clutch (OWC) in relation to the speed change elements of the respective gear units P1 to P3. ) F-1 and F-2 are provided. In the drawing, reference numeral SN1 denotes a C0 sensor for detecting the rotation of the drum of the clutch C-0, and SN2 denotes a C2 sensor for detecting the rotation of the drum of the clutch C-2. Although not shown, each clutch and brake has a hydraulic servo composed of a piston-cylinder mechanism for engaging and disengaging the friction members.

The shift range and each shift speed set by the automatic transmission 10 are as shown in the engagement operation table of FIG. 5, and each shift speed will be briefly described below. In addition,
In FIG. 5, the symbol 〇 indicates engagement, the symbol 係 合 indicates engagement during engine braking, the symbol ◎ indicates engagement without being involved in torque transmission, and the blank indicates a released state.

The output torque of the engine E is transmitted to the input shaft N of the auxiliary transmission mechanism D via the torque converter T. The torque of the input shaft N is controlled by the hydraulic control device 20 so that the clutch C-0 is engaged to
When the clutch C-1 of the main transmission mechanism M is engaged and all the other frictional engagement devices are disengaged, the input is input to the gear unit P3, and the reverse of the ring gear R3 is performed by the one-way clutch F-2. Rotation is blocked and carrier C3
Is output to the output shaft U as the first speed rotation.

In the second speed, the auxiliary transmission mechanism D is directly connected.
This is achieved when the clutch C-1 and the brake B-3 are engaged. At this time, the torque input to the gear unit P2 is controlled by using the carrier C1 of the gear unit P1 as a reaction force element and the carrier C2 of the gear unit P2 and the carrier C2. It is output to the ring gear R1 of the directly connected gear unit P1, and is output to the output shaft U as the second speed rotation.

Similarly, in the third speed, the sub-transmission mechanism D is directly connected, and the clutch C-1 and the brake B-2 are engaged.
This is achieved by releasing other frictional engagement devices.
At this time, the torque input to the ring gear R2 of the gear unit P2 is output as the third speed rotation from the output shaft U via the carrier C2 using the sun gear S2 as a reaction force element.

Further, in the fourth speed, similarly, the auxiliary speed change mechanism D
Is directly connected and the clutch C-1 and the clutch C-2 are engaged together. At this time, since the gear unit P2 is directly connected to the ring gear R2 and the sun gear S2, the input torque is output as it is. In the fifth speed, the main speed change mechanism M is in the same state as the above-described fourth speed, and in response to this, the clutch C-0 of the sub speed change mechanism D is released and the brake B-0 is engaged. The sun gear S0 is fixed so that the auxiliary transmission mechanism D
At a high speed.

In the reverse gear, the sub-transmission mechanism D is brought into the above state, and the clutch C-2 and the brake B-
4 is achieved. At this time, the torque input to the sun gear S2 of the gear unit P2 is output as reverse rotation of the carriers C2 and C3 of the gear units P2 and P3 using the ring gear R3 as a reaction force element.

The above-described friction engagement devices such as clutches and brakes are configured to have a necessary and sufficient torque capacity in a range where slippage does not occur due to applied torque. It is controlled to have a capacity. Therefore, the hydraulic control device 20
Has a built-in configuration for generating a hydraulic pressure according to the torque. One example is shown in FIG.

In FIG. 6, a primary regulator valve 61 for adjusting oil pumped by an oil pump 60 to a line pressure PL is provided. The primary regulator valve 61 is a spool type valve. An oil pump 60 is connected to an input / output port 62, and an orifice 63 is connected to the input / output port 62.
A feedback port 64 communicated with the spring 65 is formed on the opposite side of the spool 65 across the spool. A linear solenoid valve SLT is connected so that the signal pressure PSLT acts in the same direction as the pressing force of the spring 65. Therefore, the axial force generated by the spring force and the signal pressure PSLT is opposed to the feedback pressure, and when the axial force generated by the feedback pressure is larger, the input / output port 62 is communicated with the drain port 66 to provide the spring force and the signal. The line pressure PL corresponding to the pressure PSLT is generated in the line pressure oil passage 67.

The linear solenoid valve SLT is a valve that is duty-controlled and outputs a signal pressure PSLT corresponding to the duty ratio DSLT. Its duty ratio DSL
The relationship between T and the signal pressure PSLT is as shown in FIG. 6, and the signal pressure PSLT is configured to decrease as the duty ratio DSLT increases.

The items for determining the duty ratio DSLT include items determined in accordance with the throttle opening, and the duty ratio DSLT increases as the throttle opening increases.
SLT is controlled to be small. That is, when the engine output is increased due to an increase in the throttle opening, the pressure regulation level of the primary regulator valve 61 is increased. As a result, the line pressure PL is increased and the engagement pressure of the friction engagement device, that is, The capacity increases. Therefore, the torque capacity of the friction engagement device becomes a value suitable for the torque to be applied, and a predetermined gear position can be set. At the same time, excessive hydraulic pressure is not generated, so that power loss is prevented.

