JPH0932912A - Gear change control device for vehicular automatic transmission - Google Patents

Abstract

PROBLEM TO BE SOLVED: To prevent the occurrence of a gear change shock due to the change in an engine load at a gear change process by making a correction for nearing engagement pressure in the case of an occurrence of engine load change after from the inception of an engagement pressure change. SOLUTION: In the case of judgement about a change in engine load, judgement is made as to whether an engine load change or a change in throttle valve opening appears at a process for changing a gear to the second speed gear attainable with the engagement of a brake B3, or more accurately, at a process for an increase in the engagement pressure of the brake B3 (i.e., at sweep control process). A hydraulic pressure correction means has a means for detecting a time elapsed from raising control for increasing the engagement pressure of the brake B3, or from the inception of sweep control. Engagement pressure in the case of an engine load (throttle valve opening) after a change from the inception of the sweep control is calculated on the basis of the elapsed time. Then, the engagement pressure of the brake B3 is corrected so as to agree to the calculated engagement pressure. According to this construction, the engagement pressure of the brake B3 becomes an optimum value, thereby preventing the occurrence of a gear change shock.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a shift control device for an automatic transmission for a vehicle.

[0002]

2. Description of the Related Art There is known an automatic transmission of a type that achieves a desired gear stage among a plurality of gear stages by combining operations of a plurality of hydraulic friction engagement devices. In such automatic transmissions, an accumulator is often used to moderate the increase or decrease of the hydraulic pressure in the hydraulic friction engagement device in order to reduce the torque change during gear shifting. The accumulator controls changes in hydraulic pressure in the hydraulic friction engagement device in accordance with accum back pressure applied to the back surface of the piston.

However, since the above accumulator requires a relatively large volume in order to fully exert its function, the hydraulic transmission is provided with a plurality of accumulators because the size of the automatic transmission becomes large. It has been proposed to directly control the internal hydraulic pressure without using an accumulator. For example, JP-A-6-3
That is the shift control device described in Japanese Patent No. 41525.

In the above shift control device, for example, an electromagnetic valve for outputting a control pressure and an engagement pressure adjusting device for adjusting the hydraulic pressure in the hydraulic friction engagement device based on the control pressure from the electromagnetic valve. A pressure valve is provided, and when the predetermined gear stage is achieved, the engagement pressure regulating valve is operated according to the control pressure output from the solenoid valve, so that the engagement pressure of the hydraulic friction engagement device does not use an accumulator. Is directly controlled by the electronic control unit.

[0005]

By the way, in the above-mentioned conventional shift control device, for example, when shifting is achieved by engaging a predetermined hydraulic friction engagement device, the throttle valve opening and the accelerator pedal operation amount are changed. The control oil pressure is output from the solenoid valve at a magnitude and an increase rate corresponding to the magnitude of the engine load, which gradually increases the oil pressure in the predetermined hydraulic friction engagement device. The frictional engagement device is smoothly engaged at a speed according to the engine load. However, if the engine load is changed in the middle of the gear shift, then only the control for increasing the hydraulic pressure in the predetermined hydraulic friction engagement device is started with a magnitude and a change rate according to the changed engine load. Then, compared with the case where the engine load changed from the beginning, the engagement pressure was insufficient,
In some cases, the engagement pressure of the predetermined hydraulic friction engagement device does not reach an optimum value, and a shift shock may occur when the predetermined hydraulic friction engagement device is engaged.

The present invention has been made in view of the above circumstances. An object of the present invention is to change the engagement pressure of a hydraulic friction engagement device at a rate of change according to the engine load. An object of the present invention is to provide a shift control device for an automatic transmission for a vehicle, which suitably prevents a shift shock caused by a change in engine load during a shift.

[0007]

SUMMARY OF THE INVENTION To achieve the above object, the gist of the present invention is to have a plurality of hydraulic friction engagement devices that are engaged or disengaged to achieve a plurality of gear stages. In an automatic transmission for a vehicle, a hydraulic pressure in a predetermined hydraulic friction engagement device that is engaged or disengaged to achieve a predetermined gear position among a plurality of gear positions is output from a solenoid valve. Achievement of a predetermined gear stage by directly controlling the hydraulic pressure in the predetermined hydraulic friction engagement device using the engagement pressure regulating valve that regulates pressure based on the pressure and the engagement pressure regulating valve. At this time, a shift control device comprising: a shift hydraulic pressure control means for changing the engagement pressure of the predetermined hydraulic friction engagement device at a rate of change according to the engine load of the vehicle; Way of Changing Engaging Pressure of Hydraulic Friction Engagement Device An engine load change determining means for determining whether or not there is a change in the engine load, and (b) the engine load change determining means for changing the engagement pressure of the predetermined hydraulic friction engagement device while the engine is being changed. If it is determined that the load has changed, the engagement pressure in the predetermined hydraulic friction engagement device controlled by the shift hydraulic pressure control means is changed from the beginning of the change of the engagement pressure to the value after the change. Hydraulic pressure correction means for correcting the engagement pressure to approach the engagement pressure when the engine load is.

[0008]

According to the present invention, when the engine load change determining means determines that the engine load has changed during the change in the engagement pressure of the predetermined hydraulic friction engagement device, the oil pressure correcting means. Thus, when the engagement pressure in the predetermined hydraulic friction engagement device controlled by the shift hydraulic pressure control means is the engine load after the change from the beginning of the change of the engagement pressure, the engagement pressure Is corrected to approach. Therefore, according to the present invention, for example, at the time of gear shift achieved by engaging a predetermined hydraulic friction engagement device, even if the engine load is changed during the gear shift, Since the engagement pressure is corrected so as to approach the value when the engine load after the change is reached from the beginning of the control period for changing the engagement pressure, the predetermined hydraulic friction engagement device The engagement pressure has an optimum value, and shift shock is preferably prevented.

[0009]

According to another aspect of the present invention, preferably, when the engine load change determining means determines that the engine load has changed, the hydraulic pressure correcting means determines the predetermined hydraulic friction engagement. The engagement pressure in the device is corrected so as to be the engagement pressure when the engine load after the change from the beginning of the change of the engagement pressure is the engine pressure.

Further, preferably, the shift hydraulic pressure control means adds a value obtained by adding an initial value that increases according to the engine load and an increase value that increases at a change rate that increases according to the engine load, The hydraulic pressure correction means outputs the pressure as an engagement pressure during a rising period during which the engagement pressure of the predetermined hydraulic friction engagement device is increased during the shift period. Equipped with an elapsed time detecting means for detecting an elapsed time from the start time of the rising control for increasing the engagement pressure, and based on the elapsed time, the relationship when the engine load after the change is from the beginning of the rising control. The resultant pressure is calculated, and the engagement pressure of the hydraulic friction engagement device is made to match the calculated engagement pressure.

[0011]

DESCRIPTION OF THE PREFERRED EMBODIMENTS One embodiment of the present invention will be described below in detail with reference to the drawings.

FIG. 1 is a skeleton view showing an example of an automatic transmission for a vehicle, the speed of which is controlled by a shift control device according to one embodiment of the present invention. In the figure, an output of an engine 10 is input to an automatic transmission 14 via a torque converter 12,
The power is transmitted to drive wheels via a differential gear device and an axle (not shown).

The above-mentioned torque converter 12 is used for the engine 1
Pump impeller 18 connected to crankshaft 16
, A turbine runner 22 connected to the input shaft 20 of the automatic transmission 14, a lock-up clutch 24 directly connecting between the pump impeller 18 and the turbine runner 22, and a one-way clutch 26 that prevents rotation in one direction. And a stator 28 that is in the position.

The automatic transmission 14 includes a first transmission 30 for switching between high and low gears, and a second transmission 32 for switching between a reverse gear and four forward gears. The first transmission 30 includes an HL planetary gear device 34 including a sun gear S0, a ring gear R0, and a planet gear P0 rotatably supported by the carrier K0 and meshed with the sun gear S0 and the ring gear R0, a sun gear S0, and a carrier. The clutch C0 and the one-way clutch F0 are provided between the sun gear S0 and the housing 41, and the brake B0 is provided between the sun gear S0 and the housing 41.

The second transmission 32 has a first planetary gear unit 36 comprising a sun gear S1, a ring gear R1, and a planet gear P1 rotatably supported by the carrier K1 and meshed with the sun gear S1 and the ring gear R1.
, Sun gear S2, ring gear R2, and carrier K
2 and a second planetary gear unit 38, which is rotatably supported by the sun gear S2 and a ring gear R2 and is meshed with the sun gear S2 and the ring gear R2, and the sun gear S3, the ring gear R3, and the carrier K3, which are rotatably supported by the sun gear S2. S3 and a third planetary gear set 40 including a planet gear P3 meshed with the ring gear R3.

The sun gear S1 and the sun gear S2 are integrally connected to each other, the ring gear R1, the carrier K2 and the carrier K3 are integrally connected, and the carrier K3 is connected to the output shaft 42. Further, a ring gear R2 is integrally connected to the sun gear S3. A clutch C1 is provided between the ring gear R2 and the sun gear S3 and the intermediate shaft 44, and the sun gear S1 and the sun gear S3 are provided.
A clutch C2 is provided between the shaft 2 and the intermediate shaft 44. A band-type brake B1 for stopping rotation of the sun gear S1 and the sun gear S2 is provided on the housing 41.
It is provided in. A one-way clutch F1 is provided between the housing 41 and the sun gear S1 and the sun gear S2.
And the brake B2 are provided in series. The one-way clutch F1 is configured to be engaged when the sun gear S1 and the sun gear S2 try to reversely rotate in the direction opposite to the input shaft 20.

A brake B3 is provided between the carrier K1 and the housing 41, and a brake B4 and a one-way clutch F2 are provided between the ring gear R3 and the housing 41 in parallel. This one-way clutch F
2 is configured to be engaged when the ring gear R3 attempts to rotate in the reverse direction.

In the automatic transmission 14 configured as described above, for example, according to the operation table shown in FIG. 2, one of the reverse gears and the forward gears of which the gear ratios are sequentially different are switched to one of the gear stages. In FIG. 2, a circle indicates an engaged state, a blank indicates a released state, and a ● indicates an engaged state during engine braking. As is clear from FIG. 2, the brake B3 is adapted to be engaged when shifting from the first gear to the second gear.

As shown in FIG. 3, the engine 10 of the vehicle is
The intake pipe is operated by an accelerator pedal 50
First throttle valve 52 and throttle actuator 54
And a second throttle valve 56 operated by
ing. Also, the rotation speed N of the engine 10_EDetect
Engine rotation speed sensor 58, intake air of engine 10
Intake air amount sensor 60 for detecting the amount Q / N,
Temperature T_AIntake air temperature sensor 62 for detecting the
The opening θ of the throttle valve 52_THThrottle sensor to detect
Rotation speed N of the output shaft 42_OUTThat is, the vehicle speed V
Vehicle speed sensor 66 for detecting, cooling water temperature T of engine 10
_WCooling water temperature sensor 68 to detect the
Operation of brake switch 70 and shift lever 72
Position P_SHOperating position sensor 74 for detecting
Rotational speed N of the clutch C0 _C0Clutch to detect
C0 rotation sensor 73, hydraulic oil temperature T of hydraulic control circuit 84
_{OI L}An oil temperature sensor 75 for detecting
From those sensors, the engine speed N_E, Intake air
Quantity Q / N, intake air temperature T_A, Opening of the first throttle valve
θ_TH, Vehicle speed V, engine cooling water temperature T_W, Brake operation
State BK, shift lever 72 operating position P_SH,clutch
Rotation speed N of C0_C0, Hydraulic oil temperature T_OILIs a signal
Electronic control unit for engine 76 or electronic control unit for shifting 7
It is designed to be supplied to 8.

Further, as shown in FIG. 4, the shift lever 72 is operated to the P range, the R range, the N range, the D range, the 4 range, the 3 range, the 2 range, and the L range located in the front-rear direction of the vehicle. At the same time, the support mechanism is configured such that the range between the D range and the 4 range and the range between the 2 range and the L range are operated in the left-right direction of the vehicle.

The engine electronic control unit 76 shown in FIG.
A so-called microcomputer including a PU, a RAM, a ROM, and an input / output interface, and a CPU is a RAM
The input signal is processed according to a program stored in advance in the ROM while utilizing the temporary storage function of, and various engine controls are executed. For example, the fuel injection valve 80 is controlled to control the fuel injection amount, the igniter 82 is controlled to control the ignition timing, a bypass valve (not shown) is controlled to control the idle speed, and the throttle actuator is controlled to control the traction. 2nd throttle valve 56 by 54
The fuel injection valve 8 is controlled when the engine speed N _E enters into a preset overspeed region (for example, red zone).
0 is cut off to prevent further increase in engine speed N _E. The engine electronic control unit 76 is connected to a shift electronic control unit 78 so that they can communicate with each other.
A signal necessary for one is transmitted from the other as appropriate.

The shift electronic control unit 78 is also a microcomputer similar to that described above, and the CPU processes input signals in accordance with a program stored in the ROM in advance while utilizing the temporary storage function of the RAM. Of each solenoid valve or linear solenoid valve. For example,
The electronic shift control device 78 includes a linear solenoid valve SLT for generating a throttle pressure P _TH having a magnitude corresponding to the opening θ _TH of the first throttle valve 52, and a linear solenoid valve SLT for controlling the accumulator back pressure. SLN, lockup clutch 24 engagement, release, slip amount, brake B3
Drive the linear solenoid valve SLU to control the direct control of the clutch and the clutch-to-clutch shift, respectively. Further, the electronic shift control device 78 determines the actual throttle valve opening θ _TH and the vehicle speed V from the previously stored shift diagram.
The gear stage of the automatic transmission 14 is determined based on the above, and the solenoid valve S is used to obtain the determined gear stage and engagement state.
1, S2, S3 are driven, and the solenoid valve S4 is driven when engine braking is generated.

5 and 6 show the essential parts of the hydraulic control circuit 84. In FIGS. 5 and 6, the 1-2 shift valve 88 and the 2-3 shift valve 90 are provided with an electromagnetic valve S1,
A switching valve that is switched based on the output pressure of S2 at the time of shifting from the first speed gear to the second speed and at the time of shifting from the second speed to the third speed.
The numerical value indicating the switching position indicates the gear position. The forward range pressure P _D is determined by the shift lever 72 when the forward range (D,
4,3,2, a pressure that is generated from the manual valve (not shown) when being operated to L), pressurized to be higher according to the throttle valve opening theta _TH by a not-shown line pressure regulating valve regulating It has been a source pressure of the line pressure P _L that.

When a shift output for switching from the first gear to the second gear is output, the forward range pressure P is set.
_D is supplied to the brake B3 via a 1-2 shift valve 88, a 2-3 shift valve 90, an oil passage L01, a B3 control valve 92, and an oil passage L02. Incidentally, 94 is a damper for performing buffer against sudden supply of the line pressure P _L. Further, when the shift output for switching from the second gear to the third gear is output, the forward range pressure P _D becomes the 2-3 shift valve 9
At the same time as being supplied to the brakes B2 and B2 accumulators 100 through 0 and the oil passage L03, the working oil in the brake B3 is returned to the oil passages L02, B3 control valve 92, oil passage L01, 2-3 shift valve 90, and returned. Oil passage L04, 2-
3 Pressure valve is drained through the timing valve 98,
Branch oil passages L05 and B2 branched from the return oil passage L04
It is adapted to be rapidly drained through the orifice control valve 96.

Back pressure chamber 1 of B2 accumulator 100
00 of _B, accumulator back pressure P _{AC C} from accumulator back pressure control valve (not shown) for generating an accumulator back pressure P _ACC on the basis of the output pressure P _SLN output pressure P _SLT and the linear solenoid valve SLN for linear solenoid valve SLT Is supplied at each shift.

The B3 control valve 92 is connected to the oil passage L0.
Spool valve element 104 that opens and closes between valve 1 and oil passage L02
And the same as the spool valve element 104 with the spring 106 interposed.
Provided at the center and having a diameter larger than that of the spool valve element 104.
The plunger 108 and the spring 106 are housed therein,
When the 2-3 shift valve 90 is switched to the third speed side
Forward range pressure P output from it_DVia oil passage L07
Oil chamber 110 and shaft end of plunger 108
The output pressure P of the linear solenoid valve SLU _SLUTo
And a receiving oil chamber 112. Therefore, B3
In the process of establishing the second gear, the control valve 92
Output pressure P of linear solenoid valve SLU_SLUAccording to spoo
Position the valve valve 104 in the open position shown on the left side of the center line.
Perform the first fill early on and then
Supply the hydraulic oil from the oil passage L01 to the oil passage L02, or
Causes the hydraulic oil in the oil passage L02 to flow out to the discharge oil passage L06.
As a result, the engagement pressure P in the brake B3_B3The rise of
From Equation 1, the output pressure P_SLUBased on accumulator
Pressure directly, as in the buffering effect of Said lini
The solenoid valve SLU has an output pressure P_SLUIs electronic for shifting
Supplied from control device 78 to linear solenoid valve SLU
According to command value DSLU (drive duty ratio: unit is%)
Equation 1 shows that
As shown, the engagement pressure P in the brake B3_B3And Liniyasore
Output pressure P of the solenoid valve SLU_SLUAre proportional to each other
Therefore, the command value DSLU and the brake B3
Combined pressure P_B3Corresponds uniquely. Equation 1
And S₁ And S_Two Is plunger 108 and spoo
This is a cross-sectional area of the valve element 104.

[0027]

[ _Equation 1] P _B3 = P _SLU · S ₁ / S ₂

The B2 orifice control valve 96 is
Rake B2 and B2 accumulator 100 and oil passage L0
3 and at the same time discharge oil passage L06 and drain
A spool valve 114 that opens and closes between the
Attach the pool valve 114 toward the first drain position.
The spring 116 to be urged and the shaft end of the spool valve 114
And the output pressure P of the third solenoid valve S3_S33-4 shifts
And an oil chamber 120 which receives the oil through a valve 118.
You. As a result, the third solenoid valve S3 can be used, for example, during a 3 → 2 shift.
Is turned on and its output pressure P_S3Supplied to oil chamber 120
The spool valve 114 prevents the brake
B2 and between B2 accumulator 100 and oil passage L03
Open between them brake B2 and B2 accumulate
Fur that quickly discharges hydraulic oil from the
Stodrain operation is performed. Also in 1 → 2 shift
Indicates that the third solenoid valve S3 is turned off and its output pressure
P _S3Is supplied to the oil chamber 120 so that the B3 control
Operation discharged from the pressure regulating operation of the roll valve 92
A discharge oil passage L06 for discharging oil and a drain port 113
Is opened to adjust the pressure of the B3 control valve 92
Is permitted, but when the 1 → 2 shift is completed, the third solenoid valve S
3 is turned on, the discharge oil passage L06 and the drain port 1
13 is closed by closing B3 control valve
The pressure regulation operation of 92 is stopped.

The 2-3 timing valve 98 is involved in shifting from the second gear to the third gear and regulates the release pressure from the brake B3 from the linear solenoid valve SLU according to the output pressure P _SLU. Function as. That is, 2-
3 timing valve 98, 2 → 3 forward range pressure P _D is output from the 2-3 shift valve 90 when the shift is outputted 3-4 shift valve 118 and the solenoid relay valve 122
The supply port 124, the drain port 126, and the oil passage L04, which are supplied through the brake B3, are connected to the supply port 124 or the drain port 126.
Spool valve 128 that regulates the pressure P _B3 during the drain period of
And a first concentrically provided with a spool valve element 128 via a spring 130 and having the same diameter as the spool valve element 128.
The plunger 132 and the spool valve element 128 are provided concentrically with the spool valve element 128 so as to be in contact with one end thereof.
A second plunger 134 having a diameter larger than 28, houses a spring 130, an oil passage L08 to the forward range pressure P _D is output therefrom when the 2-3 shift valve 90 is switched to the second speed side And an oil chamber 138 provided at the shaft end of the first plunger 132 for receiving the output pressure P _SLU from the linear solenoid valve SLU.
And an oil chamber 140 provided at the shaft end of the second plunger 134 for receiving the hydraulic pressure P _B2 in the brake B2 and a feedback oil chamber 142 for receiving the feedback pressure.

Therefore, the sectional area of the spool valve element 128 and the first plunger 132 is S ₃ , and the spool valve element 12 is
8, the land area on the second plunger 134 side is S ₄ ,
When the cross-sectional area of the second plunger 134 and S _5, 2 → 3
Brake B in the release process when the gearshift output is output
The pressure P _B3 of No. 3 decreases according to the increase of the engagement pressure P _B2 of the brake B2 from the mathematical expression 2 by the pressure adjustment operation by the 2-3 timing valve 98, and the output pressure P of the linear solenoid valve SLU.
_The pressure is adjusted to increase according to the _SLU .

[0031]

## EQU2 ## P _B3 = P _{SLU .S 3} / (S ₃ −S ₄ ) −P _{B2.
S 5 / (S 3 -S 4} )

Further, when the forward range pressure P _D output from the 2-3 shift valve 90 switched to the second speed side is supplied to the oil chamber 136, the 2-3 timing valve 98 outputs the spool valve. The child 128 is adapted to be locked. This is also the oil chamber 138 of the 2-3 timing valve 98 and B
Since the oil chamber 112 of the third control valve 92 is connected, the change in the volume of the oil chamber 138 of the 2-3 timing valve 98 is prevented in the first speed and the second speed, and the B3 control valve 92 is closed. This is so as not to affect the pressure regulation operation.

[0033] C0 exhaust valve 150, the output pressure P _S4 of the third but brought into the closed position in accordance with the hydraulic pressure in the output pressure P _S3 and the oil passage L01 of the solenoid valve S3, the fourth solenoid valve S4
Accordance with the spool 152 is caused to position to the open position, the line pressure P _L supplied via it when the 4-5 shift valve (not shown) is in switching state of the following fourth speed, the second speed And clutch C at times other than the fifth speed
0 and the C0 accumulator 154.

In the shift control device having the above structure
Then, the 1 → 2 shift judgment is performed and the second speed gear stage is achieved.
1-2 shift valve when gear change output for
88 is switched from the first speed side to the second speed side. This
As a result, the forward range pressure P_D1-2 shift valves 88, 2
-3 shift valve 90, oil passage L01, B3 control valve 9
It is supplied to the brake B3 via 2. 1 → 2 like this
During the shift period of the shift, the B3 control valve 92 is used.
Engagement pressure P in brake B3_B3To directly control
To quickly fill the brake B3 with hydraulic oil.
Step control, engagement pressure P in brake B3_B3A little
Standby control and brake to keep the total torque generated
Smoothly increase the engagement torque of B3 regardless of the engine load.
Clutch rotation speed N to add_C0Is detected
Until the engaging pressure P in the brake B3 _B3The engine at that time
Load or throttle valve opening θ_THAccording to the size
And sweep control to increase at the rate of change, brake B3
Clutch speed N until full engagement is detected_C0Drop of
Engagement in the brake B3 so that the speed reaches a predetermined target value
Pressure P_B3Feedback control to control
You. The B3 control valve 92 is a linear solenoid.
Output pressure P of valve SLU_SLUIe its linear solenoid valve
According to command value (driving duty ratio) DSLU to SLU
Controlled.

FIG. 7 is a functional block diagram for explaining the main part of the control function of the electronic shift control device 78. In the figure, the automatic transmission 14 is a hydraulic friction engagement device (brake B3) that is engaged to achieve the second speed or is disengaged to achieve the first speed.
When the engagement pressure regulating valve and (B3 control valve 92) is provided which applies directly regulates the hydraulic P _B3 in the brake B3.

The automatic shift control means 158 determines the actual throttle valve from the prestored basic shift diagram consisting of a plurality of upshift shift lines and downshift shift lines corresponding to the type of shift between each gear. Opening θ _TH and vehicle speed V
The shift determination of the automatic transmission 14 is executed based on the above, and the output of the drive signal, that is, the shift output for driving the solenoid valves S1, S2, S3 so as to achieve the gear stage of the shift destination determined by the shift is executed. To do. The basic shift diagram is well known in a control device for an automatic transmission.

The shift hydraulic pressure control means 160 controls the B3 control valve 92 so that, for example, during the shift related to the engagement of the brake B3, the shift B3 is controlled.
A command value for controlling the control valve 92 is output to the linear solenoid valve SLU, and the engagement pressure P _{B3 of the} brake _B3 is controlled to control the step control, the standby control, the sweep control,
The feedback control is executed sequentially. Further, the shift hydraulic pressure control means 160 is configured to increase the hydraulic pressure P _B3 in the brake B3 in order to smoothly advance the engagement of the brake B3 regardless of the engine load, that is, in the rising control period, that is, the sweep control period, the engine load for example the step value DST (θ _TH) is a function that increases in response to the throttle valve opening theta _TH, increased value increases at the throttle valve opening theta _TH increases in response to the increasing rate R (θ _TH) Value obtained by adding R (θ _TH ) · t _TH DSLU [= DST
(Θ _TH ) + R (θ _TH ) · t _TH ] is output as the command value DSLU of the engagement pressure of the brake B3. Note that t _TH is the elapsed time from the start point of the engagement pressure increase control (sweep control) for a predetermined throttle valve opening θ _TH .

The engine load change determining means 162 is in the middle of shifting to the second gear, which is achieved by the engagement of the brake B3, more precisely, during the change of the engaging pressure P _B3 of the brake B3 (in the middle of the sweep control). ), It is determined whether the engine load has changed, that is, the throttle valve opening θ _TH has changed. When the engine load change determination means 162 determines that the engine load has changed during the increase change of the engagement pressure P _B3 of the brake B3, the oil pressure correction means 164 changes the hydraulic pressure control means 160.
By the engagement pressure P _B3 of the brake B3 is controlled by the engine load starts from the beginning after the change of the increase control period or sweep control period increases the hydraulic pressure P _B3 in the brake B3 (the throttle valve opening theta _TH2) as approaching the engaging pressure in case of a, as more preferably its engagement pressure, the engagement pressure P _B3 i.e. the engagement pressure P _B3 of the brake B3, which is controlled by the shift hydraulic pressure control means 160 The command value DSLU to be controlled is corrected.

The oil pressure correction means 164 is preferably
Elevation control for increasing the engagement pressure P _{B3 of the} brake B3, that is, an elapsed time detecting means 166 for detecting an elapsed time t _EL from the start point of the sweep control is provided, and based on the elapsed time, from the beginning of the sweep control, the above-mentioned The engagement pressure when the engine load after change (throttle valve opening θ _TH2 ) is calculated, and the engagement pressure P _{B3 of the} brake B3 is corrected so as to match the engagement pressure.

FIG. 8 is a flow chart for explaining the main part of the control operation by the electronic shift control device 78, 1 → 2.
9 shows a shift control routine. The flow chart corresponding to the automatic shift control means 158 is well known and therefore omitted.

In FIG. 8, at SA1, it is judged whether the automatic shift control means 158 has performed the 1 → 2 shift output. If the determination of SA1 is negative, this routine is terminated, but if the determination is affirmative, the engagement pressure P _{B3 of the} brake B3 is increased according to a predetermined procedure to engage the second gear. Execution of SA2 and below for starting the execution is started. Time point t ₁ in FIG. 9 shows this state.

At SA2, the rapid supply control is executed to maintain the command value DSLU at the preset step command value DSLU _A for the preset period T _A. The step command value DSLU _A and the period T _A are
This is experimentally obtained in advance in order to improve the responsiveness of the brake B3 by quickly supplying and filling the hydraulic oil. As a result, the B3 control valve 92
The hydraulic oil is quickly supplied to the brake B3 through the. This state is shown at time t _{2 in} FIG. 9.

At SA3, the standby pressure control is executed to maintain the command value DSLU at the preset standby command value DSLU _B for the preset period T _B. The standby command value DSLU _B and the period T _B are experimentally obtained in advance so that a slight engagement torque is generated in the brake B3. As a result, the brake B3
The engagement pressure P _B3 therein is maintained at the pressure corresponding to the standby command value DSLU _B for the period T _B. T ₃ time points in FIG. 9 shows this state. The periods T _A and T _B are time-controlled by a timer (not shown) on software.

When the standby pressure control is completed as described above, in SA4 corresponding to the elapsed time detecting means 166, the time counting operation of the timer T _S12 is started. The timer T _{S1 2} is for counting the elapsed time from the beginning of the end That sweep control subsequent control standby pressure.

Next, the engine load change judging means 1
At SA5 corresponding to 62, throttle valve opening θ _TH
It is determined whether or not the change rate dθ _TH / dt of is greater than or equal to a preset determination change rate θ ₁ . The determination reference change rate θ ₁ is experimentally obtained in advance in order to determine the change in the engine load to the extent that it is necessary to switch the sweep control to the one when the throttle is changed. SA above
If the determination of 5 is denied, the normal sweep control shown in FIG. 10, for example, is executed in SA6. This sweep control is for gradually increasing the engagement torque of the brake B3 according to the engine load. That is, by the above-mentioned sweep control, when the engine load is small, the shift shock is relatively conspicuous and therefore the engagement torque is gradually increased, while when the engine load is large, the shift shock is relatively unnoticeable. The engagement torque is immediately increased.

After the throttle valve opening θ _TH corresponding to the actual engine load is read in SA6-1 in FIG. 10, in SA6-2, the throttle valve opening θ _TH is calculated from the pre-stored relationship shown in FIG. 11, for example. step value DST (θ _TH) is calculated based on the theta _TH. Also, continue S
At A6-3, the increase rate R is calculated based on the throttle valve opening θ _TH from the relationship stored in advance as shown in FIG. 12, for example.
(Θ _TH ) is calculated. Next, at SA6-4, with the step value DST (θ _TH ) and the increase rate R (θ _TH ),
The command value DSLU for the linear solenoid valve SLU is sequentially calculated from Equation 3 based on the elapsed time t _TH from the start of the sweep control for the predetermined throttle valve opening θ _TH , and is output.

[0047]

[Formula 3] DSLU = DST (θ _TH ) + R (θ _TH ) ・ t _TH

After the normal sweep control is executed as described above, at SA8 in FIG. 8, it is determined whether or not the decrease of the clutch rotation speed N _C0 has started, for example, a turning point of the clutch rotation speed N _C0 or It is determined using a well-known algorithm for determining the upper peak. When the determination of SA8 is denied, the SA5 and subsequent steps are repeatedly executed, but when the determination of SA8 is affirmed, the rotational speed change of the rotary member such as the input shaft 20 is started by the increase of the engagement torque of the brake B3. Since it is in the open state, the feedback control of SA9 is executed. In FIG. 9, t ₄ indicates the starting point of this feedback control.

In the feedback control of SA9, the target reduction speed of the preset clutch rotational speed N _C0 is compared with the actual reduction speed, and the linear solenoid valve SLU for the linear solenoid valve SLU is eliminated so as to eliminate the difference between them. The command value DSLU is sequentially adjusted according to the magnitude of the control deviation and the changing speed of the control deviation.

Next, at SA10, it is determined whether or not the brake B3 is completely engaged, for example, the automatic transmission 14
Input shaft rotation speed N _IN, that is, N _C0 and output shaft rotation speed N
_The determination is made based on whether or not the actual gear ratio γ (= N _IN / N _OUT ) which is the ratio with _OUT matches the gear ratio γ ₂ of the second gear. When the determination of SA10 is negative, the above SA9 and subsequent steps are repeatedly executed, but when the determination is affirmative, the command value DSLU is the maximum value DSL in SA11.
U _max, and the engagement pressure P _{B3 of the} brake B3 is set to the maximum value, the forward range pressure P _D (= line pressure P _L ).
The time t _{5 in} FIG. 9 indicates this point. As above, S
A2, SA3, SA6, SA9, and SA11 correspond to the shift hydraulic pressure control means 160 because they control the engagement pressure P _B3 of the brake B3 in the 1 → 2 shift in a predetermined procedure.

When the determination at SA5 is affirmative, that is, when the change in engine load is determined, at SA7 corresponding to the hydraulic pressure correction means 164, for example, as shown in FIG.
Sweep control is executed when the throttle is changed as indicated by the one-dot chain line 4 in FIG. In the sweep control at the time of changing the throttle, in order to generate the engagement pressure when the throttle valve opening θ _TH2 after the change is made from the beginning of the sweep control,
The command value DSLU ₂ after the change of the engine load is sequentially determined based on the elapsed time t _EL , and is output in place of the command value DSLU by the normal sweep control. That is, the command value DSLU by the normal sweep control is corrected to the command value DSLU ₂ .

That is, in SA7-1 in FIG. 13, after the changed engine load, that is, the throttle valve opening θ _TH2 is read, in SA7-2, after the changed from the prestored relationship shown in FIG. 11, for example. A step value DST (θ _TH2 ) is calculated based on the throttle valve opening θ _TH2 . Further, in the subsequent SA7-3, for example, the throttle valve opening θ _{TH2 is changed} from the previously stored relationship shown in FIG.
The increase rate R (θ _TH2 ) is calculated based on Then
In SA7-4, the command value DSLU ₂ for the linear solenoid valve SLU is calculated based on the step value DST (θ _{TH 2} ) and the increase rate R (θ _TH2 ) and the elapsed time t _EL from the start of the sweep control. It is sequentially calculated from 4 and is output.

[0053]

[Formula 4] DSLU ₂ = DST (θ _TH2 ) + R (θ _TH2 )
・ T _EL

FIG. 14 is an enlarged view showing before the sweep control period.
Then, although the command value DSLU is increased by the normal sweep control as shown by the solid line, for example, at time t _A , the throttle valve opening θ _TH is larger than θ _TH2.
When the engine load is increased due to the change to, the command value DSLU ₂ determined based on the equation 4 is the command value DSL of the normal sweep control.
Instead of U, it is output as shown by the alternate long and short dash line. The command value DSLU ₂ (one-dot chain line) after this engine load change is
It has the same magnitude and increase rate as the value (two-dot chain line) when the throttle valve opening θ _TH2 after the above-mentioned change has been set since the beginning of the sweep control.

Incidentally, in the past, a predetermined throttle valve opening
θ_THTime t from the start of the sweep control for
_THCommand value for the linear solenoid valve SLU based on
Since DSLU was sequentially calculated from Equation 3,
As shown by the broken line 14 in FIG._A
From the difference of throttle valve opening Δθ_U(= Θ_TH2−
θ _TH) Change in step value ΔDST_U[=
(Θ_TH2) -DST (θ_TH)] Only after being increased
Increase rate R (θ_TH2) In the sweep control
Throttle valve opening θ after the above change from the beginning_TH2Met
Value (two-dot chain line) for the case
Hydraulic pressure P in rake B3_B3Is not enough,
Of the guard timer set to the maximum value of
Because the B3 is completely engaged, the gear shift
There was a problem.

On the contrary, even if the engine load is changed in the decreasing direction by decreasing the throttle valve opening θ _TH to a smaller value θ _TH2 within the sweep control period, according to the present embodiment. Although the output as shown in dashed line field 14, conventionally, as shown in broken line in FIG. 14, the command value DSLU from time t _a, the throttle valve opening difference Δθ _D (= θ _TH2 Change in step value corresponding to −θ _TH ) ΔDST _D [= (θ _TH2 ) -DST (θ
_TH )], then increased at an increasing rate R (θ _{TH 2} ), and becomes larger than the value when the throttle valve opening was θ _TH2 from the beginning of the sweep control.
The engagement of the brake B3 was accelerated to cause a shift shock.

As described above, according to this embodiment, the second load is determined by the SA5 corresponding to the engine load change determining means 162.
Engagement pressure P _B3 of the brake B3 for achieving the high speed gear
If it is determined that the engine load has changed during the increase of the pressure, the engagement pressure P _{B3 of the} brake B3 controlled by the shift hydraulic pressure control means 160 is changed by SA7 corresponding to the hydraulic pressure correction means 164. engine load from the change beginning after the change in the P _B3 (throttle valve opening θ
_TH2 ) is corrected so that the engagement pressure becomes the same.
Therefore, even if the engine load is changed in the middle of the 1 → 2 shift achieved by the engagement of the brake B3, the sweep control in which the engagement pressure P _{B3 of the} brake B3 changes the engagement pressure P _B3. Since the throttle valve opening θ _TH2 after the above change has been corrected from the beginning of the period so that it matches the value, the engagement pressure P of the brake B3 is corrected.
_B3 becomes an optimal value, and shift shock is preferably prevented.

Although one embodiment of the present invention has been described in detail with reference to the drawings, the present invention can be implemented in other modes.

For example, when it is determined that the engine load has changed, the value when the engagement pressure P _{B3 of the} brake B3 is the engine load after the change of the engagement pressure P _B3 from the beginning (FIG. 14 was corrected to be the value of the two-dot chain line), but even if it is not matched with the value (the value of the two-dot chain line of FIG. 14), it is more than the conventional value shown by the broken line of FIG. If the value is close to the value (value of the chain double-dashed line in FIG. 14), a tentative effect is obtained. In short, when it is determined that the engine load has changed, the engagement pressure P _{B3 of the} brake _B3 becomes
It suffices that the correction is made so as to approach the value of the chain double-dashed line in FIG.

Further, in the above-mentioned embodiment, the 1 → 2 shift in which the brake B3 is engaged has been described, but the 3 → 2 shift may be applied. Conversely, a 2 → 1 shift or a 2 → 3 shift that releases the brake B3 may be used. In short, the oil pressure P _B3 in the brake B3 during the shifting
The present invention can be applied when the hydraulic pressure P _B3 is increased or decreased at a rate of change R (θ _TH ) corresponding to the throttle valve opening θ _TH during a period in which is changed.

Further, in the above-described embodiment, the throttle valve opening θ _TH is used as the engine load, but instead of this, the operation amount of the accelerator pedal 50, the negative pressure in the engine intake pipe, the output torque of the engine, etc. Can be used.

Further, in the above-mentioned embodiment, the gear shift achieved by the engagement of the brake B3 has been described, but it may be used for the gear shift achieved by the engagement of another friction engagement device.

In addition, the step value DST of the above-mentioned embodiment
(Θ _TH ) and the rate of increase R (θ _TH ) may be corrected based on the oil temperature T _OIL of the engine 10 or the transmission 14.

Further, in the flow charts of FIGS. 8, 10 and 13 described above, it does not matter if steps are added or the contents of the steps are changed within the range where the same control function is achieved.

Although not illustrated one by one, the present invention can be carried out in various modified and improved modes based on the knowledge of those skilled in the art.

[Brief description of drawings]

FIG. 1 is a skeleton diagram illustrating a configuration of an automatic transmission for a vehicle controlled by a shift control device according to an embodiment of the present invention.

FIG. 2 is a table showing a relationship between a combination of operations of a plurality of frictional engagement devices and a gear established by the combination in the automatic transmission of FIG. 1;

FIG. 3 is a block diagram including a hydraulic control circuit and an electric control circuit for controlling the automatic transmission of FIG. 1;

FIG. 4 is a diagram illustrating an operation position of a shift lever of FIG. 3;

5 is a diagram illustrating a main part of the hydraulic control circuit of FIG.

FIG. 6 is a diagram illustrating a main part of the hydraulic control circuit of FIG. 3;

FIG. 7 is a functional block diagram illustrating a main part of a control function of the electronic shift control device of FIG.

8 is a flowchart illustrating a main part of a control operation of the shift electronic control device of FIG. 3;

9 is a time chart for explaining a main part of control operation of the electronic shift control device of FIG.

FIG. 10 is a flowchart illustrating an essential part of the operation of the normal sweep control of FIG.

11 is a diagram illustrating a relationship used when obtaining a step value DST in FIG.

FIG. 12 is a diagram illustrating a relationship used when obtaining an increase rate R in FIG.

FIG. 13 is a flowchart illustrating an essential part of an operation of sweep control when the throttle is changed in FIG.

14 is a time chart for explaining a main part of a control operation of the electronic shift control device of FIG. 3, in which a solid line shows a normal sweep control operation, and a one-dot chain line shows a sweep control operation when the throttle is changed. And the broken line indicates the conventional case.

[Explanation of symbols]

 14: Automatic transmission 92: B3 control valve (engagement pressure regulating valve) 160: Shift hydraulic pressure control means 162: Engine load change determination means 164: Hydraulic pressure correction means 166: Elapsed time detection means

Claims (1)

[Claims] 1. An automatic transmission for a vehicle having a plurality of hydraulic friction engagement devices that are engaged or disengaged to achieve a plurality of gear stages, wherein a predetermined gear stage among the plurality of gear stages is provided. An engagement pressure regulating valve that regulates the hydraulic pressure in a predetermined hydraulic friction engagement device that is engaged or disengaged to achieve the pressure based on the control pressure output from the solenoid valve; and the engagement pressure regulating valve. By directly controlling the hydraulic pressure in the predetermined hydraulic friction engagement device by using the above, when the predetermined gear stage is achieved, the engagement pressure of the predetermined hydraulic friction engagement device is set to A shift control device comprising: a shift hydraulic pressure control means for changing at a change rate according to an engine load, wherein whether or not the engine load has changed during a change in an engagement pressure of the predetermined hydraulic friction engagement device. Engine load change judgment And the engine load change determining means determines that the engine load has changed during the change of the engagement pressure of the predetermined hydraulic friction engagement device, the shift hydraulic pressure control means controls the shift. And a hydraulic pressure correction means for correcting the engagement pressure in the predetermined hydraulic frictional engagement device so as to approach the engagement pressure when the engine load is after the change from the beginning of the change of the engagement pressure. A shift control device for an automatic transmission for a vehicle, comprising: