JP2884906B2 - Shift control method for automatic transmission for vehicle - Google Patents

Abstract

PURPOSE:To take measures not to sense any incompatibility to speed change feeling by restraining the overshoot of an engine speed below an allowable maximum engine speed stably, even in the case of the generation of endurance inferiority of a friction engagement element and the start up delay of a supply oil pressure and the like. CONSTITUTION:An elapsed time Tst from a basic point before the start point of an actual speed change upto the actual speed change start point is measured and the measured elapsed time is compared with a prescribed discriminated value. When the elapsed time is bigger than the prescribed discriminated value, the increase rate of an oil pressure supplying to a second friction engagement element is increased so as to wait the actual speed change start. Accordingly, the generation of the transmission torque of a connection side clutch or the start up of the supply oil pressure and also the reach to the speed change start point of an inertia phase become quick and then the excess revolution of an engine is prevented. An excess revolution watching routine is operated as the backup of an usual speed change control and an abnormal speed change control by which an overshoot amount becomes big can be eliminated and the engine maximum speed in a speed change time can be restrained below an allowable maximum engine speed stably.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a shift control method for an automatic transmission for a vehicle, and more particularly to a shift control method for preventing an engine speed from rising during an upshift.

[0002]

2. Description of the Related Art An automatic transmission for a vehicle is provided with a number of frictional engagement elements such as clutches and brakes which are operated by hydraulic pressure.
For example, when performing an upshift, the friction engagement element of the established lower gear is disengaged while the friction engagement element of the higher gear to be established is engaged. Thus, the switching of the friction engagement element is performed.
When such engagement of the friction engagement elements is switched, if the engagement of the engagement side frictional engagement element is too early, a shift shock occurs, while the engagement of the release side frictional engagement element is completely released. If the engagement of the friction engagement element on the side does not start, the engine speed may increase during the gear shift, and may exceed the allowable maximum speed. In view of this, the transmission torque of each of the frictional engagement elements on the disengagement side and the coupling side is optimized to prevent the above-described overshoot of the engine speed and the occurrence of the shift shock during the shift.

In order to frictionally engage the frictional engagement element on the coupling side, the frictional engagement element must be disengaged from an idle stroke section from a standby position where the engagement is completely released to a standby position to a position immediately before transmission torque is generated. Then, the hydraulic pressure supplied to the coupling-side frictional engagement element is gradually increased at a predetermined increasing rate to start the engagement of the coupling-side frictional engagement element. Then, wait for the input shaft rotation speed of the automatic transmission to reach the actual shift start point, and when the input shaft rotation speed reaches the actual shift start point, feedback-control the transmission torque of the coupling-side friction engagement element, The input shaft rotation speed is reduced toward the synchronous rotation speed so that the input shaft rotation speed change rate matches the target change rate.

[0004]

By the way, in the conventional shift control method, the rise of the hydraulic pressure supplied to the friction engagement element at the time of the power-on upshift and the variation of the friction coefficient of the friction plate of the friction engagement element are caused. As shown by the broken line in FIG. 1, the engagement start delay of the engagement-side frictional engagement element may occur. In such a case, if the engagement of the disengagement side frictional engagement element is completely released, the engine speed may exceed the allowable maximum speed (upper limit) determined by the design specifications, and may overshoot. In addition, the engine speed temporarily jumps into a so-called red zone, and the overshoot amount of the engine speed varies.

The present invention has been made in order to solve such a problem, and even when the durability of the friction engagement element is degraded or the rising of the supply oil pressure is delayed, the engine speed is stably reduced. It is an object of the present invention to provide a shift control method for an automatic transmission for a vehicle in which an overshoot is suppressed to a maximum allowable rotational speed or less and a feeling of strangeness does not occur in a shift feeling.

[0006]

According to the present invention, a first friction engagement element and a high-speed side, each of which is driven by a hydraulic pressure and establishes a low-speed side shift stage, are achieved according to the present invention. A second frictional engagement element for establishing a gear position,
During the upshift, the first frictional engagement element is disengaged, and the second frictional engagement element is stroked for a predetermined invalid stroke section and then supplied to the second frictional engagement element. A shift control method for an automatic transmission for a vehicle, in which a hydraulic pressure is increased at a predetermined increase rate, an actual shift is started, a feedback shift control is started after the actual shift is started, and the input shaft rotational speed is reduced toward a synchronous rotational speed. In, measuring the elapsed time from the reference time point before the actual shift start point to the actual shift start point, comparing the measured elapsed time with a predetermined determination value,
A shift control method for an automatic transmission for a vehicle, characterized in that when the elapsed time is greater than a predetermined determination value, the increase rate of the hydraulic pressure supplied to the second frictional engagement element is increased to wait for the actual shift start. Provided.

[0007]

In the shift control method of the present invention, referring to FIG. 1, the actual shift start point Ns from a reference time point, for example, a shift command time point or a hydraulic pressure start point of the second (coupling side) friction engagement element.
The elapsed time t until is measured. This actual shift start point Ns
Is the point at which the transmission torque is switched to the coupling-side frictional engagement element, the shaft torque is restored, and the shift is actually started, and the input shaft rotation speed starts to decrease. More specifically, the so-called torque phase shift starts at the same time as the engagement by the coupling-side frictional engagement element is started, the transmission torque of the coupling-side frictional engagement element is increased, and the input shaft rotation speed is reduced to the above-mentioned actual speed. Starting point N
When s is reached, the so-called inertia phase shift starts.
In the shift of the inertia phase, it means a state in which the transmission torque of the coupling-side friction engagement element is to be feedback-controlled.For example, the input shaft rotation speed is synchronized so that the input shaft rotation speed change rate matches the target change rate. Decrease towards rotational speed.

Thus, by comparing the elapsed time measured as described above with a predetermined determination value, it is possible to determine whether or not the input shaft rotational speed exceeds the allowable maximum rotational speed and overshoots. When the time is greater than the predetermined discrimination value, the increase rate of the hydraulic pressure supplied to the coupling-side frictional engagement element is increased to speed up the increase of the transmission torque of the coupling-side frictional engagement element, thereby reaching the actual shift start point. Time required

[0009]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS An embodiment of the present invention will be described below in detail with reference to the accompanying drawings. FIG. 2 shows a schematic configuration of an automatic transmission of an automobile that implements the shift control method according to the present invention. Reference numeral 1 in the figure denotes an internal combustion engine, and the output of the engine 1 is transmitted to driving wheels (not shown) via an automatic transmission 2. The automatic transmission 2 includes a torque converter 4, a gear transmission 3, a hydraulic circuit 5, a controller 40, and the like. The gear transmission 3 is, for example, a forward four-stage reverse 1
It is provided with a friction engagement element of a hydraulic clutch or a hydraulic brake for performing a shift operation by switching a shift speed and a shift position. The hydraulic circuit 5 has a duty solenoid valve (hereinafter, simply referred to as a solenoid valve) corresponding to each of the friction engagement elements described above.
Operate each clutch and brake independently of each other. Each solenoid valve is electrically connected to an output side of a controller (ECU) 40, which will be described later, and adjusts a working oil pressure supplied to the friction engagement element according to a drive signal from the ECU 40.

FIG. 3 is a partial structural view of the gear transmission 3, in which a first drive gear 31 and a second drive gear 32 are rotatably arranged around an input shaft 3a. The input shaft 3a between the first drive gear 31 and the second drive gear 32 has
Hydraulic clutches 33 and 34 are fixedly provided as shift friction engagement elements. The drive gears 31 and 32 rotate integrally with the input shaft 3a by engaging with the clutches 33 and 34, respectively.

An intermediate transmission shaft 35 arranged in parallel with the input shaft 3a is connected to a drive axle via a final reduction gear unit (not shown). A first driven gear 36 and a second driven gear 37 are fixed to the intermediate transmission shaft 35, and the driven gears 36 and 37 are engaged with the driving gears 31 and 32, respectively. Therefore, when the clutch 33 and the first drive gear 31 are engaged, the rotation of the input shaft 3 a is controlled by the clutch 33, the first drive gear 31, the first driven gear 36, and the intermediate transmission shaft 35. And the first gear (for example, the second speed) is achieved.
When the clutch 34 and the second drive gear 32 are engaged, the rotation of the input shaft 3a is controlled by the clutch 34, the second
Drive gear 32, second driven gear 37, intermediate transmission shaft 3
5 and the second gear (for example, the third speed) is achieved.

From the state in which the second speed side clutch 33 is engaged, the third speed
By engaging the high speed clutch 34, the automatic transmission 2
Shifts up from the second speed to the third speed. Conversely, by disengaging the clutch 34 from the state in which the clutch 34 is engaged and engaging the clutch 33,
The automatic transmission 2 shifts down from the third speed to the second speed.

The clutches 33 and 34 described above are, for example,
A hydraulic multi-plate clutch is used, and FIG.
3 shows an example. The clutch 33 includes a number of friction engagement plates 5.
When the hydraulic oil is supplied into the clutch 33 from the oil passage 14 through the port 51, the piston 52 moves forward to frictionally engage the friction engagement plates 50. On the other hand, when the piston 52 moves backward while being pressed by the return spring 53 and discharging the hydraulic oil to the oil passage 14 through the port 51, the frictional engagement between the friction engagement plates 50 is released. The clutch 34 is configured similarly to the clutch 33.

The ECU 40 has built-in storage devices such as a ROM and a RAM, a central processing unit, an input / output device, and a counter (all not shown). On the input side of the ECU 40, various sensors, for example, an input shaft of the gear transmission 3, that is, a turbine speed sensor (Nt sensor) 21 for detecting a speed Nt of a turbine of the torque converter 4, a transfer drive (not shown) No sensor 2 that detects the number of rotations of the gear (the number of rotations of the output shaft of the gear transmission 3) No
2. Arranged in the middle of an intake passage (not shown) of the engine 1,
As a parameter value representing the engine load, a throttle valve opening sensor (θt
Sensor) 23, Ne for detecting the rotational speed Ne of the engine 1
The sensor 24 and the like are electrically connected. These sensors 21 to 24 supply detection signals to the ECU 40. Note that the ECU 40 can calculate the vehicle speed based on the rotation speed No of the transfer drive gear detected by the No sensor 22.

Next, the shift control performed by the ECU 40 during an upshift will be described by taking, as an example, the case of shifting up from the first gear to the second gear. FIG.
1 shows a schematic procedure of upshift shift control that is executed when an upshift from the first gear to the second gear is determined.
First, in step S10, the ECU 40 sets the duty ratio of the solenoid valve on the release side to 0 until the transmission torque of the clutch 33, which has established the first speed, becomes substantially 0.
, The hydraulic pressure supplied to the clutch 33 is released, and the duty ratio of the solenoid valve of the clutch 34 for establishing the second speed is set to 100%, and the engagement of the clutch 34 is started from its standby position. The stroke of the invalid stroke section up to the position immediately before the stroke is performed. That is, the transmission torque of the release clutch 33 and the engagement clutch 34 is switched.

When the coupling side clutch 34 completes the stroke in the invalid stroke section (when the so-called rattling operation is completed), the ECU 40 proceeds to step S12, and sets the duty ratio Da of the coupling side solenoid valve 34 to a predetermined value Dao. Then, the solenoid valve 34 is driven at the set duty ratio Da. This predetermined value Dao is, for example,
4 is set experimentally or learned at the time of each shift control to a value that holds the stroke position 4 immediately before the start of the above-described coupling.

Next, the ECU 40 detects the turbine rotation speed Nt detected by the Nt sensor 21 (step S1).
4) It is determined whether or not the detected turbine rotational speed Nt has become equal to or lower than a value lower than the first-speed synchronous rotational speed by a predetermined rotational speed ΔN1 (for example, 20 rpm), that is, whether or not an out-of-synchronization has been detected (step). S16). Immediately after the above-described backlash operation is completed, the out-of-synchronization is not normally detected, so the determination result in step S16 is negative (No).
And the process proceeds to step S18.

In step S18, the duty ratio Da of the coupling-side solenoid valve is set to a value obtained by adding a predetermined increment ΔDao to the previous value Da, and the coupling-side solenoid valve is driven with the value. Then, returning to step S14 again, the duty ratio Da of the coupling-side solenoid valve is increased at a predetermined increasing rate while repeating the detection of the out-of-synchronization. Accordingly, the hydraulic pressure supplied to the coupling-side clutch 34 is also increased at the predetermined increasing rate α. (See FIG. 1), and waits for detection of loss of synchronization. The value of the predetermined increment ΔDao is set by an overspeed monitoring routine described later.

When the hydraulic pressure supplied to the coupling side clutch 34 is gradually increased and the out-of-synchronization is detected, the determination result of step S16 becomes affirmative (Yes), and the process proceeds to step S20 to start the feedback shift control. The feedback shift control method is not particularly limited, but various known methods can be applied. For example, for example,
The supply hydraulic pressure of the coupling-side clutch 34, that is, the transmission torque is feedback-controlled so that the time rate of change of the turbine rotation speed Nt matches the target change rate, and the turbine rotation speed Nt is controlled.
t is decreased toward the second-gear synchronous speed.

Next, in step S22, it is determined whether or not the turbine rotational speed Nt has substantially reached the second-gear synchronous rotational speed. More specifically, the absolute value of the deviation between the turbine rotational speed Nt and the second-gear synchronous rotational speed is determined as follows. It is determined whether or not it has become equal to or less than a predetermined determination value ΔN2 (for example, 50 rpm). If the determination result is negative, step S20 is repeatedly executed to continue the feedback shift control.

If the second speed synchronization is detected and the determination result of step S20 is affirmative, the process proceeds to step S24, where the hydraulic pressure of the disengagement side clutch 33 is completely released while the solenoid valve of the engagement side clutch 34 is operated. Duty ratio Da is 10
At 0%, the clutch 34 is completely engaged. Thus, the upshift from the first gear to the second gear is completed.

FIG. 6 shows an overspeed monitoring routine in which the above-mentioned increment ΔDao is corrected as follows. The ECU 40 executes this routine each time the upshift shift control is executed. First, in step S30, it is determined whether or not the supply of the hydraulic pressure to the coupling side clutch 34 has been started. The start point of the supply of the hydraulic pressure to the clutch 34 is a reference for measuring the elapsed time described later. After the shift command of the upshift is output and the invalid time until the hydraulic pressure is actually supplied to the clutch 34 after the shift command of the upshift is output, the
It is considered that the supply of the hydraulic pressure has been started.

In step S30, if the supply of the hydraulic pressure to the coupling side clutch 34 has not been started, step S30 is repeatedly executed to wait until the supply of the hydraulic pressure is started.
When the supply of the hydraulic pressure to the clutch 34 is started, the process proceeds to step S32, in which a timer is started to start measuring the elapsed time from the start of the hydraulic pressure supply to the coupling side clutch 34, which is the reference time.

Then, the program proceeds to a step S34, wherein it is determined whether or not the actual shift start point Ns is detected. The Ns point means the shift start point of the inertia phase, but actually, the ECU
Reference numeral 40 indicates a time when the previous value of the input shaft rotation speed Nt is compared with the current value, and when the current value is lower than the previous value, that is, when the input shaft rotation speed Nt starts decreasing for the first time after the shift command. When detected, it is regarded as the Ns point and detected. If the determination result of step S34 is negative, Ns
Wait until a point is found.

In step S34, when the Ns point is detected, the elapsed time Tst is read from the timer (step S36), and the read elapsed time Tst is compared with a predetermined reference time XST (step S38). The timer is reset after reading the elapsed time Tst. The reference time XST is an allowable upper limit time from the above-mentioned reference time point to the shift start point of the inertia phase, and the time required to reach the shift start point of the inertia phase is the allowable upper limit time XST.
If it is larger, it means that the engine rotation speed Ne may exceed the allowable maximum rotation speed (upper limit value) and the overshoot amount may become large as shown by the broken line in FIG. Therefore, if the result of the determination in step S38 is negative, the routine is terminated without doing anything. However, if the result of the determination is positive, the small value Δd is added to the above-mentioned increment ΔDao, and this is added to the new increment. Stored as ΔDao (step S4
0).

As described above, each time one shift control is executed, the elapsed time Tst required from the reference time to the shift start point of the inertia phase is monitored, and the increment ΔDao is corrected according to the elapsed time Tst. The corrected increment ΔDao increases the rate of increase in the hydraulic pressure supplied to the coupling side clutch 34 after the backlash operation from that at the time of the previous shift control.

In the above-described embodiment, the reference time at which the timer is started is set as the hydraulic pressure supply start time of the coupling side clutch. Alternatively, the time at which the shift command is output may be set as the reference time. Further, the point at which the input shaft rotation speed starts to decrease is detected as the actual shift start point Ns, but the point at which the engine speed Ne starts to decrease (FIG. 1)
(Medium, Ne1, Ne2 points) may be detected and used as the Ns point.

The hydraulic circuit according to the above-described embodiment is provided with a solenoid valve for each of the friction engagement elements (clutch, brake) on the coupling side and the release side, and is supplied to each friction engagement element by the corresponding solenoid valve. Operating hydraulic pressure was controlled. The transmission control method of the present invention is not limited to the automatic transmission having the hydraulic circuit having this configuration. FIG.
Shows a second embodiment of the hydraulic circuit to which the method of the present invention is applied, and this hydraulic circuit supplies the operating hydraulic pressure supplied to the coupling-side and release-side friction engagement elements 28 and 30 to one solenoid valve 6.
7 illustrates an example of control by the control unit 7. The construction of a hydraulic circuit for controlling the operating oil pressure of the coupling-side and release-side friction engagement elements by one solenoid valve is disclosed in Japanese Patent Application Laid-Open No. 58-46.
No. 258, Japanese Patent Application Laid-Open No. 60-215143 and the like have already been known.

The operation of the kick-down brake 30 for establishing the second gear is controlled by a kick-down servo 80 as a reciprocating hydraulic actuator. The kick-down servo 80 includes a housing for defining a stepped cylinder hole 80c. A stepped piston 80e slidably fitted in a stepped cylinder hole 80c;
e, a piston rod extending out of the housing,
That is, the actuator rod 80f is provided, and the tip of the actuator rod 80f is
The kick down brake 30, that is, the brake shoe wound on the peripheral surface of the kick down drum 52 can be abutted and engaged. The first and second pressure chambers 80a, 8a are provided in the stepped cylinder hole 80c.
The first pressure chamber 80a is defined between the step surface of the piston 80e and the step surface of the stepped cylinder hole 80c, as is apparent from FIG.

The first pressure chamber 80 a of the kick down servo 80 is connected to the 2-3 shift valve 5 via an oil passage 65.
5 is connected, and a shift control valve 69 is connected to the 2-3 shift valve 55 via an oil passage 62. An oil passage 83 is branched from the middle of the oil passage 65, and the oil passage 83 is connected to a 1-2 shift valve 84. The 1-2 shift valve 84 is further connected to bifurcated oil passages 85 and 86, and one of the two oil passages 85 is provided with a kick down servo 80.
The other oil passage 86 is connected to the front clutch 28 that establishes the first gear. In FIG. 7, the front clutch 28 is shown only schematically.

Here, the 2-3 shift valve 55 and the 1-2 shift valve 84 are spool-type on-off valves which are opened and closed by the pressure supplied to their operation control ports 87 and 88. The pressure to 87 and 88 is specifically led from a switching valve (not shown). For example, in the third gear, the 2-3 shift valve 5
Unlike the case shown in FIG. 7, the spool 55a of No. 5 does not receive the switching pressure through its operation control port 87 and is in a state of being displaced to the left end. Therefore, in this case, the oil passage 65 is connected to the oil discharge port EX of the 2-3 shift valve 55.
To the kick-down servo 80
Is connected to the low pressure side.
As a result, the piston 80e of the kick down servo 80
Is returned to the right in FIG. 7 by the spring force of the compression coil spring 80d in the second pressure chamber 80c, and the engagement of the kickdown brake 30 with the kickdown drum 52 is released. At this time, the switching pressure is not supplied to the 1-2 shift valve 84 through the operation control port 88, and therefore, the spool 84a is connected to the spool 84a as shown in FIG.
In the middle, it is displaced to the left end as shown. Therefore, in this case, the oil passage 86 leading to the front clutch 28
-2 is connected to the low pressure side through the oil discharge port 90 of the shift valve 84, whereby the front clutch 2
8 has been released. In this case, the oil passage 85,
Since 86 is always in communication, the second pressure chamber 80c of the kick down servo 80 is also connected to the low pressure side.

In the second speed, the 2-3 shift valve 55 is switched to the illustrated switching position.
The 1-2 shift valve 84 has also been switched to the illustrated position. Therefore, in this case, when the pressurized liquid is supplied to the first pressure chamber 80a of the kick down servo 80 through the oil passages 62 and 65, the piston 80e, that is, the actuator rod 80f moves leftward to kick. The down brake 30 is engaged, while the hydraulic fluid in the front clutch 28 can be discharged through the oil passage 86 and the oil discharge port 90, whereby the engagement of the front clutch 28 is released. .

Further, in the first speed, the 2-3 shift valve 55 remains at the switching position shown in the figure, whereas the 1-2 shift valve 84 has the spool in the right direction. The oil passage 83 and the oil passages 85 and 86 are communicated via the 1-2 shift valve 84, and the oil discharge port 90 is closed. In this case, through the 2-3 shift valve 55, the oil passage 83
Is supplied to the front clutch 28 through the 1-2 shift valve 84 and through the oil passage 86, whereby the front clutch 28 is brought into the engaged state. On the other hand, in the kick down servo 80, since the oil passages 86 and 85 are always in communication,
The pressure fluid supplied to the front clutch 28 is also supplied to the second pressure chamber 80b, and at the same time, the same pressure fluid is supplied to the first pressure chamber 80a through the oil passage 65. . In this case, since the piston 80e of the kick down servo 80 is a stepped piston as described above, the piston 80e is formed from the difference in pressure receiving area at both ends thereof.
Displaced rightward with the actuator rod 80f,
As a result, the engagement of the kickdown brake 30 is released.

Further, when the shift speed is shifted from the first speed to the second speed, each of the 2-3 shift valve 55 and the 1-2 shift valve 84 is in the illustrated switching position.
Regarding the kick down servo 80, the first pressure chamber 8
0a is supplied with the pressurized liquid, so that the piston 80e,
In other words, the actuator rod 80f is displaced in the direction for engaging the kickdown brake 30, while the oil passage 86, the 1-2 shift valve 8
The engagement of the front clutch 28 is released by releasing the pressurized liquid through the hydraulic fluid 4 and the oil drain port 90. At this time, the engagement of the front clutch 28 is released by the piston 80e of the kick down servo 80 as described later. Is displaced, it is controlled by the back pressure generated in the second pressure chamber 80b.

For this reason, in order to generate an appropriate back pressure in the second pressure chamber 80b, a predetermined throttle 91 is provided at the oil discharge port 90 of the 1-2 shift valve 84.
A predetermined throttle 9 is also provided in the oil passage 84 guided to the -2 shift valve 84.
2 are provided. Then, the above-described shift control valve 69
7, an oil passage 61 is connected to the left end thereof, and an oil passage 63 is connected thereto. The oil passage 61 is connected to an oil pump, and in the middle thereof, a solenoid valve 67 that opens and closes the oil passage 61 and controls the pressure of the hydraulic fluid supplied through the shift control valve 69.
Is inserted. The solenoid valve 67 is electrically connected to an electronic control unit (ECU) 40, and the electronic control unit 40 controls the switching operation of the solenoid valve 67 by duty control. Also, in the oil passage 63,
An operating oil pressure adjusted to a predetermined pressure is supplied from the above-described oil pump. The hydraulic fluid in the oil passage 61 is discharged to the low pressure side through a solenoid valve 67 that opens and closes according to the duty ratio. Therefore, the hydraulic pressure according to the duty ratio acts on the left end surface of the spool 69a of the shift control valve 69. Will do. As a result, the shift control valve 69 regulates the oil pressure from the oil passage 63 to generate a predetermined oil pressure PKD in the oil passage 62.

Then, the ECU 40 is supplied with detection signals of the operating state of the vehicle from the same various sensors as shown in FIG. 2, executes the same shift control as described above, and executes the aforementioned 1-shift control.
Two upshifts are performed. In the case of the second mode, when the supply of the operating oil pressure to the kick-down servo 80 on the coupling side is started, the duty ratio of the solenoid valve 67 is gradually changed, and the pressing force of the kick-down brake 30 is reduced to a predetermined value. Increase at the rate of increase and wait for the loss of synchronization. Then
When the out-of-synchronization is detected, the turbine rotation speed Nt is monitored, and the duty ratio of the solenoid valve 67 is feedback-controlled so that the change rate matches the target speed. In the second embodiment, the increasing rate of the pressing force of the kick-down brake 30 is increased in accordance with the elapsed time Tst detected in the overspeed monitoring routine to wait for loss of synchronization (actual shift start). Become.

[0037]

As described above, according to the method of the present invention, the elapsed time from the reference time point before the actual shift start point to the actual shift start point is measured, and the measured elapsed time is compared with the predetermined determination value. However, when the elapsed time is greater than the predetermined determination value, the increase rate of the hydraulic pressure supplied to the second frictional engagement element (engagement side clutch) is increased to wait for the start of the actual shift. The generation of the transmission torque or the rise of the supply hydraulic pressure is accelerated, and the reaching of the shift start point in the inertia phase is also accelerated, thereby preventing the engine from overrunning. The overspeed monitoring routine of the present invention operates as a backup for normal shift control. For example, it operates effectively when the friction coefficient of the clutch or brake decreases due to deterioration in durability, and an abnormal amount of overshoot increases. The shift control can be eliminated, and the maximum engine speed at the time of gear shifting can be stably suppressed to the allowable maximum speed or less. Also, according to the present invention,
The shift control for increasing the increase rate of the supplied oil pressure is control until the start of the inertia phase, so that it does not affect the inertia phase. Therefore, even when this control is operated, there is no unnatural feeling of the shift feeling. Further, since the above-described correction of the increase rate of the hydraulic pressure is performed during one shift control, it is possible to instantaneously respond to disturbance during the shift.

[Brief description of the drawings]

FIG. 1 is a graph showing a relationship among time changes of an input shaft rotation speed, an engagement side clutch supply pressure, an engine rotation speed, and a shaft torque of an automatic transmission that is controlled by the method of the present invention.

FIG. 2 is a schematic configuration diagram of an automatic transmission for a vehicle to which a method according to the present invention is applied;

FIG. 3 is a schematic configuration diagram showing a part of a gear train of the gear transmission 3 of FIG. 2;

FIG. 4 is a sectional view showing the hydraulic clutch of FIG. 2;

FIG. 5 is a flowchart showing a shift control procedure during an upshift according to the method of the present invention.

FIG. 6 is a flowchart of an overspeed monitoring routine.

FIG. 7 is a hydraulic circuit diagram showing a second mode of the hydraulic circuit of the automatic transmission to which the method of the present invention is applied.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 1 Engine 2 Automatic transmission 3 Gear transmission 5 Hydraulic circuit 22 Transfer drive gear rotation speed sensor 23 Throttle valve opening sensor 24 Engine rotation speed Ne sensor 28 Friction engagement element (clutch) 30 Friction engagement element (brake) 33 Friction Engagement element (clutch) 34 friction engagement element (clutch) 40 controller

Claims (1)

(57) [Claims] A first frictional engagement element for establishing a low-speed gear position and a second frictional engagement element for establishing a high-speed gear position, each of which is driven by hydraulic pressure; While disengaging the first frictional engagement element, the second
After the first frictional engagement element is stroked for the predetermined invalid stroke section, the hydraulic pressure supplied to the second frictional engagement element is increased at a predetermined increase rate, and the start of the actual shift is waited. A shift control method for an automatic transmission for a vehicle, in which a shift control is started to reduce an input shaft rotation speed toward a synchronous rotation speed, wherein an elapsed time from a reference time point before the actual shift start point to an actual shift start point is calculated. The measured elapsed time is compared with a predetermined determination value, and when the elapsed time is greater than the predetermined determination value, the increase rate of the hydraulic pressure supplied to the second friction engagement element is increased to start the actual shift. A shift control method for an automatic transmission for a vehicle, comprising waiting.