As shown in FIG. 6, the hydraulic control device 20 is provided with an accumulator control valve 68. The accumulator control valve 68 is connected to the accumulator 6 connected to the clutches C-1 and C-2 and the brakes B-0 and B-2.
A pressure control valve for controlling the back pressure of 9, 70, 71, 72. The pressure control valve is configured to change the transient pressure control level by the linear solenoid valve SLN and change the basic oil pressure by the linear solenoid valve SLT. ing. That is, the accumulator control valve 68 has a spool for selectively communicating an input port 73 to which the line pressure PL is input with an output port 74 and a drain port 75, and a spring 76 at one end thereof.
And a linear solenoid valve SLT is connected to generate a pressing force by the signal pressure PSLT in the same direction as the spring 76. On the opposite end, a feedback port 78 is formed which communicates with the output port 74 via an orifice 77. The output port 74 is connected to each of the accumulators 69, 70,
The back pressure chambers 71 and 72 communicate with each other. Therefore, when the signal pressure PSLT of the linear solenoid valve SLT increases,
It is configured such that the pressure regulation level increases and the output pressure, that is, the accumulator back pressure increases. Note that the pressure regulating action by the linear solenoid valve SLN is the same as that known in the art, and a description thereof will be omitted.

In the control items for the duty ratio of the linear solenoid valve SLT, in addition to the items based on the throttle opening described above, items based on the detected or estimated engine torque obtained from the engine control computer 50 are set. During shifting, the signal pressure PSLT of the linear solenoid valve SLT is increased or decreased according to the detected or estimated engine torque to increase or decrease the accumulator back pressure, that is, the oil pressure of the friction engagement device. An example of the control will be described below.

FIG. 1 shows a control routine for obtaining a difference ΔTE between the basic engine torque TE0 and the estimated engine torque TE, and this routine is repeated every several tens of msec. First, in step 1, the difference t between the basic engine torque TE0 and the estimated engine torque TE at that time is calculated.
Calculate ΔTE. Here, the basic engine torque TE0 is
It is calculated based on the engine speed NE and the throttle opening, and is a value uniquely set for each type or type of engine E. On the other hand, the estimated engine torque TE is a value determined on the basis of the intake air amount, the engine speed, the map, and the like at that time. The value is smaller than the torque.

Each time the routine shown in FIG. 1 is executed, the above-mentioned difference tΔTE is obtained, and the difference tΔTE is calculated a plurality of times (4 in the example of FIG. 1).
Are averaged to obtain a difference ΔTE between the basic engine torque TE0 and the estimated engine torque TE (step 2).
This is to eliminate an error due to disturbance or the like. Then, the value ΔTE is sequentially stored in the storage device RAM (step 3). Therefore, the difference ΔTE between the engine torque uniquely determined based on the throttle opening and the like and the actual engine torque is sequentially obtained and stored.

FIG. 2 shows a control routine for correcting the oil pressure of the friction engagement device at the time of shifting, based on the deviation ΔTE of the engine torque obtained as described above. That is, when the shift output (step 11) is performed, the engine torque deviation ΔTE for a predetermined time T0 seconds from that point is compared with the reference value A (step 12). The reference value A is used to divide the engine torque deviation ΔTE into a plurality of regions and set a correction value for each region. In step 12, the engine torque deviation ΔTE at a predetermined time T0 seconds before the shift output is calculated. It is determined whether the value is larger than the reference value A. If an affirmative determination is made in step 12, the torque actually output from the engine E at the time of shifting is greatly reduced with respect to the basic engine torque obtained from the throttle opening and the engine speed. In this case, the correction value ΔDSLT of the duty ratio DSLT
The largest value α among the values set in advance is adopted (step 13).

As described above, the linear solenoid valve S
Since LT has a configuration in which the output signal pressure PSLT decreases as the duty ratio DSLT increases, it is estimated that the engine torque deviation ΔTE is large, that is, the actual engine torque is greatly reduced from the basic engine torque. As a result, the duty ratio DSLT is increased, and the signal pressure PSLT is reduced accordingly, so that the pressure regulation level of the accumulator control valve 68 is reduced. The correction is made to be lower than the pressure determined based on the opening.

On the other hand, if the result of the determination in step 12 is negative due to the fact that the engine torque shift ΔTE is equal to or smaller than the reference value A, the routine proceeds to step 14, where it is determined whether or not the engine torque deviation ΔTE is larger than another reference value B. Is determined. The reference value B is smaller than the reference value A. Therefore, if the result of the determination in step 14 is affirmative, the engine torque deviation .DELTA.TE is in an intermediate region. In this case, the duty ratio A value β smaller than the above value is adopted as the correction value ΔDSLT of DSLT (step 15). As a result, since the engine torque deviation ΔTE is smaller than the above case, it is determined that the degree of decrease of the engine torque from the basic engine torque is small. Therefore, the correction value of the duty ratio DSLT is reduced, and
The amount of decrease in the pressure regulation level at the accumulator control valve 68, that is, the amount of correction for decreasing the oil pressure of the friction engagement device is made smaller than in the above case. If the engine torque deviation ΔTE is equal to or smaller than the second reference value β, it is determined that it is not necessary to correct the oil pressure of the friction engagement device, and the correction value ΔDSLT of the duty ratio DSLT is set to zero (step 1).
6).

Therefore, according to the control shown in FIG. 2, the hydraulic pressure of the friction engagement device at the time of shifting is corrected based on the estimated value of the engine torque at a point before (for example, about 0.5 seconds) the shifting output. Therefore, the estimated value of the engine torque does not include the value due to the transient torque fluctuation at the time of the shift, and as a result, the estimated value of the engine torque at the time of the shift becomes a value close to the actual engine torque, and at the same time, the friction engagement is performed. The hydraulic pressure of the device becomes a pressure suitable for the actual engine torque (input torque of the automatic transmission). Therefore, during shifting, the hydraulic pressure of the friction engagement device changes to a high value, and the output torque changes, that is, the shift shock increases. On the other hand, the hydraulic pressure of the friction engagement device is too low, causing a shift delay, or It is possible to prevent a shock due to the so-called end contact of the accumulator from occurring.

In the above example, the line pressure and the accumulator back pressure are corrected based on the deviation between the basic engine torque and the estimated engine torque.
According to the present invention, the oil pressure of the friction engagement device may be adjusted based on the estimated value of the actual engine torque. In short, the estimated value of the actual engine torque before (preferably immediately before) the shift output is obtained by: It is only necessary to reflect the change in the oil pressure of the friction engagement device at the time of shifting, so that an amount related to the input torque of the automatic transmission such as the engine torque may be detected. Therefore, steps 1 and 2 described above correspond to the means for detecting the amount related to the input torque of the present invention. The shift may be detected based on the running state and the shift map. Therefore, the above-described step 11 corresponds to the means for detecting a shift in the present invention.

According to the present invention, the oil pressure at the time of gear shifting may be controlled based on the engine torque estimated value before the gear shifting output. As described above, the engine torque deviation is divided into a plurality of regions and each oil pressure is controlled. Instead of setting the correction value, set the function of the oil pressure with the amount related to the input torque such as the engine torque estimated value as a parameter,
The hydraulic pressure may be continuously corrected according to the amount related to the input torque. Therefore, the above steps 13, 15, and 16 including such control correspond to the means for controlling the hydraulic pressure of the present invention. Further, the means for controlling the oil pressure of the friction engagement device at the time of gear shifting is not limited to the means for controlling the back pressure of the accumulator described above. The control amount may be corrected by an amount related to the input torque. The present invention can also be applied to a control device for an automatic transmission other than the automatic transmission having the gear train or the hydraulic circuit described above. Further, the power source to which the automatic transmission according to the present invention is connected is not limited to the engine, but may be another power source such as a motor.

[0036]

As described above, according to the control device of the present invention, when controlling the oil pressure of the friction engagement device during shifting based on the amount related to the input torque such as the engine torque, the control of the control is performed. Since the amount before the gear shift is used as the amount related to the input torque such as the base engine torque, the transient fluctuation due to the gear shift is not included in the amount related to the input torque and accurately reflects the actual situation As a result, the control accuracy of the hydraulic pressure of the friction engagement device is improved, and shift shock, shift delay, and the like can be reliably prevented. Further, the control device of the present invention executes the control of the hydraulic pressure based on the amount related to the input torque only at the time of shifting, so that useless control can be omitted, and the control system is simplified and the cost is reduced. Can be achieved.

[Brief description of the drawings]

FIG. 1 is a flowchart conceptually showing an example of a control routine for obtaining a deviation between a basic engine torque and an estimated engine torque value.

FIG. 2 is a flowchart conceptually illustrating an example of a control routine for determining a correction value of a duty ratio based on an engine torque deviation.

FIG. 3 is a block diagram showing an overall control system of the automatic transmission targeted by the present invention.

FIG. 4 is a skeleton diagram showing an example of a gear train of the automatic transmission.

FIG. 5 is a table showing an engaged / disengaged state of a friction engagement device for setting each shift speed.

FIG. 6 is a hydraulic circuit diagram schematically illustrating a part of a hydraulic circuit that controls a line pressure and an accumulator back pressure.

[Explanation of symbols]

 E Engine 10 Automatic transmission 20 Hydraulic control device 30 Automatic transmission control computer 50 Engine control computer 61 Primary regulator valve 68 Accumulator control valve B-0, B-2 Brake C-1, C-2 Clutch SLT Linear solenoid valve

Claims (1)

[Claims] An oil pressure control device for an automatic transmission for setting a gear position by engaging a friction engagement device by oil pressure and controlling the oil pressure based on an amount related to the input torque. Means for detecting a related amount; means for detecting a shift; and means for controlling a hydraulic pressure during the shift based on an amount related to an input torque detected before the shift when the shift is detected. A hydraulic control device for an automatic transmission, comprising: