JPH06331016A - Gear change control device for automatic transmission - Google Patents

Abstract

PURPOSE:To carry out down shift operation of clutch to clutch without causing gear change shock with a small time lug by maintaining oil pressure in the abrasion engaging device of an engaging side to a prescribed value in low pressure before input rotary speed of a transmission exceeds the synchronous point of a gear change step, and then boosting oil pressure thereof. CONSTITUTION:In the automatic transmission 2 of forward 5 steps and backward 1 step provided with a torque converter 111, a plural gear change part 112 and a main gear change part 113, transmission between a second speed step and a third speed step is set as clutch to clutch transmission of brakes B2 and B3, and each clutch and brake are engaged and disengaged through an oil pressure control device 20 by command of a computer 30. In this case, at the time of down shift of clutch to clutch, engaging pressure of the abrasion engaging device of an engaging side is maintained to a prescribed low pressure value after input rotary speed of a transmission exceeds the synchronous point of a gear change step, so that opening pressure in the abrasion engaging device of an opening side is maintained to a high pressure value. When input rotary speed of the transmission exceed the synchronous point of the gear change step, opening pressure is reduced while raising engaging pressure.

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 which may perform a downshift of a clutch to clutch relating to two friction engagement devices.

[0002]

2. Description of the Related Art When performing a specific shift of an automatic transmission, it is often necessary to simultaneously engage and disengage two friction engagement devices (including a clutch and a brake in a broad sense). So-called clutch to clutch shift).
In this case, if the engagement and disengagement of the friction engagement devices are not properly synchronized, the output shaft torque will drop or the engine will blow up.

For this reason, conventionally, in order to perform such control, a one-way clutch having a function substantially equivalent to that of one friction engagement device is generally provided so that such a problem does not occur. Was considered.

However, in the method of synchronizing the friction engagement devices by using the one-way clutch in this way, the cost increases due to the attachment of the one-way clutch.
In addition, there are problems such as an increase in weight and an occupied space.

In view of such a point, in recent years, against the background of improvements in various sensor technologies and in electronic control technologies for hydraulic control devices, "clutch-to-clutch shift" can be directly executed without using a one-way clutch. Attempts to do so have become active again.

Conventionally, as a technique for controlling the downshift of the clutch-to-clutch, for example, Japanese Patent Publication No. 52-1834.
No. 4 publication discloses that an orifice kick-down valve is used, a governor pressure (which changes according to the vehicle speed) and a spring load are applied to the spool of the orifice kick-down valve so as to face each other. A technique has been proposed in which the supply is performed via a large-diameter orifice, while the supply is performed via a small-diameter orifice at a predetermined vehicle speed or higher. According to this, the frictional engagement device on the engagement side can be engaged to some extent near the point where the input and output rotations are synchronized.

[0007]

However, in this prior art, the supply of the engagement pressure to the friction engagement device on the engagement side depends on the drain opening condition of the throttle opening and the release pressure of the release side friction engagement device. However, since the downshift is executed by only two-step control by switching the orifice depending on the vehicle speed, there is a large deviation from the actual synchronization point, which may cause engine injection or excessive tie-up. was there. That is, since only the governor pressure (vehicle speed) is taken into consideration, it is not possible to properly cope with variations in the actual synchronization of input / output rotation depending on other running conditions. There was a problem.

In view of such circumstances, the applicant first
In Japanese Patent Application No. 3-344123 (unknown), in the downshift of the clutch-to-clutch, the friction adjustment on the engagement side is performed before the input speed of the automatic transmission reaches the synchronization point of the low speed stage by the pressure adjusting means. A technique has been proposed in which the engagement pressure of the device is made to stand by at a predetermined low pressure value, and after it is detected that the synchronization point is reached, the engagement pressure is gradually increased based on the input torque. According to this technique, the engagement pressure can be adjusted from before the synchronization point according to the input rotation, and after the synchronization, the engagement pressure can be increased based on the input torque.
The engagement timing of the engagement-side friction engagement device can be optimized, and the amount of torque transmitted from the release-side friction engagement device to the engagement-side friction engagement device can be optimized.

However, this Japanese Patent Application No. 3-34412
According to the technique proposed in 3, the engagement pressure at the time of waiting at a low pressure value is large or small due to the variation of the spring load of the control valve, which is the pressure adjusting means, and the variation of the output pressure of the linear solenoid. If the output torque varies, the output shaft torque may drop, and in extreme cases, a gear shift other than the intended gear shift may occur during standby, resulting in a large gear shift shock.

The present invention has been made in view of such circumstances, and further improves the technique proposed in the above-mentioned Japanese Patent Application No. 3-344123 so that the time lag is small regardless of various variations. The purpose is to achieve a downshift of the clutch-to-clutch with a small shift shock.

[0011]

As shown in FIG. 1, the present invention is directed to a shift control device for an automatic transmission which may execute a downshift of a clutch-to-clutch related to two friction engagement devices. In the automatic transmission, the synchronization detecting means for detecting whether or not the input speed of the automatic transmission has reached a synchronization point of a low speed stage, and the friction engagement device on the engagement side before the input speed of rotation reaches the synchronization point. Overlap control means for maintaining the engagement pressure of No. 1 to a predetermined low pressure value and the release pressure of the frictional engagement device on the disengagement side to a predetermined high pressure value capable of forming a torque phase; A means for increasing the engagement hydraulic pressure from the predetermined low pressure value and decreasing the release hydraulic pressure from the predetermined high pressure value after it is detected that the input rotation speed has reached the synchronization point, and the series of overlap control. Formed by A means for detecting a drop in the output shaft torque during the torque phase, and a learning means for changing the predetermined low pressure value in the low pressure standby state on the engagement side depending on the detection of the drop in the output shaft torque. By having
This is a solution to the above problem.

[0012]

In the present invention, when performing downshifting of the clutch-to-clutch, the above-mentioned Japanese Patent Application No. 3-3 is basically used.
The downshift is controlled by the method disclosed in 44123. That is, before the input speed of the automatic transmission reaches the synchronization point of the low speed stage by the pressure adjusting means, the engagement pressure of the frictional engagement device on the engagement side is kept at a predetermined low pressure value, and after the synchronization, the engagement is performed. The pressure is gradually increased.

However, in this technique, for example, when the hydraulic pressure (low pressure value) at the time of standby varies due to various variations, the torque capacity on the engagement side becomes relatively large, There was a problem that the output shaft torque dropped sharply and the shift shock increased.

On the contrary, when the low pressure value varies to the small side, the piston on the engaging side cannot complete its stroke to reach the state of having the torque capacity in the standby state, and thus the synchronous state cannot be achieved. Even if an attempt is made to increase the engagement pressure by detection, the increase is delayed, and as a result, the hydraulic pressure on the release side decreases before the friction engagement device has a sufficient torque capacity. A phenomenon occurs in which a gear shift occurs temporarily, and then the frictional engagement device on the engagement side has a capacity to settle at the final gear stage. As a result, a very large shift shock may occur.

In the present invention, this shift control is basically realized by "overlap shift". Here, the term "overlap speed change" means a state in which the torque capacity of the disengagement side friction engagement device is too large with respect to the torque capacity of the engagement side friction engagement device, that is, the engagement side friction engagement device engagement. The shift control is executed in a state that is relatively too fast with respect to the release of the friction engagement device on the release side. Overlap gear shifting can be easily achieved by setting the release pressure on the release side to be slightly higher.

When the downshift is realized by the overlap shift, the "torque phase" can be formed in the portion corresponding to the overlap. The "torque phase" here is a state in which there is no change in rotational speed for gear shifting (for gear ratio shifting), that is, the input / output rotation of the automatic transmission is uniformly defined by the gear ratio. State, especially in a narrow sense,
Although the force balance of each rotary member is broken (due to gear ratio shifting), the actual rotational speed change due to the broken force balance is not yet generated.

In general, a downshift at a low vehicle speed has a small rotation change width, and therefore, it is better to use the overlap shift than the underlap shift. Therefore,
Virtually no disadvantages arise due to the overlap shifting.

When the basic control is adjusted to the overlap control, when the output shaft torque drops in the torque phase, it can be estimated that the cause of the drop is tie-up. Therefore, when a drop in the output shaft torque during the torque phase is detected in this way, the hydraulic pressure level of the low pressure standby is reduced, while when no drop is detected, it is actually (to the extent that the output shaft torque drops). Since the overlap control was not performed, learning control is performed to maintain or slightly increase the hydraulic pressure level, so that synchronization is always performed at the optimum level regardless of various variations. It becomes possible to stand by, and it becomes possible to extremely smoothly increase the hydraulic pressure after the synchronization. As a result, it is possible to execute a downshift with a small shift shock and a small time lag.

[0019]

Embodiments of the present invention will now be described in detail with reference to the drawings.

First, a concrete example of an automatic transmission to which the present invention is applied is shown by a skeleton in FIG. The automatic transmission 2 includes a torque converter 111, a sub transmission unit 112, and a main transmission unit 113.

The torque converter 111 has a lockup clutch 124. The lockup clutch 124 is provided between a front cover 127 that is integrated with the pump impeller 126 and a member (hub) 129 to which a turbine runner 128 is integrally attached.

The crankshaft (not shown) of the engine 1 is connected to the front cover 127. The input shaft 130 connected to the turbine runner 128 includes an overdrive planetary gear mechanism 13 that constitutes the auxiliary transmission unit 112.
It is connected to one carrier 132.

A clutch C0 and a one-way clutch F0 are provided between the carrier 132 and the sun gear 133 in the planetary gear mechanism 131. The one-way clutch F0 is adapted to be engaged when the sun gear 133 makes a positive rotation relative to the carrier 132 (rotation in the rotation direction of the input shaft 130).

On the other hand, a brake B0 for selectively stopping the rotation of the sun gear 133 is provided. Further, a ring gear 134 which is an output element of the auxiliary transmission section 112 is connected to an intermediate shaft 135 which is an input element of the main transmission section 113.

The sub-transmission unit 112, when the clutch C0 or the one-way clutch F0 is engaged, the planetary gear mechanism 1
Since the whole 31 rotates integrally, the intermediate shaft 135
Rotates at the same speed as the input shaft 130. Further, in the state where the brake B0 is engaged and the rotation of the sun gear 133 is stopped, the ring gear 134 is accelerated with respect to the input shaft 130 and rotates normally. That is, the subtransmission unit 112 can set switching between high and low.

The main transmission unit 113 has three sets of planetary gear mechanisms 140, 150, 160, and these gear mechanisms 140, 150, 160 are connected as follows.

That is, the sun gear 141 of the first planetary gear mechanism 140 and the sun gear 151 of the second planetary gear mechanism 150 are integrally connected to each other, and the ring gear 143 of the first planetary gear mechanism 140 and the second planetary gear mechanism 150 are connected. Carrier 1
52 and the carrier 162 of the third planetary gear mechanism 160 are connected together. Further, the output shaft 170 is connected to the carrier 162 of the third planetary gear mechanism 160. Further, the ring gear 153 of the second planetary gear mechanism 150 is connected to the sun gear 161 of the third planetary gear mechanism 160.

The gear train of the main transmission unit 113 can set one reverse gear and four forward gears, and clutches and brakes therefor are provided as follows.

That is, the ring gear 153 of the second planetary gear mechanism 150 and the sun gear 161 of the third planetary gear mechanism 160.
A clutch C1 is provided between the intermediate shaft 135 and the intermediate shaft 135, and a clutch C2 is provided between the sun gear 141 of the first planetary gear mechanism 140 and the sun gear 151 of the second planetary gear mechanism 150 and the intermediate shaft 135.

A brake B1 for stopping the rotation of the sun gears 141 and 151 of the first planetary gear mechanism 140 and the second planetary gear mechanism 150 is arranged. Also, these sun gear 1
A one-way clutch F1 and a brake B2 are arranged in series between 41 and 151 and the casing 171.
The one-way clutch F1 is adapted to be engaged when the sun gears 141, 151 try to rotate in the reverse direction (rotation in the direction opposite to the rotation direction of the input shaft 135).

The carrier 142 of the first planetary gear mechanism 140
A brake B3 is provided between the casing and the casing 171. In addition, the ring gear 16 of the third planetary gear mechanism 160
A brake B4 and a one-way clutch F2 as elements for stopping the rotation of No. 3 are arranged in parallel with each other between the casing 171. The one-way clutch F2 is adapted to be engaged when the ring gear 163 tries to rotate in the reverse direction.

In the automatic transmission 2 described above, it is possible to perform a shift of one reverse speed and five forward speeds as a whole. FIG. 3 shows an engagement operation table of each clutch and brake for setting these shift speeds. In addition, in FIG. 3, a circle mark indicates an engaged state,
● indicates the engaged state when the engine is braked, and the blank indicates the released state.

As is apparent from this figure, the shift between the second speed and the third speed is the clutch-to-clutch shift of the brake B2 and the brake B3.

Engagement or disengagement of each clutch and brake is executed by driving an electromagnetic valve or a linear solenoid in the hydraulic control device 20 based on a command from the computer 30. The computer 30 includes signals from various sensor groups 40, for example, a vehicle speed signal from a vehicle speed sensor 41 (a signal indicating an output shaft rotation speed N0), a throttle opening signal (accelerator opening signal) from a throttle sensor 42, and a pattern select switch. Basic pattern signals such as a pattern select signal from 43 (a signal selected by the driver for driving with emphasis on power, driving with emphasis on fuel consumption, etc.), a shift position signal from the shift position switch 44, and a foot brake signal from the brake switch 45. In addition, the rotational speed signal of the clutch C0 from the C0 sensor 46 is input. Since the rotation speed of the clutch C0 is the same as the turbine rotation speed (the input shaft rotation speed of the automatic transmission) Nt during the shift between the second speed and the third speed, the rotation speed of the clutch C0 is detected. The turbine rotation speed Nt can be grasped.

The specific hydraulic control of the release and engagement of the brakes B2 and B3 is performed by the hydraulic circuit as shown in FIG.

B3 control valve 202 is 2-3
Oil passage L1 from shift valve 201 to brake B3,
It is arranged in L2 in parallel with the orifice 214 with a check ball. In the valve hole 202a of the B3 control valve 202, a port 221 communicating with the oil passage L1 via a large diameter orifice 202b, a port 222 communicating with the oil passage L2, and an input port 223 also communicating with the oil passage L2. A port 224 is provided to which the signal hydraulic pressure regulated by the SLU linear solenoid valve 203 controlled by the control signal from the computer 30 is supplied. Reference numeral 225 is an input port for the signal pressure of the third speed stage, and 226 is a drain port.

The spool 202c of the B3 control valve 202 has a pair of large diameter lands 227, 228 and a small diameter land 229. One end of the large diameter land 228 is in contact with the pressure receiving piston 202e via the spring 202d, and the outer end of the large diameter land 227 receives the supply pressure from the port 223. The outer end of the pressure receiving piston 202e receives the signal pressure from the input rotation signal port 224, and the inner end of the pressure receiving piston 202e receives the signal pressure from the port 225.

On the other hand, the B2 accumulator 204 is connected to an oil passage L3 leading to the brake B2 via an orifice 204a. Therefore, by controlling the back pressure of the B2 accumulator 204 by the accumulator back pressure control device 206, it is possible to adjust the drain pressure of the oil passage L3 even when the brake B2 is released.

In the figure, reference numeral 210 is a B2 orifice control valve for speeding up (first fill) hydraulic pressure supply to the brake B2, and 211 is an upshift from the second speed stage to the third speed stage. 2 to 3, a 2-3 timing valve for adjusting the drain pressure from the brake B2 by a signal from the SLU linear solenoid valve 203 will be described. However, since these are not directly involved in the downshift control which is the subject of the present invention, detailed description will be given. The description of the configuration is omitted.

In the hydraulic control device constructed as described above, in the third speed stage, the drive range pressure D passed through the manual valve (not shown) becomes 1-2 shift valve 21.
It is supplied to the brake B2 via the 2-3 shift valve 201 and the oil passage L3, and as a result, the brake B2 is brought into the engaged state. The force balance at this time is shown in FIG. Brake B
If the torque capacity TB2 of 2 is ρ2 · Tin or more, the third speed is maintained, and the output shaft torque becomes (1 + ρ2) · Tin due to the dynamic balance (described later).

Here, a shift solenoid valve (not shown) is operated in accordance with the running condition of the vehicle, and the 2-3 shift valve 201 of FIG. 4 is in the second gear position (the oil passage in the valve is indicated by a solid line). The pressure of the brake B2 starts to be drained from the oil passage L3 through the 2-3 shift valve 201. In this embodiment, the drop of the drain pressure is caused by controlling the back pressure of the B2 accumulator 204 by the accumulator back pressure control device 206 so that the input rotational speed Nt.
Is controlled so as to be a predetermined ratio. In this case, initially, at least the torque capacity of the brake B2 is {ρ1 /
(1 + ρ1)} · TB3 + ρ2 · Tin or more is set to be secured. Note that this will be described later.

On the other hand, the drive range pressure D is the brake B3 from the 2-3 shift valve 201 via the oil passages L1 and L2.
Is supplied to. The hydraulic pressure supply to the brake B3 follows the procedure of the flowcharts shown in FIGS. 5 and 6 (FIG. 6 is a continuation of FIG. 5).

This procedure will be described below with reference to FIG. 5 and FIG. 6 and still with reference to other drawings.

First, in step 302, the 2-3 shift valve 201 is in a communication state (brake B) shown by the solid line in FIG. 4 by the control of a shift solenoid valve (not shown).
2 is connected to the drain circuit and brake B3 is connected to the supply circuit). At this time, as shown in step 304, the back pressure of the B2 accumulator 204 is feedback-controlled as described above so that the input rotation speed basically becomes the target rotation speed change.
The drain pressure of 2 is feedback controlled. This feedback control is basically continued until the second speed synchronization is achieved.

At the initial stage of this change in the number of revolutions, the spring of the B3 control valve 202 causes the spool 202c to take the upper half position of FIG.
A rapid hydraulic pressure supply through the large orifice 202b communicating with the piston 2 causes a piston stroke of the brake B3, and the piston is immediately displaced to a position where the friction material can be engaged, so-called "first fill" operation is performed.

As will be described later in step 306, in step 308, the SLU linear solenoid valve 20 supplied to the port 224 of the B3 control valve 202 is supplied.
The signal pressure by 3 is controlled to a predetermined low pressure value (predetermined by the previous shift). That is, the brake B3 is in a standby state in which the supply pressure (hereinafter referred to as B3 pressure) to the brake B3 is kept low so that the brake B3 slightly transmits the torque. This low pressure standby basically prevents a sudden change in rotational speed due to an increase in B3 pressure and excessive tie-up of both brakes B2 and B3. This low-pressure standby state is specifically achieved by draining the excess B3 pressure through the B2 orifice control valve 210 by releasing the port 226 by applying a feedback pressure from the port 223.

During this low pressure standby, the hydraulic pressure of the brake B2 is maintained slightly higher (than the hydraulic pressure estimated to be the proper timing) to form the torque phase. As a result, the torque transfer is performed under the overlap control. Steps 310-320 and 330-
340 corresponds to a step for learning control (according to the present invention) executed based on this state.

Now, skipping these steps and explaining the basic flow of steps 322, 328 and 342 to 346 for the sake of convenience, in step 322, from the rate of change of the input speed Nt in the computer 30, Calculation (estimation) of the time t until the synchronization of the second gear is performed, and this value is compared with the set value t0. While the result t of this comparison exceeds t0, the process returns from step 328 to step 304, and the flow of steps 304 to 322 is repeated.

When the value of the time t becomes less than the set value t0, the operation for increasing the B3 pressure is performed at step 342 at an increasing rate according to the input torque Tin. Due to this boost, both brakes B2 and B3 cooperate to share the torque,
A smooth transition to the synchronization point is possible due to the displacement of the torque sharing ratio. This state continues until it is confirmed in step 344 that the second gear is synchronized.

When the synchronization is finally confirmed, step 34
At 6, the SLU linear solenoid valve is set to high output and B
3 The port 221 of the control valve 202 is completely in communication with the port 222, and the brake B3 is rapidly engaged, while the signal pressure is supplied from the solenoid valve 213 to the B2 orifice control valve 21.
0 opens, and a rapid drain of B2 pressure is performed. This series of steps completes the downshift from the third speed to the second speed.

Next, the learning control of a predetermined value in the low pressure standby state according to the present invention, which has been skipped above, will be described.

First, as a preliminary explanation, the flow of the output shaft torque at the time of downshifting from the third speed to the second speed will be described with reference to FIGS. 7 to 9.

FIG. 7 shows the force balance of each part in the state of the third speed. As described above, if the torque capacity TB2 of the brake B2 is equal to or greater than the transmission torque ρ2 · Tin, the third speed is maintained and the output shaft torque becomes (1 + ρ2) · Tin due to the dynamic balance.

FIG. 8 shows the state of the torque phase from the third speed stage to the second speed stage. When the brake B3 starts to be engaged, the output shaft torque increases from (1 + ρ2) · Tin to (1 + ρ2) · Tin as the torque capacity TB3 of the brake B3 increases.
-TB3 / (1 + ρ1), which is reduced by TB3 / (1 + ρ1).

However, when the torque capacity TB2 of the brake B2 is greater than or equal to (ρ1 / (1 + ρ1)) · TB3 + ρ2 · Tin, there is no change in rotational speed for shifting from the third speed to the second speed. , The torque phase is maintained.

FIG. 9 shows the force balance in the inertia phase (a state in which the rotational speed changes due to the rotation member shifting from the third speed to the second speed).

The torque capacity TB2 of the brake B2 is (ρ1 /
(1 + ρ1)) ・ TB3 + ρ2 ・ Tin When it becomes less than or equal to Tin, an imbalance of torque occurs between the input system and the output system, and the input system is accelerated to shift to the inertia phase.

Here, ρ1 is the ratio of the number of teeth of the ring gear 143 of the front planetary gear mechanism 140 and the sun gear 141, and ρ2 is the ratio of the number of teeth of the ring gear 153 of the central planetary gear mechanism 150 and the sun gear 151, respectively. Shows.

As is apparent from this description, in this embodiment, the hydraulic pressure (drain pressure) of the brake B2 during the low pressure standby is set slightly higher in order to basically perform the overlap control. , If the standby hydraulic pressure (B3 pressure) during low pressure standby becomes too high due to the overwhelming relationship,
It can be seen that the amount of tie-up increases (because both hydraulic pressures are too high), and the drop in output shaft torque increases correspondingly. Therefore, by considering how much the overlapped state should be applied, by giving a proper threshold value to the drop of the output shaft, it is judged whether or not the predetermined value at the time of low pressure standby is proper (how it varies). Will be able to.

The output shaft torque during shifting is proportional to the acceleration of the output shaft rotation speed from the equation of motion. Therefore, the same effect as monitoring the output shaft torque can be obtained by monitoring the angular acceleration dω0 of the output shaft rotation speed. Further, even if the angular acceleration of the rotational speed of a member that rotates in the same manner as the output shaft, for example, the front or intermediate sun gear mechanism 140, 150 sun gear 141 (that is, clutch C2) is observed, the output shaft torque changes as a result. Can be monitored.

When actually detecting the drop in the output shaft torque, it is desirable not to execute the learning control in the following cases in order to improve the reliability.

1) When the brake is depressed 2) When it is judged that the vehicle is traveling on a rough road 3) When it is judged that the vehicle is turning a curve 4) Fluctuation of output shaft torque (acceleration of output shaft rotation speed) Is large 5) The bending line of the drop in the output shaft torque is 2 during the torque phase.
When detected more than once

In such a case, it is advisable to stop the learning control based on the current shift and use the already-obtained predetermined value for the low pressure standby state as it is.

Since this learning control intentionally forms the overlap state, it is not appropriate to execute it at high vehicle speed, and therefore it is executed on condition that the vehicle speed is low. However, as a result of this learning, good clutch-to-clutch shifting can be executed even at high vehicle speeds.

FIG. 10 shows a waveform at the time of gear shifting with the throttle depressed, and the output shaft torque also increases as the input torque Tin increases. Even when the input torque is constant, the control can be executed because the output shaft torque (output shaft acceleration) drops and the bending point appears at the stage when the brake B3 reaches the completion point of the piston stroke.

The learning control described above is executed by the steps described below.

Since the specific contents of each step have already been described in detail, the learning control procedure will be mainly described here.

First, at step 306, it is judged if the vehicle speed is low. When the answer is "No", the following learning control is not performed.

At step 310, it is confirmed that the value of the flag F is 2. In this flowchart, in steps 312, 314, 316, and 320, it is determined whether or not the state is suitable for executing the learning control (as described above), and if even one of them is determined to be not in the suitable state, the flag F is determined. Is set to 2 (step 324). Therefore, the process proceeds to step 312 and subsequent steps on condition that the flag F is not 2 in step 310.

In step 318, the angular acceleration dω0 of the output shaft is calculated from the rotation of the output shaft.

After that, the value of the flag F is confirmed once again in step 330 through steps 322 and 328 (which have already been described).

When the value of the flag F is 2, the learning control steps of steps 332 to 338 are bypassed.

When the value of the flag F is not 2, the routine proceeds to step 332, where the bending point is detected from the change in the value of the output angular acceleration dω0, and the maximum value dω0 of the bending points is thereby obtained.
The max and the minimum value dω0 min are obtained.

In step 334, the number of bending points is confirmed, and when two or more bending points are detected, it is possible that some large disturbance is mixed in at the time of detection, and it is judged that the reliability of the detected value itself is low. However, learning control is not performed.

At steps 336 and 338, the difference k between the maximum value dω0 max and the minimum value dω0 min of the bending points is obtained, the degree of the drop in the output shaft torque is judged from this, and based on this difference k, the next shift step A predetermined value (standby pressure) of the brake B3 at 308 at low pressure standby is learned and corrected.

In this embodiment, at the time of downshifting at a low vehicle speed, a torque phase is formed by positively setting the overlap state, and a drop in the output shaft torque in this torque phase is confirmed to reduce the low pressure. Since the suitability of the predetermined value during standby is determined and this is learned and corrected for the next shift, the predetermined value during low pressure standby is set to the most appropriate value of the automatic transmission at that time regardless of various variations. You will be able to set the value.
Further, since this learning control is executed only in the situation determined to be suitable for executing the learning control, the reliability of the learning control can be further improved.

The learning result is naturally reflected not only in the low-speed 3 → 2 downshift, but also in the high-speed 3 → 2 downshift. Further, it is also reflected when the same brake is used in other shifts (for example, the brake B3 of 1 → 2 shift).

[0078]

As described above, according to the present invention, the downshift of the clutch-to-clutch can be executed with a small time lag and a small shift shock regardless of various variations. The effect is obtained.

[Brief description of drawings]

FIG. 1 is a block diagram showing the gist of the present invention.

FIG. 2 is a block diagram schematically showing an automatic transmission for a vehicle to which the present invention is applied.

FIG. 3 is a diagram showing an operating state of each friction engagement device of the automatic transmission.

FIG. 4 is a hydraulic circuit diagram in which a portion related to a downshift from the third speed stage to the second speed stage is extracted and shown in the hydraulic control device for the automatic transmission.

FIG. 5 is a flow chart showing a basic control flow of the above-mentioned speed change.

6 is a flowchart showing a continuation of the control flow of FIG.

FIG. 7 is a diagram showing force balance in the third speed state.

FIG. 8 is a diagram showing force balance in the torque phase.

FIG. 9 is a diagram showing force balance in the inertia phase.

FIG. 10 is a diagram showing a shift characteristic at the time of downshifting from the third speed to the second speed.

[Explanation of symbols]

 B2, B3 ... Brake 20 ... Hydraulic control device 30 ... Computer 40 ... Various sensor parts

Front Page Continuation (72) Inventor Kunihiro Iwatsuki 1 Toyota-cho, Toyota-shi, Aichi Toyota Motor Co., Ltd. (72) Inventor Mamoru Niimi 10 Takane, Fujii-cho, Anjo-shi, Aichi Aisin AW shares In-house (72) Inventor Masahiko Ando 10 Takane, Fujii-cho, Anjo City, Aichi Prefecture Aisin AW Co., Ltd. (72) Inventor Yoshihisa Yamamoto 10, Takane Fujii-cho, Aichi Prefecture Aisin AW Shares In-house (72) Inventor Akira Fukatsu 10 Takane, Fujii-cho, Anjo-shi, Aichi Aisin AW Co., Ltd.

Claims (1)

[Claims] 1. A shift control device for an automatic transmission, which may execute a downshift of a clutch-to-clutch related to two friction engagement devices, wherein an input speed of the automatic transmission reaches a synchronization point of a low speed stage. Synchronous detection means for detecting whether or not, before the input speed reaches the synchronization point, while maintaining the engagement pressure of the friction engagement device on the engagement side at a predetermined low pressure value,
Overlap control means for maintaining the release pressure of the friction engagement device on the release side at a predetermined high value capable of forming a torque phase, and the synchronization detection means detected that the input rotation speed has reached the synchronization point. After that, means for increasing the engagement hydraulic pressure from the predetermined low pressure value and decreasing the release hydraulic pressure from the predetermined high pressure value, and a drop in the output shaft torque in the torque phase formed by the series of overlap control are performed. An automatic transmission comprising: a detecting unit; and a learning unit that changes the predetermined low pressure value during low pressure standby on the engagement side depending on detection of a drop in the output shaft torque. Shift control device.