JPH04185959A - Hydraulic control device for automatic transmission - Google Patents

Abstract

PURPOSE:To improve accuracy of line pressure control during a speed change by correcting a study correction amount, set in a line pressure correcting means, in accordance with whether the speed change of performing correction of the correcting means is the speed change or not of performing engine output torque decrease control by a torque decreasing means. CONSTITUTION:A controller 70 inputs signals from a turbine speed sensor 71, throttle opening sensor 72, oil temperature sensor 73 and a speed change mode select switch 74 and a signal for showing execution and non-execution of fuel cut control from an engine controller 75, to perform speed change control and lockup control based on a turbine speed, throttle opening and a selected speed change mode. Line pressure control during speed change and non-speed change is performed in accordance with output signals from the sensors 71 to 73, switch 74 and the engine controller 75. Here in the line pressure control during the speed change, correction control of a line pressure by a study of speed change time is performed and also correcting a line pressure correction amount in accordance with execution and non-execution of engine output torque decrease control during the speed change.

Description

DETAILED DESCRIPTION OF THE INVENTION (Field of Industrial Application) The present invention relates to a hydraulic control device for an automatic transmission, and particularly to a hydraulic control device for an automatic transmission that performs variable control of line pressure during gear shifting.

(Prior Art) Automatic transmissions generally installed in automobiles combine a torque converter and a speed change gear mechanism, and the power transmission path of this speed change gear mechanism is controlled by selectively operating multiple friction engagement elements such as clutches and brakes. This type of automatic transmission is equipped with a hydraulic control circuit that controls the supply and discharge of hydraulic pressure to the actuators of each of the friction engagement elements. .

This hydraulic control circuit includes a regulator valve that adjusts the discharge pressure of the oil pump to a predetermined line pressure, a manual valve that switches the range by manual operation, and a manual valve that operates depending on the operating state to switch the oil paths leading to each of the above actuators. Accordingly, a plurality of shift valves and the like are provided to selectively engage the plurality of frictional engagement elements. Also,
In recent years, a duty solenoid valve is used to change the line pressure adjustment value by the regulator valve according to operating conditions such as the throttle opening of the engine, and an ON-OFF valve is used to operate the shift valve when changing gears.
The precision of speed change control is improved by providing solenoid valves and the like and controlling these electrically.

On the other hand, in this type of automatic transmission, there is a problem in that a so-called shift shock occurs due to the engagement or release operation of the frictional engagement element during a shift, and this causes discomfort to the occupant. Therefore, in an automatic transmission in which line pressure is electrically controlled as described above, control is performed to reduce the line pressure during gear shifting to reduce shift shock.

In line pressure control during such a shift,
For example, as disclosed in Japanese Unexamined Patent Publication No. 2-57759, line pressure correction control is sometimes performed by learning the shift time. This control measures the time required for gear shifting, compares this shifting time with a preset target time, and corrects the line pressure during gear shifting to eliminate the deviation. It is.

(Problem to be Solved by the Invention) By the way, when performing correction control of line pressure by learning the shift time as described above, it is important to note that the shift at the time of measuring the shift time is a shift performed under a predetermined operating condition. It is necessary that there be. In other words, when shifting, in order to reduce shift shock, control is sometimes performed to reduce the engine output torque by, for example, stopping or reducing the amount of fuel supplied to the engine. In this case, even if the type of shift is the same, the shift time will be different from the original shift time under normal driving conditions. Therefore, unless the shift time is learned only when shifting in a predetermined operating state, line pressure correction will be performed incorrectly.

Therefore, in the invention disclosed in the above-mentioned publication, correction of the line pressure by learning the shift time is prohibited in the case of either a shift in which the output control of the drive source is performed or a shift in which the output control of the drive source is not performed. This is designed to avoid such erroneous learning. However, in this case, there will be fewer opportunities for learning, and the accuracy of line pressure control during gear shifting will not be sufficiently improved.

Therefore, when performing correction control based on learning of shift time regarding line pressure control during gear shifting, the present invention makes it possible to perform this learning operation regardless of the presence or absence of torque reduction control during gear shifting while avoiding erroneous learning. The object of the present invention is to further improve the accuracy of line pressure control during gear shifting.

(Means for Solving the Problems) In order to solve the above problems, the present invention is characterized by having the following configuration.

That is, the hydraulic control device for an automatic transmission according to the present invention has the following features:
A line pressure adjustment means provided in the hydraulic control circuit to adjust the line pressure supplied to the friction engagement element, and a line pressure adjustment means that adjusts the line pressure adjusted by the line pressure adjustment means between an actual shift time and a target shift time when changing the ratio. In an automatic transmission having a line pressure correction means for performing a learning correction according to a deviation, and a torque reduction means for reducing engine output torque under predetermined conditions during gear shifting, the learning correction amount set by the line pressure correction means is The present invention is characterized in that a correction amount correcting means is provided for making corrections depending on whether or not the corrected speed change is a speed change in which engine output torque reduction control is performed by the torque reduction means.

(Function) According to the above configuration, when shifting, the line pressure is corrected based on the learning of the shifting time, and the subsequent shifting is performed using the corrected line pressure. Pressure control is performed with high precision, and gear shifts are performed in an optimal shift time. Particularly, according to the present invention, the line pressure correction amount by the learning correction is corrected depending on whether or not the gear shift for which the correction was performed was performed with the engine output reduced. Therefore, appropriate correction will be made in either case. Therefore,
It is possible to increase the number of learning corrections while avoiding erroneous learning, and line pressure control during gear shifting can be performed with higher accuracy.

(Example) Examples of the present invention will be described below.

First, the mechanical configuration of the automatic transmission according to this embodiment will be explained with reference to FIG. It has a mechanism 30, a plurality of frictional engagement elements 41 to 46 such as clutches and brakes, and one-way clutches 51 and 52 that switch the power transmission path of the mechanism 30, and these act as a driving range. Each of the H ranges, 1st to 4th speeds in the D range, 1st to 3rd speeds in the 2nd range, and 1st to 2nd speeds in the lunge are available.

The torque converter 20 includes a pump 22 fixedly installed in a case 21 connected to the engine output shaft 1, and a turbine 23 disposed opposite to the pump 22 and driven by the pump 22 via hydraulic oil. , a stator 25 that is interposed between the pump 22 and the turbine 23 and is supported by the transmission case 1 via the one-way clutch 24 to increase torque; and the case 21 and the turbine 2
3, and a lock-up clutch 26 that connects the engine output shaft 1 and the turbine 23 via the case 21. Then, the turbine 23
The rotation is outputted to the speed change gear mechanism 30 side via the turbine shaft 27. Here, a pump shaft 12 passing through a turbine shaft 27 is connected to the engine output shaft 1, and the shaft 12 drives an oil pump 13 provided at the rear end of the transmission.

On the other hand, the speed change gear mechanism 30 is composed of a Ravigneau type planetary gear device, and includes a small sun gear 31 with a small diameter loosely fitted on the turbine shaft 27, and a small sun gear 31 that is loosely fitted onto the turbine shaft 27.
1, a large diameter large sun gear 32 loosely fitted on the turbine shaft 27, a plurality of short pinion gears 33 meshed with the small sun gear 31, and a front half meshed with the short pinion gear 33. The rear half is composed of a long pinion gear 34 meshed with the large sun gear 32, a carrier 35 rotatably supporting the long pinion gear 34 and the short pinion gear 33, and a ring gear 36 meshed with the long pinion gear 34. .

A forward clutch 41 and a first one-way clutch 51 are interposed in series between the turbine shaft 27 and the small sun gear 31, and a coast clutch 42 is interposed in parallel with these clutches 41 and 51. In addition, the turbine shaft 27 and carrier 35
A 3-4 clutch 43 is interposed between the turbine shaft 27 and the large sun gear 32, and a reverse clutch 44 is interposed between the turbine shaft 27 and the large sun gear 32. Further, a 2-4 brake 45 consisting of a band brake for fixing the large sun gear 32 is provided between the large sun gear 32 and the reverse clutch 44, and the carrier 35
A second one-way clutch 52 that receives and stops the reaction force of the carrier 35, and a low reverse brake 46 that fixes the carrier 35 are provided in parallel between the transmission case 1) and the transmission case 1). The ring gear 36 is connected to the output gear 14, and rotation is transmitted from the output gear 14 to left and right wheels (not shown) via a differential gear.

Table 1 summarizes the relationship between the operation of each of the friction engagement elements 41 to 46 and the one-way clutches 51 and 52 and the gear position.

(Hereinafter, blank space) On the other hand, as shown in FIG. 2, this automatic transmission has the above-mentioned frictional engagement elements 41 to 46 selectively operated according to Table 1 to change gears according to the operating state. A hydraulic control circuit 60 is provided for forming. This hydraulic control circuit 6
0 includes first to third solenoid valves 61 to 63 for changing gears that switch the engagement pressure supply oil passages leading to each of the friction engagement elements 41 to 46, and a fourth solenoid valve 64 for controlling the lock-up clutch 26. A duty solenoid valve 65 for controlling line pressure is provided.

A controller 70 that controls these solenoid valves 61 to 65 is provided, and the controller 70 includes a turbine rotation speed sensor 71 that detects the turbine rotation speed of the torque converter 20 (or a vehicle speed sensor that detects the vehicle speed of the vehicle). , a signal from the throttle opening sensor 72 that detects the throttle opening of the engine, a signal from the oil temperature sensor 73 that detects the temperature of the hydraulic oil, and a shift mode selection switch operated by the driver. A signal from the sensor 71.74 and a signal indicating execution or non-execution of fuel cut control from the engine controller 75 that controls the engine are inputted to the sensor 71.7.
Shift control and lock-up control are performed based on the turbine rotational speed (or vehicle speed), throttle opening, and selected shift mode indicated by output signals from 2 and switch 74, and each sensor 71 to 73 and switch According to output signals from the engine controller 74 and the engine controller 75, line pressure control is performed during shifting and during non-shifting. Here, in this embodiment, there is an economy mode where the shift point is set on the low turbine rotation speed (low vehicle speed) side, which emphasizes fuel efficiency, and an economy mode where the shift point is set on the high turbine rotation speed (high vehicle speed) side. It is said that it is possible to select a power mode that emphasizes dynamic performance.

Next, as shown in FIG. 3, a hydraulic control circuit 60 supplies and discharges hydraulic pressure to the actuators of each of the frictional engagement elements 41 to 46.
I will explain about it.

This hydraulic control circuit 60 is equipped with a regulator valve 81 that adjusts the pressure of hydraulic oil discharged from the oil pump 13 to the main line 101 shown in FIG. 1 to a predetermined line pressure. The line pressure is variably controlled by the duty solenoid valve 65. In other words, the hydraulic pressure kept at a constant pressure by the solenoid reducing valve 82 is adjusted by the duty solenoid valve 65 to a pilot pressure corresponding to the duty rate (ON time ratio during one ON-OFF cycle), and this pilot pressure is introduced into the pressure booster boat 81a of the regulator valve 81 via the line 102, whereby the line pressure is adjusted to a value corresponding to the pilot pressure or the duty rate.

This line pressure is supplied to the input port a of the manual valve 83 that is manually operated by the driver.
shift position (range) P.

According to R, N, D, 2.1, output port b.

It is now selectively output to c, d, and e. That is, in the D range and 2 range, the signal is output to boats b and c, in the Lungin mode, it is output to boats b and d, and in the R range, it is output to boat e.

Boat b, to which line pressure is output in each of the forward ranges D and 2.1, is communicated with a 1-2 shift valve 84 via a line 103. This 1-2 shift valve 8
4 is switched and operated by the first solenoid valve 61, and in the first gear, the solenoid valve 61 is OFF.
(closed), the spool 84a is closed (
The same applies hereafter) The line 104 located on the left side and leading to the apply chamber 45a of the 2-4 brake 45 is drained. Also, in the 2nd to 4th speeds, the first solenoid valve 61 is turned on.
(open), the spool 84a is positioned on the right side, the line 103 led from the boat b is communicated with the line 104, and the line pressure is introduced into the apply chamber 4ga of the 2-4 brake 45. . Furthermore, this one
-2 shift valve 84 is supplied with line pressure from the boat d through a line 105 provided with a low reducing valve 85 in the first gear of the lunge, and is supplied to the low reverse brake 46 through lines 106 and 107. supply to.

Further, the line pressure from the boat b of the manual valve 83 is supplied to 2-3 via lines 103, 108, 109.
Pilot pressure is supplied to one end of the shift valve 86, and line pressure is supplied to the 2-3 shift valve 86 from the boat C via line 1)0. and,
The pilot pressure is supplied and discharged by the second solenoid valve 62, so that the solenoid valve 62 is turned ON (open) in the 1st and 2nd speeds, the 2-3 shift valve 8
6 spool 86a is located on the right side, the line 1) 1 leading to the 3-4 clutch 43 is drained, and the OF
In the 3rd and 4th speeds, which are F (closed), the spool 86a is located on the left side and supplies the line pressure from the line 1)0 to the 3-4 clutch 43 via the line 1)1.

The line 1) 1 is also led to a 3-4 shift valve 87. This 3-4 shift valve 87 is the third
The supply and discharge of pilot pressure is controlled by the solenoid valve 63, and the solenoid valve 63 is turned ON (open).D
In the 1st, 2nd, and 4th speed ranges and the 1st speed in the 2nd range, the spool 87a is located on the right side, thereby draining the line 1)2 leading to the release chamber 45b of the 2-4 brake 45. Furthermore, in the 3rd gear of the D range, the 2nd and 3rd gears of the 2nd range, and the 1st and 2nd gears of the lunge, when the solenoid valve 63 is 0FF (closed), the spool 87a of the 3-4 shift valve 87 is located on the left side. 2-3 shift valve 86
The line 1) 1 connected to the above line 1) 2 is connected to the above line 1) 2.

Therefore, line pressure is supplied to the release chamber 45b of the 2-4 brake 45 depending on the position of the spool 86a of the 2-3 shift valve 86.

Further, this 3-4 shift valve 87 connects or disconnects the line 1) 3 connected to the boat b of the manual valve 83 via the line 103° 108 and the line 1) 4 leading to the coast clutch 42. By doing so, the operation of the coast clutch 42 is controlled.

In this way, the first to third solenoid valves 61 to 6
3, each frictional engagement element is selectively engaged according to the required gear position according to Table 1. This hydraulic control circuit 60
In order to reduce shift shock, etc., in addition to the above-mentioned valves, 1-2 accumulator 88, 2-3 accumulator 89, 2-3 timing valve 90, N-D accumulator 91, N-R accumulator 92. A 3-2 timing valve 93, a bypass valve 94, etc. are provided. Furthermore, the supply and discharge of pilot pressure is controlled by the fourth solenoid valve 64, and a lock-up control valve 95 engages or disengages the lock-up clutch 26, and a converter relief valve adjusts the operating pressure supplied to the torque converter 20. 96
etc. are provided.

Next, the specific operation of line pressure control by the controller 70 and the duty solenoid valve 65 will be explained according to the flowcharts shown in FIG. 4 and subsequent figures.

The flowchart in FIG. 4 shows the main routine of line pressure control. In this routine, first, step S1
If it is determined that it is not the time to shift, the line pressure control during non-shifting in step S2 is performed, and if it is determined that it is the time to shift,
The line pressure control during the shift in step S3 is executed. If it is determined that it is time to shift, it is determined in step S4 whether or not the shift is an upshift, and when the shift is up, line pressure correction control based on learning of the shift time in step S5 is performed. At the time of downshifting, line pressure correction control is performed by learning the revving rotation speed in step S6.

The line pressure control during non-shifting in step S2 is performed as follows according to a subroutine shown in a flowchart in FIG.

In this subroutine, in step S□1+812, the throttle opening degree of the engine and the turbine rotation speed of the torque converter are read based on the signals from the respective sensors 72, 71 shown in FIG. A target line pressure value P corresponding to the opening degree and turbine rotation speed is read from a preset map.

Then, in step S14, the duty rate of the duty solenoid valve 65 corresponding to this target line pressure P is determined according to a subroutine described later, and in step S15, the duty ratio of the duty solenoid valve 65 to be output to the duty inlenoid valve 65 is determined. The frequency is set to, for example, 35 Hz, and in step S16, the ON time of the duty solenoid valve 65 during the 10N-OFF cycle is calculated based on the duty rate and the period which is the reciprocal of the drive frequency. Then, in step S17, a duty drive signal is output to the duty solenoid valve 65 so that the ON time obtained as described above is achieved. As a result, the pilot pressure supplied to the pressure booster boat 81a of the regulator valve 81 shown in FIG. 3 is adjusted, and the line pressure is controlled to the target line pressure P determined in step S13 above.

On the other hand, the line pressure control during the gear change in step S3 of FIG. 4 is performed as follows according to the subroutine shown in the flowchart of FIG.

In this subroutine, it is determined in step S21 whether or not the current shift is an upshift. If it is an upshift, the throttle opening is read in step S2□, and the throttle opening is changed in step S23. The target line pressure P is set based on the type of the current shift, that is, the combination of the gears before and after the shift. In other words, as shown in FIG. 7, the memory of the controller 70 has a map set in advance of the target line pressure P according to the throttle opening and the type of shift. The line pressure P is read.

This focuses on the fact that the required torque capacity of each frictional engagement element during upshifting varies not only depending on the throttle opening corresponding to the input torque, but also depending on which gear stage the shift is being made. This is to ensure that the optimum torque capacity is always obtained during upshifts between each gear stage. As shown in FIG. 8, this target line pressure P is set lower than the conventional line pressure P', and has different characteristics depending on the type of shift, but details are shown in step S5 in FIG. The line pressure correction control based on the learning of the shift time is set according to a subroutine to be described later, and in that case, the map shown in FIG. 7 is set differently depending on whether or not fuel cut control is executed during the shift. .

Further, if the current shift is a downshift, it is determined in step S24 whether or not it is a 3-2 downshift, and if it is not a 3-2 downshift, in step S25, as shown in FIG. Control similar to the control in step S2, that is, line pressure control during non-shifting according to the subroutine of FIG. 5 is executed.

On the other hand, if the current shift is a 3-2 downshift, the turbine rotation speed is read in step S26, and the baseline pressure Pa corresponding to this turbine rotation speed is determined in step S27 as shown in FIG. Read from a preset map as shown in . Here, the reason why such a baseline pressure Po is set only during the 3-2 downshift is that the release operation of the 3-4 clutch 43 and the 2-4 brake shown in FIGS. 1 and 3 are performed during this shift. 45 are performed at the same time, it is necessary to adjust the timing of these operations, and in particular, the optimal engagement timing of the 2-4 brake 45 varies depending on the turbine rotation speed, so the line pressure must also be adjusted to correspond to the turbine rotation speed. This is to change the situation.

Note that this baseline pressure P. Specifically, as line pressure correction control by learning the revving speed shown in step S6 in FIG. 4, it is set according to a subroutine to be described later.

During this 3-2 downshift, the rate of change in throttle opening is then calculated in step 321, and in step S29, a characteristic as shown in FIG. A correction coefficient 01 is obtained from a preset map, and this correction coefficient 01 is used as the baseline pressure P. The target line pressure P is set by multiplying by This is because when the throttle opening change rate is large, the speed at which the turbine rotation speed increases due to downshifting also becomes large, so the 2-4 brake 45 is adjusted accordingly.
This is to speed up the timing of the conclusion of the contract.

When the target line pressure P is set in step 821 or step S29 as described above, steps 314 to 314 of the flow chart of FIG. In the same manner as S17, the duty ratio is determined according to the target line pressure P, the frequency of the duty drive signal is set, the duty ON time is calculated, and the duty solenoid valve 65
By performing each operation of outputting a duty drive signal to the gear shifter, the line pressure during gear shifting is controlled to a value suitable for each gear shifting. During this shift, the frequency of the duty drive signal is set to, for example, 70 Hz in order to improve the responsiveness of line pressure control.

Next, step S14 in FIG. 5 and step S in FIG.
,. The subroutine for determining the duty ratio will be explained with reference to the flowchart shown in Fig. 1).

In this subroutine, in step S41, the target line pressure P set in the subroutine of FIG. 5 or 6 is read, and then in step S42, the automatic gear shift is Read the machine's hydraulic oil temperature. Then, in step S43, the pace duty rate Do corresponding to the oil temperature and line pressure at that time is determined from a map set in advance for a plurality of oil temperatures as shown in FIG. Here, pace duty rate D
The reason why oil temperature is used as a parameter for setting o is that the duty rate to obtain the target line pressure P varies depending on the oil temperature, and the actual oil temperature is When the temperature does not correspond to the temperature, the pace duty rate Do is determined by linear interpolation using these maps.
is required.

Next, in step S44, the elapsed time after the engine key is turned on is measured, and in step S45, a correction coefficient C2 corresponding to this elapsed time is read from a preset map as shown in FIG. That is, immediately after the engine or automatic transmission is activated, the characteristics of the control pressure with respect to the duty rate are different from those in normal times due to the presence of air in the hydraulic control circuit, so this is corrected.

Then, in step S46, a final duty rate is calculated by multiplying the pace duty rate Do by this correction coefficient 02, and this is used in the line pressure control shown in FIGS. 5 and 6.

In addition, the line pressure correction control by learning the revving rotation speed in step S6 of FIG. 4, that is, the correction control for the baseline pressure P0 during the 3-2 downshift set in step S27 of FIG. This is carried out by a subroutine whose flowchart is shown in FIG.

In this subroutine, the turbine rotation speed is read in step SSt, and then in step S52, the target turbine rotation speed No. after the shift is calculated from the turbine rotation speed before the shift. Next, in step S53, the rate of change in the turbine rotational speed is calculated, and in step S54, it is determined whether or not this rate of change has become equal to or less than a predetermined value, that is, the shift end timing for the change in the turbine rotational speed shown in FIG. Point X corresponding to
It is determined whether o has been reached. When this point Xo is reached, the turbine rotation speed N at that time is determined in step SSS.
is read, and in step 356, the deviation ΔN (=N-NO) of the turbine rotational speed N at the end of the shift with respect to the target turbine rotational speed N after the shift is calculated.

Next, in step S57, the correction coefficient 03 corresponding to the deviation ΔN is read from a preset map as shown in FIG. 16, and in step S5g, this correction coefficient 03 is read from step S2? in FIG. Baseline pressure P set at

The baseline pressure Po is corrected by multiplying by . In that case, the target rotation speed N of the turbine rotation speed N at the end of the shift. If the deviation ΔN is 0, the correction coefficient 03=
1, and when the deviation ΔN is positive (when the turbine rotation speed at the end of the shift is larger than the target rotation speed No, as shown in N1 in FIG. 15), the correction coefficient C3>1 is set, and the deviation ΔN
When N is negative (when the turbine rotational speed at the end of the shift is smaller than the target rotational speed No, as in N2 in FIG. 15), the correction coefficient C3 is set to <1. In other words, when the turbine speed or engine speed rises at the end of the 3-2 downshift, the baseline pressure PO is corrected to be higher, and when the speed is pulled down, the baseline pressure PO is increased. 3-4 clutch 43 during the 3-2 downshift.
The releasing operation and the engaging operation of the 2-4 brake 45 are performed at optimal timing.

On the other hand, in the line pressure correction control by learning the shift time performed in step ≦5 of the flowchart in FIG. 4 that constitutes the characteristic part of hydrogen, that is, in the line pressure control during the shift shown in FIG. 6, it is performed as step S23. Control for setting the target line pressure P during upshifting is performed as follows by a subroutine shown in a flowchart in FIG.

In this control, first, step S61. In s6□, the turbine rotation speed and throttle opening degree are read, and then in step 86N, the target turbine rotation speed after the shift is calculated from the turbine rotation speed before the shift, and in step S64, it is determined whether or not the shift operation has ended, i.e. It is determined whether or not the turbine rotational speed has decreased to the target rotational speed due to the upshift, and when the turbine rotational speed has decreased to the target rotational speed, it is determined that the speed change operation has been completed.

Next, in step S65, the shift time T from the start to the end of the shift operation is calculated, and in step S66, the average throttle and torque opening during the shift ((opening at the start of shift + opening at the end of shift) is calculated. The opening degree)/2) is calculated, and at the same time, in step S67, the corresponding base shift time To is read from a map preset according to the type of shift. and,
In step 868, the correction coefficient 04 corresponding to the average throttle opening is read from a preset map as shown in FIG. 18, and the optimum shift time T1 is calculated by multiplying the base shift time To by this correction coefficient 04. calculate. Here, al in FIG. 18 showing the above correction coefficient 04,
The numerical values such as a2, . . . are larger as they are larger than 1 and the subscript is smaller. Therefore, the optimum shift time T1 becomes longer when the throttle opening is low. This is to deal with the fact that the gear shift shock appears more pronounced when the throttle opening is low.

Then, in Stella 7S69, the actual shift time T and the optimum shift time T1 are compared, and the deviation ΔT (=T-71) is calculated, and in step 370, the amount of line pressure correction corresponding to the deviation ΔT is calculated. ΔP is read from a preset map as shown in FIG. In this case, the numerical values of s, , and b2 in FIG. 19 are larger than 0, and the larger the subscript, the larger the value. Therefore, when the deviation ΔT is positive (when the actual shift time T is longer than the optimum shift time T1), the line pressure correction amount ΔP is set to a larger value as the deviation ΔT is larger, and when the deviation ΔT is negative (the actual shift time T1 is longer than the optimum shift time T1), the line pressure correction amount ΔP is When the shift time T is shorter than the optimum shift time T, the smaller the deviation ΔT (the larger the absolute value), the larger the negative value.

By learning the shift time in this manner, the line pressure correction amount ΔP is set in accordance with the type of current shift and the average throttle opening during the shift. Next, this line pressure correction amount ΔP is corrected based on the oil temperature of the hydraulic oil of the automatic transmission. In other words, in step S71, the oil temperature is read, and then in step S72, a correction coefficient corresponding to the oil temperature is corrected. 05 to 20th
The correction amount ΔP is corrected by reading from a map preset with the characteristics shown in the figure and multiplying the line pressure correction amount ΔP obtained in step S70 by this correction coefficient 05. In that case, as shown in Figure 20,
The above correction coefficient 05 is below a predetermined oil temperature (for example, 60°C),
It is set so that it becomes smaller than 1 and approaches O as the oil temperature decreases. Therefore, the effect of correcting the line pressure by learning the shift time is weakened when the oil temperature is low.

Further, in step S3, the shift mode selected by the mode selection switch 74 shown in FIG. 2 is read, and in step S74, it is determined whether fuel cut control has been performed during the shift. This fuel cut control during gear shifting is
This is to reduce shift shock by stopping the fuel supply to the engine during gear shifting and reducing engine output. In this example, when the throttle opening is 4/8 or more, which increases engine output It is executed when changing gears, and fuel cut is not performed when the engine water temperature is low or when the battery voltage is low.

Then, the line pressure correction amount ΔP obtained in step S72 is further corrected according to the shift mode, the presence or absence of fuel cut control during shift, and the average throttle opening during shift, and this is used to correct the line pressure correction amount ΔP. Set pressure P.

The final line pressure P is set by steps 375 to S77 if the fuel cut control is not being executed, or by steps 378 to SaO if the fuel cut control is being executed.
The process is carried out in the same manner as above, and first, the corresponding map is selected from a preset map as shown in FIG. Correction coefficient 06 or 06
'('' indicates the value at the time of execution of fuel cut control. The same applies hereinafter) is read (step 5751378). In that case, the 21st.
CP in the map shown in Figure 22! + CF2. ”'.

cE1+ CE2+ ”'+ CP!' + CP2'
+ ”' + Numerical values such as CElo・Ci3°, etc. are less than 1 and the smaller the subscript, the smaller the value.Also, for the same average throttle opening, the value in economy mode is It is smaller than the value in power mode (for example, C21<cpt), and the value when fuel cut control is executed is smaller than the value when it is not executed (for example, CE
4'<Cl3). This is to reduce the line pressure correction amount when the throttle opening is low, when in economy mode, and when fuel cut control is executed. Note that the map shown in FIG. 22 when the fuel cut control is executed is set only for throttle openings of 4/8 or more where the control is performed.

Next, the baseline pressure Po or Poo at the time of shifting according to the throttle opening is set in advance as shown in Fig. 23.24 for each of the execution and non-execution of fuel cut control. Read from map (Step S)
. , 579), this baseline pressure P, or Poo;
Using the correction coefficient C6 or C6' and the line pressure correction amount ΔP, calculate the final line pressure P by using one of the following formulas:
Output (Step S)7. S8°).

When fuel cut control is not executed: p = po + Δpxc6 When fuel cut control is executed: p = po' + ΔPxC 6° With the above, the final line pressure P or Po is set by learning the shift time, and this determines the type of shift and the current shift time. The new target line pressure corresponding to the throttle opening is registered in the corresponding area of the memory storing the map shown in FIG. Then, during subsequent gear changes, the corresponding target line pressure is read from this map, and the aforementioned control is performed to achieve this line pressure.
In that case, the target line pressure registered in the map shown in FIG. 7 is set depending on whether fuel cut control is executed or not during gear shifting, and the target line set when the control is executed is:
The target line pressure that is used during a subsequent shift when the control is executed, and which is set when the control is not executed, is used during a subsequent shift when the control is not executed.

In the above embodiment, the target line pressure as the final line pressure is memorized and updated, but it is also possible to memorize and update the line pressure correction amount and add it to the baseline pressure to obtain the target line pressure. It may be configured as follows, and the same effect can be obtained in this case as well.

(Effects of the Invention) As described above, according to the present invention, line pressure correction control is performed by learning the shift time during line pressure control during gear shifting, and the line pressure correction amount obtained by this control is Since the correction is made depending on whether the engine output torque reduction control is executed or not during the process, an appropriate correction will be made in either case. As a result, in this type of automatic transmission, it is possible to increase the number of learning corrections while avoiding erroneous learning, and line pressure control during gear shifting can be performed more accurately, resulting in appropriate shifting times. In addition, a shift with less shift shock can be realized.

[Brief explanation of drawings]

The drawings show an embodiment of the present invention; Fig. 1 is a schematic diagram showing the mechanical configuration of an automatic transmission, Fig. 2 is a control system diagram, Fig. 3 is a hydraulic control circuit diagram, and Fig. 4 is a line diagram. Flowcharts showing the main routine of pressure control, Figures 5 and 6 are flowcharts showing subroutines of line pressure control during non-shifting and shifting, respectively, and Figures 7 to 10 are maps used in the subroutine of Figure 6. Figure 1) is a flowchart showing a subroutine for determining the duty rate; Figures 12 (a), (b), (c) and 13 are illustrations of maps used in this subroutine; The figure is a flowchart showing a subroutine for line pressure correction control by learning the speed of racing, Figure 15 is an explanatory diagram of the operation of this subroutine, Figure 16 is an explanatory diagram of the map used in this subroutine, and Figure 17 is a shift diagram. A flowchart showing a subroutine for line pressure correction control based on time learning, and FIGS. 18 to 24 are explanatory diagrams of maps used in this subroutine. 10... Automatic transmission, 60... Hydraulic control circuit, 65
．． 81... Line pressure adjustment means (duty solenoid valve, regulator valve) 1, 70... Line pressure correction means, correction amount correction means (controller), 75...
- Torque reduction means (engine controller). s 1 z! 1E2 Fig. 4 Fig. 5 Fig. 6 Fig. 7 Fig. 8 Fig. Throttle/1 Torr angle Fig. 14 Fig. 15 Fig. 16 Turbini/rotation speed and relation (-N) 2Q ■ 21 ■ lI22 Fig. Shin 23 Figure 24

Claims (1)

[Claims] (1) Line pressure adjustment means provided in the hydraulic control circuit to adjust the line pressure supplied to the friction engagement element, and the actual shift time and target shift time when changing the line pressure adjusted by the line pressure adjustment means. A hydraulic control device for an automatic transmission, comprising: a line pressure correcting means for learning correction according to a deviation from the line pressure correcting means; and a torque reducing means for reducing engine output torque under predetermined conditions during gear shifting, the line pressure correcting means A correction amount correcting means is provided for correcting the learning correction amount set by the learning correction amount according to whether or not the speed change for which the correction was performed is a speed change for which the reduction control of the engine output torque is performed by the torque reduction means. A hydraulic control device for an automatic transmission characterized by: