JPH05306760A - Clutch to clutch speed change controller for automatic transmission - Google Patents

Abstract

PURPOSE:To execute speed change with a speed change shock kept small by providing a means for determining the proceeding state of speed change, a means for regulating pressure so as to decrease disengaged oil pressure on a disengaging side as engaged oil pressure on an engaged side increases, and a means for early correcting the dis engaged oil pressure somewhat higher according to the proceeding state of the speed change. CONSTITUTION:An automatic transmission speed change controller to execute clutch to clutch speed change by disengaging and engaging frictional engaging elements has a means for determining the proceeding state of speed change and a pressure regulating means to regulate the disengaged oil pressure of the frictional engaging elements on disengaged side during the torque phase of the speed change so as to decrease the disengaged pressure at a fixed decreasing rate with an increase in engaging oil pressure according to the engaging oil pressure of frictional engaging elements on an engaged side. The disengaged oil pressure is corrected according to the proceeding state of the speed change by a correcting means so that it may be somewhat higher than the disengaged oil pressure set up on the basis of the above relationship at the early stage of the torque phase.

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 that executes clutch-to-clutch shift by releasing and engaging a friction engagement element.

[0002]

2. Description of the Related Art When performing a specific gear shift of an automatic transmission, it is often necessary to simultaneously release and engage two friction engagement elements (including a clutch and a brake in a broad sense). So-called clutch to clutch shift).
In this case, if the release and engagement of each friction engagement element 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 clutch is generally provided so as to prevent such a problem from occurring. Was there.

However, the method of synchronizing each clutch by using the one-way clutch in this way is as follows.
The cost increases because the one-way clutch is attached, and
There are problems such as an increase in weight and an occupied space.

In view of these points, 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.

For example, in Japanese Patent Laid-Open No. 64-65354, the frictional engagement on the side where the timing valve is actuated to release when the engagement hydraulic pressure of the frictional engagement element on the engagement side exceeds a predetermined value. A technique has been proposed in which a large-diameter orifice that drains the release hydraulic pressure of an element at once is opened to control the timing of engagement and release.

In this case, the publication also discloses that the throttle pressure and the engagement pressure are applied to the valve spool so as to face each other, and the orifice of the drain pressure is switched at a timing according to the engine load. ..

[0008]

However, even in the technique disclosed in Japanese Patent Laid-Open No. 64-65354 described above, the switching timing of the timing valve is determined by the initial setting. Cannot be corrected and the drain pressure is discharged all at once at a predetermined timing, so once the discharge is started, the drain pressure must be recovered after that. There has been a problem that it is extremely difficult to (decrease the drain speed), so that it is impossible to promptly correct the switching timing corresponding to the change of the input torque.

In view of such a point, the applicant first adjusts the engagement pressure and the drain pressure (release pressure) to a predetermined relationship according to the progress state of the gear shift, so that the extremely low temperature or the gear shift can be achieved. We proposed a hydraulic control method that can adjust the timing of release and engagement so that it can respond quickly to changes in the throttle opening, and that can suppress the occurrence of gear shift shock (Japanese Patent Application No. Hei 3). -344124: unknown).

That is, in this technique, the release hydraulic pressure of the released frictional engagement element is reduced at a predetermined decrease rate according to the engagement hydraulic pressure of the engaged frictional engagement element as the engagement hydraulic pressure increases. The release hydraulic pressure is constantly lowered linearly with respect to the engagement hydraulic pressure by adjusting the pressure to 1. Thus, it is possible to favorably deal with various variations and changes in input torque.

However, in this technique, when the friction coefficient μ of the friction engagement element on the release side becomes low, the margin ratio (margin allowance) of the release hydraulic pressure of the friction engagement element on the release side that is set decreases. There is a problem that the friction engagement element on the disengagement side may slip out early in the initial phase of the torque phase, which may cause the engine to squirt up.

That is, this margin is originally the friction coefficient μ
It is provided to prevent the disengagement side frictional engagement element from slipping in the initial stage of the torque phase even if there are variations in the valve, manufacturing errors in the valve, etc. If you simply set
There arises a problem that the drainage of the friction engagement element in the general time (normal time) becomes too slow. Further, at the end of the torque phase, the timing at which the torque capacity on the disengagement side becomes 0 and the timing at which the torque capacity on the engagement side becomes equal to or greater than the total torque capacity to be transmitted are largely deviated from each other. Of course, just adjusting various constants that are proportional to the engagement hydraulic pressure on the engagement side causes the timing at the end of the shift to shift, so there is a limit to the range of change.

The present invention has been made in view of the above-mentioned problems, and the releasing hydraulic pressure of the releasing side friction engaging element and the engaging hydraulic pressure of the engaging side friction engaging element in the torque phase of the clutch-to-clutch shift. Regulates and in a more rational relationship,
The clutch-to-clutch shift is executed without generating engine injection and with a smaller shift shock, thereby solving the above problem.

[0014]

SUMMARY OF THE INVENTION According to the present invention, in a clutch-to-clutch shift control device of an automatic transmission for executing clutch-to-clutch shift by releasing and engaging a friction engagement element, the progress state of the shift is obtained. Means for decreasing the release hydraulic pressure of the disengagement side frictional engagement element in accordance with the engagement hydraulic pressure of the engagement side frictional engagement element at a predetermined decrease rate in accordance with the engagement hydraulic pressure of the engagement side frictional engagement element during the torque phase of gear shifting. By providing a means for adjusting the pressure and a means for correcting the release hydraulic pressure so that the release hydraulic pressure becomes higher in the initial phase of the torque phase than the release hydraulic pressure set by the relationship, according to the progress state of the shift, This is a solution to the above problem.

[0015]

In the present invention, the disengagement hydraulic pressure (drain hydraulic pressure) is controlled so as to decrease with respect to the engagement hydraulic pressure so that the release hydraulic pressure decreases at a predetermined decrease rate with the increase of the engagement hydraulic pressure. Is maintained at a proper value. Therefore, even if there is a temperature change or a throttle change during shifting, the relationship between the two hydraulic pressures can always be maintained properly.

Further, in the present invention, the progress state of the shift is obtained, and the release hydraulic pressure is set to be higher in the initial stage of the torque phase than the release hydraulic pressure set by the above relation in accordance with the progress state (torque phase). By the end of, I am trying to make a correction so that it becomes close to the above relationship. As a result, the drain speed becomes too slow during normal operation (when shifting the friction engagement element independently), or the release side friction engagement occurs at the end of the torque phase during the clutch-to-clutch shift. The torque displacement remains in the elements, or the frictional engagement elements cannot secure the torque capacity to be transmitted. It becomes possible to secure a sufficient margin.

Therefore, it is possible to prevent the release side frictional engagement element from slipping out early without causing any trouble.

[0018]

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 in 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 rotates forward relative to the carrier 132 (rotates in the rotational 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 has the planetary gear mechanism 1 when the clutch C0 or the one-way clutch F0 is engaged.
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 to rotate 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 speed and four forward speeds, and a clutch and a brake 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 and 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.

The automatic transmission 2 described above is adapted to perform a shift of 1 reverse speed and 5 forward speeds, and FIG. 3 shows an engagement operation table of each clutch and brake for setting these shift speeds. Show. In addition, in FIG. 3, a circle indicates an engaged state, and
The mark indicates the engaged state during engine braking, and the blank indicates the released state.

As is apparent from this figure, in this embodiment, for example, the upshift from the second speed to the third speed is the release of the brake B3 and the clutch-to-clutch shift by the engagement of the brake B2. I understand.

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 of output shaft rotation speed N0), a throttle opening signal from a throttle sensor 42 (accelerator opening signal), and a pattern select switch. The basic signal 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.

FIG. 4 shows the hydraulic control device 20 of this automatic transmission.
Of these, parts related to the clutch-to-clutch shift from the second speed stage to the third speed stage are extracted and shown.

In the figure, 201 is a 2-3 shift valve, 202 is a 2-3 timing valve, 203 is an SLU linear solenoid valve, 204 is a B3 accumulator, and 205 is an accumulator control valve for controlling its back pressure.

The structure of each part will be described in detail. The import 2 which communicates with the return oil passage L1 of the 2-3 shift valve 201 is provided in the valve hole 202a of the 2-3 timing valve 202.
21, a drain pressure input port 222 communicating with the oil passage L1 via an orifice 202e, and a supply pressure input port 223 communicating with the supply oil passage L2 from the 2-3 shift valve 201 to the brake B2 via an orifice 202f.
And an input torque signal port 224 to which a signal hydraulic pressure regulated by the SLU linear solenoid valve 203 controlled by a control signal from the computer 30 is supplied by a signal corresponding to the input torque, and a drain port 225. There is.

The spool 202b inserted into the valve hole 202a is located at one end thereof and has a drain port 225.
Is located in the middle with the land 226 for controlling the opening degree of the drain pressure input port 222 and the partition between the import 221.
A land 227 that forms a drain pressure receiving surface on the drain pressure input port 222 side and a land 227 that is located at the other end and forms a supply pressure receiving surface, and has a small diameter that separates the supply pressure input port 223 and the drain pressure input port 222. The land 226 is provided, and the land 226 on one end side is in contact with the pressure receiving piston 202d via the spring 202c. The pressure receiving piston 202d constitutes a pressure receiving surface for receiving the signal pressure from the input signal port 224.

B3 accumulator 204 is a brake B
B2 accumulator 20 for brake B2 is connected to the oil passage L3 leading to 3 through an orifice 204a.
As in the case of 8, the pressure of the supply oil passage L3 can be adjusted even during the drain by the back pressure control. In the figure, reference numeral 209
Indicates a B2 orifice control valve which constitutes a first fill means for speeding up the hydraulic pressure supply to the brake B2 in its initial stage, and 210 is involved in shifting from the third speed to the second speed, and the spool is at the upper end of the figure. A B3 control valve that locks to the left shown in the section is shown. However, since they are not directly involved in the hydraulic control which is the subject of the present invention, a detailed description of the configuration is omitted.

In the hydraulic control circuit configured as described above, in the state of the second speed, the drive range line pressure passing through the manual valve (not shown) causes the 1-2 shift valve 211, 2-3 shift valve 201, and the supply oil. It is supplied to the brake B3 via the path L3, which is in the engaged state.

Here, a shift solenoid valve (not shown) is operated by a control signal of the computer 30 according to the traveling condition of the vehicle, and the 2-3 shift valve 201 is in the third gear position (the oil passage in the valve is indicated by a solid line). The hydraulic pressure of the brake B3 starts to flow from the supply oil passage L3 to the drain oil passage L1 through the 2-3 shift valve 201, and a part of the orifice 21 has a very small diameter for compensating the valve stick.
While being drained through 2, it flows into the port 222 through the port 221 and the orifice 202e.

On the other hand, the drive range line pressure is 2-3
Brake B from shift valve 201 through oil supply path L2
Supplied to 2. At this time, the B2 orifice control valve 209 receives the spring force and switches to the position shown in the upper half of the figure, and the orifice 213 with a check ball
Is supplied to the brake B2 all at once from a large-capacity in-valve oil passage that bypasses the valve 209, and the piston of the brake B2 is stroked all at once until the brake B2 is engaged, and the play is removed.

At the end of this fast fill operation, the valve 209 is displaced to the position shown in the lower half of the figure by the feedback of the oil pressure in the supply oil passage L2 (hereinafter referred to as B2 engagement pressure), and slowly passes through the orifice 213 with a check ball. B2 pressure starts to gradually increase with the supply of a sufficient hydraulic pressure, and pressure is accumulated in the accumulator 208, while the pressure increase at the port 223 of the 2-3 timing valve 202 causes the spool 202b to move to the lower half of the drawing as shown in the lower half of the figure. When the drain port 225 is displaced to the right and the drain port 225 is released, the pressure in the drain oil passage (that is, the oil passage L3) corresponding to the pressure increase in the port 223 (hereinafter referred to as B3
This is called release pressure).

At this time, the B2 engagement pressure is input and the B3 release pressure has a predetermined relationship, that is, the B3 release pressure decreases at a predetermined decrease rate with the increase of the B2 engagement pressure with respect to the B2 engagement pressure. Therefore, as shown in FIG. 5, the operation of linearly decreasing the B3 release pressure from the start of the shift to at least the end of the torque phase is performed. In the figure, the solid line shows the theoretical characteristics derived from the calculation of the relationship between the B2 engaging pressure and the B3 releasing pressure with the input torque as a parameter, and the one-dot chain line shows the flow resistance change due to temperature change and the friction coefficient μ of the friction engaging element. In consideration of variations and the like, the actual characteristics are shown in which the ratio of B3 pressure is slightly increased so that slipping does not occur at the initial stage of the torque phase. The difference between the theoretical characteristic and the actual characteristic, which corresponds to the above-mentioned "margin ratio (or margin margin)".

When the throttle is operated during the above-described descent control, the change in the input torque corresponding to the operation is calculated by the computer 30, and the change is reflected in the SLU linear solenoid valve 203. As a result, it is applied from the port 224. The piston 202d is displaced by the change of the signal pressure, the spring load of the spring 202c is changed, the spring 202c is further compressed, and the B3 release pressure is the signal pressure of the port 224 when the 2-3 timing valve 202 and the piston 202d are connected. Therefore, the characteristic line is changed to the upper or lower side by changing the parameters shown in FIG.

For example, in the case of power-on (a state in which the accelerator is depressed and power is transmitted from the engine side to the wheel side), the B3 release pressure is applied by the back pressure of the accumulator 204 so as to prevent the engine from rising. The pressure is increased and the overlap state is reached. On the contrary, when the power is off (the accelerator is released and the power is transmitted from the wheel side to the engine side), there is no concern that the engine will blow up, so the spring load of the spring 202c is reduced and the brake B3 release pressure is reduced. Reduced to an underlap condition.

FIG. 6 shows the changes over time in the actual brake B2 engagement pressure, B3 release pressure, output torque, and input shaft rotation speed during gear shifting. In the figure, x indicates a first fill period, y indicates a pressure adjusting period according to the present invention, and a broken line indicates a conventional release pressure drop state. The drop in the output torque curve at the end of the torque phase is due to the torque pull-in due to the brake tie-up.

Next, FIG. 7 shows a control flow in which the margin of the B3 release pressure is increased at the initial stage of the torque phase.

First, after various input processes are performed in step 302, it is determined in step 304 whether or not the shift from the second speed stage to the third speed stage in the power-on state is determined. In this embodiment, the present invention is applied only during the clutch-to-clutch shift from the second speed stage to the third speed stage. Therefore, in the case of other gear shifts or when no particular gear shift determination is made, the control flow is left as it is.

When the shift from the second gear to the third gear is judged, the routine proceeds to step 306, where the 2-3 shift valve 201 is switched to the third gear position and the brake B2 is turned on as described above. The engagement pressure and the release pressure of the brake B3 are aP _B2
The pressure adjustment is controlled so that + bP _B3 = (c + α) P _SLU . This pressure regulation is continued until it is determined in step 308 that the torque phase has started.

Eventually, the torque phase starts in step 308.
If it is determined that the pressure adjustment formula is aP _B2+ BP_B3= {C +
γ (t)} P_SLUIs changed to.

Note that, as shown in FIG. 8, γ (t) is a value that has a relationship of decreasing from α with time. As a result, as shown in FIG. 5, it is possible to obtain such a characteristic that the margin ratio is large at the initial stage of the torque phase and becomes small at the end of the torque phase.

Incidentally, the reason why the margin ratio is increased only in the initial stage of the torque phase is that the 2-3 shift valve 2 is used as shown in FIG.
01 to 2-3 Port 223 of timing valve 202
It can also be achieved by providing a pressure regulating valve 400 in the oil passage L2 leading to.

That is, as shown in FIG. 9, when the pressure regulating valve 400 is provided in this manner, the 2-3 timing valve 202 is operated until the engagement hydraulic pressure P _B2 of the brake B ₂ reaches a predetermined value and thereafter. The balance equation can be changed, and the release pressure of the brake P _B3 can be increased until the engagement pressure P _{B2 of the} brake B2 becomes a predetermined value or more.

As described above, according to the present invention, basically, the engagement pressure P _B2 of the brake _B2 and the release pressure P _B3 of the brake B3 are reduced at a predetermined rate as the release pressure P _B3 increases as the engagement pressure P _B2 increases. The purpose is to set the release pressure P _B3 of the brake B 3 given by this relational expression to a higher value only in the initial phase of the torque phase while adjusting the pressure in a decreasing relationship, and it is not always necessary to leave a linear margin with time. It is not essential to reduce the rate.

According to the apparatus of the above embodiment, the release pressure is adjusted in a predetermined relationship with the increase of the engagement pressure, and the input pressure is engaged by the solenoid pressure output from the linear solenoid valve 203. It is configured to change the relationship of release pressure. Further, the linear solenoid valve 203 for controlling the pressure regulation of the 2-3 timing valve 202 is
It is independent of the control of the engagement pressure by the 3-shift valve 201, and the switching control can be performed regardless of the characteristic of the engagement. Further, regarding the relationship between the engagement pressure and the release pressure, the 2-3 timing valve 202 is set so that the margin ratio is high near the start of the torque phase and becomes low at the end of the torque phase. This prevents the brake B3 from slipping out early in the initial phase of the torque phase even if there is a change in the μ characteristic of the friction engagement element, the engine torque, or the like over time, or other variations, thus preventing the engine from rising. In addition to being able to prevent it, each brake can be controlled so as to have a predetermined torque capacity at the end of the torque phase, and the variation at the end of the torque phase can be reduced.

[0056]

As described above, according to the present invention,
Even if there is a change in the input torque or the like during clutch-to-clutch shifting, the engagement hydraulic pressure on the engagement side and the release hydraulic pressure on the disengagement side can always be controlled in an appropriate relationship, and at the beginning of the torque phase. Since the larger margin ratio is set, the disengagement side friction engagement element may slip off early in the initial phase of the torque phase while reducing the variation in the torque capacity that each friction engagement element should have at the end of the torque phase. Can be prevented.

[Brief description of drawings]

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

FIG. 2 is a block diagram showing an outline of 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 partial circuit diagram showing a main part of a hydraulic control circuit of the automatic transmission.

FIG. 5 is a diagram showing the relationship between the hydraulic pressure of the brake B2 and the hydraulic pressure of the brake B3.

FIG. 6 is a specially-made gear shift transient diagram during clutch-to-clutch shifting from the second speed stage to the third speed stage of the automatic transmission.

FIG. 7 is a flow chart showing a control flow relating to the change of the pressure adjusting method during the clutch-to-clutch shifting from the second speed stage to the third speed stage.

FIG. 8 is a diagram showing the relationship between α and γ (t) in the above flow chart.

FIG. 9 is a partial hydraulic circuit diagram showing another embodiment of the present invention.

[Explanation of symbols]

30 ... Computer, 201 ... 2-3 shift valve, 202 ... 2-3 timing valve, 203 ... SLU linear solenoid valve, 204 ... B3 accumulator, 205 ... Accumulator control pressure control solenoid valve, B2 ... Brake (engagement side friction member) B3 ... Brake (release side frictional engagement element).

 ─────────────────────────────────────────────────── ─── Continuation of the front page (72) Inventor Masahiko Hayabuchi 10 Takane, Fujii-cho, Anjo-shi, Aichi Aisin AW Co., Ltd.

Claims (1)

[Claims] 1. A shift control device for a clutch-to-clutch of an automatic transmission which executes clutch-to-clutch shift by releasing and engaging a friction engagement element, means for determining a progress state of the shift, and a torque phase of the shift. A means for adjusting the release hydraulic pressure of the disengagement side frictional engagement element to a relationship in which the disengagement hydraulic pressure of the disengagement side frictional engagement element decreases at a predetermined decrease rate as the engagement hydraulic pressure increases; A clutch toe for an automatic transmission, comprising: means for correcting the release hydraulic pressure to be higher in the initial phase of the torque than the release hydraulic pressure set by the relationship according to the state. Clutch shift control device.