JPH09229180A - Control device of automatic transmission - Google Patents

Abstract

PROBLEM TO BE SOLVED: To optimize the initial oil pressure of an engage side frictional engagement device in an automatic transmission to execute so-called equal speed shifting. SOLUTION: This control device concerned is for an automatic transmission to execute the initial hydraulic control of an engage side frictional engagement device at the time of gear shifting and is coupled with a drive motive to raise the revolving speed temporarily at the time of specified down-shifting in the automatic transmission. The arrangement further includes a shift judging means to judge shift operation (Step 2), an equal speed shift judging means to judge whether or not the shift is such as to raise temporarily the revolving speed of the drive motive (Step 3), and an initial hydraulic control means to alter the controlling contents of the initial hydraulic control (Step 10) in accordance with the result from judging by the equal speed shift judging means.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a control device for an automatic transmission, and particularly to a hydraulic pressure for so-called clutch-to-clutch shift or direct pressure control depending on the control of a driving force source such as an engine or a motor. The present invention relates to a device for controlling the electric field.

[0002]

2. Description of the Related Art Since a shift shock in a vehicle equipped with an automatic transmission causes a loss of riding comfort, various controls have been conventionally performed to improve the shift shock.
As an example of such control, there is known a control for improving a shock caused by an abrupt increase in braking force accompanying a downshift. For example, when a downshift in a power-off state is performed, an engine throttle valve is electrically operated. Control is performed to temporarily increase the throttle opening, thereby increasing the engine speed to the synchronous speed at the gear after the gear shift, and executing the gear shift in that state to execute the gear shift. This control prevents abrupt changes in the engine speed that accompany it. This type of control is sometimes called constant velocity shift, and an example thereof is described in Japanese Patent Laid-Open No. 5-231525.

In the invention described in this publication, the line hydraulic pressure is prevented from increasing with the increase of the throttle opening, whereby the engagement of the friction engagement device is achieved. It prevents the shock caused by the sudden increase in pressure.

Further, since gear shifting in an automatic transmission is executed by controlling engagement / disengagement of a multi-disc clutch or a multi-disc brake, a gear shift shock depends on how hydraulic pressures supplied to and discharged from these friction engagement devices are controlled. Is greatly affected. Therefore, for example, two friction engagement devices should be engaged at the same time.
In the case of so-called clutch-to-clutch shift or direct pressure control that is executed by releasing the hydraulic pressure, the hydraulic pressure to be supplied to the friction engagement device on the engaging side is temporarily increased at the same time as or immediately after the shift output. Thus, by filling the so-called pack clearance, the initial hydraulic pressure control is executed to immediately control the frictional engagement device to have the torque capacity by supplying the hydraulic pressure higher than that.

This initial hydraulic pressure control is a control for putting the frictional engagement device into a standby state in which it can be substantially engaged immediately. In other words, if the initial oil pressure is low, such as insufficient control, there will be a time delay until the frictional engagement device engages substantially, which will deteriorate the speed response, and if the initial oil pressure is too high, However, the friction engagement device may have a torque capacity of a predetermined value or more, and as a result, the control of the subsequent low pressure standby may not be successful.

[0006]

By the way, recently, there is a case where the above-mentioned so-called constant speed shift is executed. In the case of this shift, the shift is performed in the power-off state at the same time as the shift is executed or simultaneously with the execution of the shift. Since the engine speed is increased irrespective of the operation of the accelerator pedal, the input torque to the automatic transmission is different from that in the normal power off state. Therefore, in the past, when performing constant-speed shifts such as clutch-to-clutch shift or direct pressure control, for example, it is not possible to cope with an increase in engine speed and the initial hydraulic pressure becomes inadequate, and subsequent engagement pressure control may be delayed. There was a nature.

The present invention has been made in view of the above circumstances, and it is possible to improve the shift control by optimizing the initial hydraulic pressure control in the clutch-to-clutch shift or direct pressure control. An object is to provide a control device. This object is achieved by changing the content of the initial hydraulic pressure control depending on the presence / absence of the constant speed shift.

[0008]

In order to achieve the above object, the invention described in claim 1 disengages a predetermined friction engagement device and engages another friction engagement device. Drive for temporarily increasing the rotational speed during a predetermined downshift by executing initial hydraulic pressure control that raises the hydraulic pressure supplied to the engagement-side frictional engagement device to a predetermined pressure after gearshift output during a gearshift In a control device for an automatic transmission connected to a power source, a shift determining means for determining a shift for engaging and disengaging the two friction engagement devices, and the shift determined by the shift determining means, A constant-velocity shift determination means for determining whether or not the shift is for temporarily increasing the rotation speed of the internal combustion engine, and an initial stage for changing the control content of the initial hydraulic control according to the determination result of the constant-speed shift determination means. With hydraulic control means And it is characterized in Rukoto.

Therefore, according to the first aspect of the invention, when the rotational speed of the driving force source is increased during the downshift, the initial hydraulic pressure control for the engagement side friction engagement device at the start of the gear shift is performed. The initial hydraulic pressure control is performed differently from the case where the increase control of the rotation speed is not performed, and is suitable for the driving state of the driving force source. Therefore, the engagement-side frictional engagement device that achieves the shift is in a standby state in which the rotational speed of the internal combustion engine can be immediately changed, even when the rotational speed increase control of the internal combustion engine is being performed. As a result, so-called clutch-to-clutch shift or direct pressure control can be appropriately executed.

The invention described in claim 2 is the same as claim 1.
The initial hydraulic pressure control means in the above configuration maintains the maintenance time of the hydraulic pressure set at substantially the same time as the shift output or the hydraulic pressure set at substantially the same time as the shift output, and the constant speed shift determination means determines the rotational speed of the driving force source. It is characterized in that when a shift that is temporarily increased is determined, it is increased compared to when a shift other than the shift is determined.

Therefore, in the second aspect of the present invention, the frictional engagement device on the engagement side is set to the state immediately before the engagement in accordance with the increase in the rotation speed of the driving force source, so that the so-called clutch-to-clutch shift is performed. Alternatively, direct pressure control or the like can be appropriately executed.

[0012]

Next, the present invention will be described more specifically with reference to the drawings. First, the overall control system will be described. FIG. 6 is a control system diagram for the engine (driving force source) 1 and the automatic transmission 3, and a signal corresponding to the depression amount of the accelerator pedal 20 is an electronic control unit for the engine. 21 is input. The intake duct of the engine 1 is provided with an electronic throttle valve 23 driven by a throttle actuator 22. The electronic throttle valve 23 is controlled by the engine controller 21 from the throttle actuator according to the depression amount of the accelerator pedal 20. A control signal is output to 22 and the opening is controlled according to the control amount.

Further, an engine rotation speed sensor 24 for detecting the rotation speed of the engine 1, an air flow meter 25 for detecting the intake air amount, an intake air temperature sensor 26 for detecting the temperature of the intake air, and an opening degree of the electronic throttle valve 23. Throttle sensor 27 for detecting θ, vehicle speed sensor 28 for detecting vehicle speed V from the rotational speed of output shaft 17, cooling water temperature sensor 29 for detecting cooling water temperature of engine 1, brake switch 30 for detecting brake operation, shift An operation position sensor 32 that detects the operation position of the lever 31 is provided. From those sensors, engine speed N, intake air temperature T a, electronic throttle valve 2
A signal representing the opening degree θ of 3, the vehicle speed V, the engine cooling water temperature THw, the brake operating state BK, and the operation position Psh of the shift lever 31 is supplied to the engine electronic control unit 21 or the shift electronic control unit 33. It has become. It should be noted that the shift electronic control unit 33 receives the opening θ of the electronic throttle valve 23, the vehicle speed V, the engine cooling water temperature THw, the signal of the brake operating state BK, and the signal of the operating position Psh of the shift lever 31. Has been done.

Further, a signal representing the turbine rotation speed NT is supplied from the turbine rotation speed sensor 34, which detects the rotation speed of the turbine runner, to the electronic shift control device 33. Further, a signal indicating the kick down operation is output from the kick down switch 35 that detects that the accelerator pedal 20 has been operated to the maximum operation position.
3 has been entered. Further, a sports mode switch 39 which is manually operated to output a shift signal and a constant speed shift switch 40 are connected to a shift electronic control unit 33 (for example, Japanese Patent Application Laid-Open No. 6-307527, Japanese Patent Application No. 7-2).
15892).

Here, the sport mode switch is a switch for selecting a mode in which a shift is performed by a manual operation or a switch for outputting a shift signal by a manual operation, and is arranged on a shift device (not shown) or an instrument panel. . The constant speed shift switch is a switch for downshifting one step by manual operation, and is attached to an appropriate position such as a central portion of a steering wheel (not shown). When a downshift is performed by operating these switches, the electronic throttle valve 23 is opened more than the depression amount of the accelerator pedal 20 based on the output signal from the engine electronic control unit 17,
The so-called constant speed shift control is executed to increase the engine speed Ne to the synchronous speed at the gear after the downshift.

The engine electronic control unit 21 is a so-called microcomputer equipped with a central processing unit (CPU), a storage device (RAM, ROM), and an input / output interface. The CPU uses the temporary storage function of the RAM. Meanwhile, the input signal is processed according to a program stored in advance in the ROM, and various engine controls are executed. For example, the fuel injection valve 36 is controlled to control the fuel injection amount,
Controls the igniter 37 for ignition timing control, controls a bypass valve (not shown) for idle speed control, and controls all throttles including traction control.
The throttle actuator 22 controls and executes the electronic throttle valve 23.

The gear shift electronic control unit 33 is also a microcomputer similar to the engine electronic control unit 21 described above, and the CPU uses the temporary storage function of the RAM and receives an input signal according to a program stored in the ROM in advance. As well as processing, each solenoid valve or linear solenoid valve of the hydraulic control circuit 38 is driven.
For example, the electronic shift control device 33 includes a linear solenoid valve SLT for generating an output pressure PSLT having a magnitude corresponding to the opening of the throttle valve 23, a linear solenoid valve SLN for controlling the accumulator back pressure, and The linear solenoid valves SLU are driven to control the slip amount of the lock-up clutch, and to control the engagement pressure of a predetermined clutch or brake during a gear shift transition according to the progress of gear shift and according to the input torque.

Further, the electronic shift control device 33 uses the basic throttle opening θ (throttle opening converted to the depression amount of the accelerator pedal with a predetermined non-linear characteristic), the vehicle speed V, and a shift diagram using these as parameters. The gear position of the automatic transmission 3 and the engagement state of the lockup clutch are determined based on the above, and No. in the hydraulic control circuit 38 is determined so as to obtain the determined gear stage and engagement state. 1 to No. When the solenoid valves SOL1, SOL2 and SOL3 of No. 3 are driven to generate engine braking, No.
4 solenoid valve SOL4.

The automatic transmission 3 in this embodiment is constructed so that it can set 5 forward gears and 1 reverse gear, which is shown in a skeleton diagram as shown in FIG. That is, in FIG. 7, the automatic transmission 3 is connected to the engine 1 via the torque converter 2. The torque converter 2 includes a pump impeller 5 connected to a crankshaft 4 of the engine 1, a turbine runner 7 connected to an input shaft 6 of the automatic transmission 3, and a direct connection between the pump impeller 5 and the turbine runner 7. The lock-up clutch 8 and the stator 10 whose one-way clutch 9 prevents rotation in one direction.

The automatic transmission 3 has two types of high and low.
It is provided with a sub-transmission unit 11 that switches the gears, and a main transmission unit 12 that can switch between a reverse gear and four forward gears. The auxiliary transmission unit 11 includes a sun gear S0, a ring gear R0,
And an HL planetary gear unit 13 composed of a pinion P0 rotatably supported by the carrier K0 and meshed with the sun gear S0 and the ring gear R0, and a clutch C0 and a one-way clutch provided between the sun gear S0 and the carrier K0. F0 and a brake B0 provided between the sun gear S0 and the housing 19 are provided.

The main transmission unit 12 includes a sun gear S1, a ring gear R1, and a first planetary gear unit 14 which is rotatably supported by a carrier K1 and comprises a pinion P1 meshed with the sun gear S1 and the ring gear R1.
A second planetary gear unit 15 including a sun gear S2, a ring gear R2, and a pinion P2 rotatably supported by the carrier K2 and meshed with the sun gear S2 and the ring gear R2, a sun gear S3, and a ring gear R3.
, And a third planetary gear unit 16 comprising a pinion P3 rotatably supported by the carrier K3 and meshed with the sun gear S3 and the ring gear R3.

The sun gear S1 and the sun gear S2 are integrally connected to each other, and the ring gear R1, the carrier K2, and the carrier K3 are integrally connected to each other.
Is connected to the output shaft 17. Also, the ring gear R2
Are integrally connected to the sun gear S3. A first clutch C1 is provided between the ring gear R2 and the sun gear S3 and the intermediate shaft 18, and a second clutch C2 is provided between the sun gear S1 and the sun gear S2 and the intermediate shaft 18.

As the braking means, the housing 19 is provided with a band-type first brake B1 for stopping the rotation of the sun gear S1 and the sun gear S2. Also, the sun gear S1 and the sun gear S2 and the housing 19
Between the first one-way clutch F1 and the brake B
2 are provided in series. This first one-way clutch F
1 is configured to be engaged when the sun gear S1 and the sun gear S2 try to reversely rotate in the direction opposite to the input shaft 6.

A third brake B3 is provided between the carrier K1 and the housing 19, and a fourth brake B4 and a second one-way clutch F2 are provided in parallel between the ring gear R3 and the housing 19. Has been. This second
The one-way clutch F2 is configured to be engaged when the ring gear R3 attempts to rotate in the reverse direction. The clutches C0, C1, C2, the brakes B0, B1, B2
, B3, B4 are hydraulic friction engagement devices in which friction materials are engaged by the action of hydraulic pressure.

In the above automatic transmission, five forward gears and reverse gears can be set, and the engagement / disengagement state of each friction engagement device for setting these gears is shown in FIG. It is shown in the operation table. Note that, in FIG. 8, the mark ◯ indicates the engaged state and the mark x indicates the released state.

FIG. 9 shows the operating position of the shift lever 31. In the figure, a combination of six operation positions in the front-rear direction of the vehicle and two operation positions in the left-right direction of the vehicle allows the shift lever 31 to be operated at eight operation positions by a supporting device (not shown) so that the shift lever 31 can be operated. It is supported. And P is the parking range position, R
Is the reverse range position, N is the neutral range position,
D is a drive range position, "4" is a "4" range position for setting a gear up to the fourth speed, "3" is a "3" range position for setting a gear up to the third speed, and "2" is The "2" range position is used to set the gears up to the second speed, and L is the low range position where the upshift to the first or higher gears is prohibited. A sports mode switch 39 is arranged between the "2" range position and the low range position on the vehicle rear side with respect to these.

As shown in FIG. 8, the automatic transmission 3 described above is
The shift between the second speed and the third speed is a clutch-to-clutch shift in which both the engagement states of the third brake B3 and the second brake B2 are switched. In the shift control, it is necessary to control the friction engagement device involved in the shift to the underlap or overlap state according to the power on / off state and the shift up / down state. It is necessary to control the hydraulic pressure of the second brake B2 in accordance with the input torque, and the hydraulic pressure of the third brake B3 in accordance with the progress of the shift. Therefore, the hydraulic control circuit 38 incorporates the circuit shown in FIG. 10 in order to smoothly and quickly execute this shift, and its configuration will be briefly described below.

In FIG. 10, reference numeral 70 indicates a 1-2 shift valve, reference numeral 71 indicates a 2-3 shift valve, and reference numeral 72 indicates a 3-4 shift valve. The communication state of each port of these shift valves 70, 71, 72 at each shift speed is determined by the respective shift valves 70, 71,
72 is as shown below. The numbers indicate the respective gears. A third brake B3 is connected via an oil passage 75 to a brake port 74 that communicates with the input port 73 at the first speed and the second speed among the ports of the 2-3 shift valve 71. An orifice 76 is interposed in this oil passage, and the orifice 76 and the third brake B
3 and a damper valve 77 is connected. The damper valve 77 sucks a small amount of hydraulic pressure to perform a buffering action when the line pressure is suddenly supplied to the third brake B3.

Reference numeral 78 is a B-3 control valve, and the engagement pressure of the third brake B3 is directly controlled (direct pressure control) by this B-3 control valve 78. That is, the B-3 control valve 78 is equipped with a spool 79, a plunger 80, and a spring 81 interposed therebetween.
The oil passage 75 is connected to the input port 82 which is opened and closed by the output port 83 which is selectively communicated with the input port 82 is connected to the third brake B3.
Further, the output port 83 is connected to a feedback port 84 formed on the distal end side of the spool 79.
On the other hand, among the ports 85 of the 2-3 shift valve 71, a port 86 that outputs the D range pressure at the third or higher speed is provided through a hydraulic passage 87 to the port 85 that opens at the place where the spring 81 is disposed. Are in communication. Also the plunger 80
A lockup clutch linear solenoid valve SLU is connected to a control port 88 formed on the end side of the.

Therefore, the B-3 control valve 78
Is configured such that the pressure regulation level is set by the elastic force of the spring 81 and the hydraulic pressure supplied to the port 85, and the elastic force of the spring 81 increases as the signal pressure supplied to the control port 88 increases. There is.

Further, in FIG. 10, reference numeral 89 is a 2-3 timing valve, and the 2-3 timing valve 89 is
A spool 90 having a small-diameter land and two large-diameter lands, a first plunger 91, a spring 92 arranged between them, and a spool 90 are arranged on the opposite side of the first plunger 91. And a second plunger 93. An oil passage 95 is connected to an intermediate port 94 of the 2-3 timing valve 89.
Is the port 9 of the 2-3 shift valve 71 that is communicated with the brake port 74 at the third or higher speed.
6 is connected.

Further, the oil passage 95 is branched on the way to open a port 97 between the small diameter land and the large diameter land.
Is connected via an orifice. The port 98, which is selectively communicated with the port 94 at the intermediate portion, is the oil passage 9
9 is connected to the solenoid relay valve 100. The lock-up clutch linear solenoid valve SLU is connected to the port opened at the end of the first plunger 91, and the second brake B2 is passed through the orifice at the port opened at the end of the second plunger 93. Connected.

The oil passage 87 is for supplying and discharging hydraulic pressure to and from the second brake B2, and a small diameter orifice 101 and an orifice 102 with a check ball are interposed in the middle thereof. Further, the oil passage 103 branched from the oil passage 87 has a large diameter orifice 10 provided with a check ball that opens when the pressure is exhausted from the second brake B2.
4 is interposed, and this oil passage 103 is connected to an orifice control valve 105 described below.

The orifice control valve 105 is a valve for controlling the exhaust pressure speed from the second brake B2, and the port 107 formed in the intermediate portion so as to be opened and closed by the spool 106 has the second brake B2.
The oil passage 103 is connected to a port 108 formed below the port 107 in the figure. A port 109 formed above the port 107 to which the second brake B2 is connected in the drawing is a port that is selectively communicated with the drain port, and the port 109 is connected to the port 109 via an oil passage 110. The port 111 of the B-3 control valve 78 is connected. The port 111 is a port that is selectively communicated with the output port 83 to which the third brake B3 is connected.

Of the ports of the orifice control valve 105, a control port 112 formed at the end opposite to the spring for pressing the spool 106 is connected to the port 114 of the 3-4 shift valve 72 via the oil passage 113. ing. The port 114 is a port that outputs the signal pressure of the third solenoid valve SOL3 at the third and lower gears and outputs the signal pressure of the fourth solenoid valve SOL4 at the fourth and higher gears. Further, an oil passage 115 branched from the oil passage 95 is connected to the orifice control valve 105, and the oil passage 115 is selectively connected to the drain port.

Incidentally, the port 11 for outputting the D range pressure at the second or lower speed in the 2-3 shift valve 71.
6 through the oil passage 11 at the port 117 opening at the position where the spring 92 is arranged in the 2-3 timing valve 89.
8 are connected. Also 3-4 shift valve 72
Of these, a port 119, which is communicated with the oil passage 87 at a speed lower than the third speed, is connected to the solenoid relay valve 100 via an oil passage 120.

In FIG. 10, reference numeral 121 indicates an accumulator for the second brake B2, the back pressure chamber of which is
An accumulator control pressure regulated according to the hydraulic pressure output from the linear solenoid valve SLN is supplied. The accumulator control pressure is controlled according to the input torque, and the linear solenoid valve S
The lower the LN output pressure, the higher the pressure. Therefore, the transitional hydraulic pressure of the engagement / disengagement of the second brake B2 is changed to a higher pressure as the signal pressure of the linear solenoid valve SLN is lower. Further, by temporarily lowering the signal pressure of the linear solenoid valve SLU, the engagement pressure of the second brake B2 can be temporarily raised.

Reference numeral 122 represents a C-0 exhaust valve, and reference numeral 123 represents an accumulator for the clutch C0. C-0 exhaust valve 1
Numeral 22 operates to engage the clutch C0 to apply the engine brake only in the second speed in the second speed range.

Therefore, according to the hydraulic circuit described above,
If the port 111 of the B-3 control valve 78 communicates with the drain, the engagement pressure of the third brake B3 can be directly regulated by the B-3 control valve 78, and the regulation level can be adjusted by the linear solenoid valve. SLU
Can be changed by If the spool 106 of the orifice control valve 105 is in the position shown in the left half of the figure, the second brake B2 can be communicated with the oil passage 103 through this orifice control valve 105, so that the large diameter orifice 104 is used. Exhaust pressure is possible and therefore the drain speed from the second brake B2 can be controlled.

Second speed and third speed in the automatic transmission 3 described above.
The shift between the high speed and the second speed is the second brake B2 and the third brake B.
The clutch-to-clutch shift is executed by simultaneously switching the engagement / disengagement state with 3 and executing the downshift from the 3rd speed to the 2nd speed, for example, the second brake B2 engaged at the 3rd speed. Is gradually released according to the input rotation speed to cause a rotation change, and the input rotation speed changes toward the second speed synchronous rotation speed, so that when the predetermined rotation speed is reached, the third brake B3 The second engagement speed is achieved by rapidly increasing the engagement pressure.

In this way, the third brake B3 on the engaging side is
It is necessary to immediately engage in accordance with the change in rotation accompanying the progress of gear shift. On the other hand, in the general friction engagement device including the third brake B3, a slight clearance is generated between the friction plates or between the friction plates and the piston of the hydraulic servo mechanism in the released state. It has no torque capacity until is blocked. Therefore, the hydraulic pressure is rapidly supplied to the engagement side frictional engagement device at the same time as or immediately after the shift output of the clutch-to-clutch shift, so that the torque capacity is set to a state immediately before the engagement.
That is, the initial hydraulic pressure control is performed. In that case, the state in which the hydraulic pressure is immediately engaged due to further increase depends on the engine output or the input rotation speed to the automatic transmission, and thus the above-described control device controls as follows.

FIG. 1 is a flow chart for explaining the downshift from the third speed to the second speed by dividing it into three modes. After processing the input signal (step 1),
A downshift from the third speed to the second speed, which is so-called clutch-to-clutch shift, is determined (step 2). Therefore, this step 2 corresponds to the shift determining means of the present invention. If a negative determination is made in step 2, the process returns without performing any control, and if an affirmative determination is made, it is determined whether or not the sport mode is set (step 3).

As described above, the sport mode is a gear shift mode in which gear shifting is executed based on switch operation. As a form of the switch, each gear position is provided in the shift device, and a shift lever is provided at each gear position. Equipped with a switch that can be turned on, or by setting the sport mode state and operating the upshift switch or downshift switch with the shift lever in that state, and further up the steering wheel or instrument panel etc. Some include down switches. Therefore step 3
The determination may be made by determining whether or not signals are output from these switches. Therefore, this step 3 corresponds to the constant velocity shift determination means of the present invention.

When a negative determination is made in step 3 because the downshift is accompanied by a change in the running state, it is determined whether or not the power is on (step 4). That is, it is determined whether the electronic throttle valve 23 is open and the vehicle is being driven by the output of the engine 1, and this can be determined based on the throttle opening θ.

If an affirmative decision is made in step 4 due to the power-on state, a shift is executed by the release control of the second brake B2 and the engagement control of the third brake B3 as described above (step 5). ). This is shown in FIG.
2-3 shift valve 71 shown in FIG. 3 is switched, the linear solenoid valve SLU regulates the engagement pressure of the third brake B3, and accordingly, the second brake B2 is regulated by the accumulator 121 and discharged. It is executed by In this case, since the engine 1 is in the driving state, the engine speed, that is, the input speed NC0 increases due to the decrease in the engagement pressure of the second brake B2 which has set the third speed. Therefore, the engagement pressure PB3 of the third brake B3 for achieving the second speed is controlled according to the pattern I shown in FIG. 2 (step 6).

This will be briefly described. From the third speed to the second speed
At a time point t2 when a predetermined T1 second has elapsed from the time point t1 when the speed change judgment is established, the duty ratio of the linear solenoid valve SLU that determines the pressure regulation level of the third brake pressure PB3 is increased for a predetermined time until the time point t3. , Performs initial hydraulic pressure control to reduce pack clearance. That is, the duty ratio is set to D1 and maintained for T2 seconds. Thereafter, the duty ratio is maintained at the low value D2 until the time t4 when the input speed NC0 increases to a speed smaller by the predetermined speed Δα than the synchronous speed of the second speed, and the third brake pressure PB3 stands by at a low pressure. To be done. And the linear solenoid valve S
The duty ratio of LU is increased stepwise to gradually increase (sweep up) the third brake pressure PB3, and the input speed N
At time t5 when C0 reaches the synchronous speed, the third brake B
Fully engage 3.

By controlling in this way, the third brake B is activated when the engine speed Ne reaches the predetermined speed.
3 starts to engage substantially immediately and the engine speed Ne
However, abrupt changes before and after reaching the second-speed synchronous rotation speed after the downshift are prevented, and the shift shock is improved.

In the case of the downshift in the power-on state described above, the engine torque reduction control by retarding the ignition timing and the release pressure of the second brake B2 on the release side are temporarily set at the end of the shift. Control for increasing the pressure and reducing the output torque is executed.

On the other hand, if the determination in step 4 is negative because the power is off, a downshift in the power off state is executed (step 7) and the engagement pressure of the third brake B3 is increased. Is the pattern shown in FIG.
Controlled according to II (step 8).

That is, the initial hydraulic pressure control for maintaining the state where the duty ratio of the linear solenoid valve SLU is set to D1 for T2 seconds is executed, and after the end time point t3, the input rotational speed NC0 is compared with the synchronous rotational speed of the second speed. The duty ratio of the linear solenoid valve SLU is gradually increased at time t6 when the rotational speed is considerably low, and the engagement pressure PB3 of the third brake B3 is gradually increased. Then, after the time point t7 when the third brake pressure PB3 becomes sufficiently high, the input rotational speed NC0, that is, the engine rotational speed reaches the second speed synchronous rotational speed. Therefore, in the case of downshifting from the third speed to the second speed in the power-off state, the engagement pressure PB3 of the third brake B3 is increased at an early stage to input the rotation speed of the engine 1 from the output shaft side. The torque is increased by the torque to accelerate the speed change, and the input speed NCO is smoothly changed toward the synchronous speed to improve the speed change shock. Further, in this case, since the third brake B3 on the engaging side is made to stand by in a state of immediately engaging according to the input rotation speed, the shift shock does not deteriorate. In this case, since the power is off, neither the retard control of the ignition timing nor the control of temporarily increasing the release pressure of the second brake B2 on the release side to reduce the output torque at the end of the shift is performed.

If the affirmative determination is made in step 3 due to the downshift in the sports mode, the constant speed shift is executed (step 9). This constant speed shift is a shift control in which the engine speed Ne is increased to the synchronous speed at the shift stage after the downshift, and the downshift is executed in this state. The engagement of the second brake B2 and the third brake are performed. This is performed by temporarily opening the electronic throttle valve 23 by the engine electronic control unit 21 in accordance with the engagement of B3. Furthermore, the third brake pressure P
B3 is controlled according to pattern III shown in FIG. 4 (step 10).

To briefly explain this, the duty ratio of the linear solenoid valve SLU in the initial hydraulic pressure control is
The value D1 which is larger than D1 is set on the basis of the value D1 which is almost normal to increase the hydraulic pressure supplied to the third brake B3.
Here, D3 is set based on D1, and when D1 is changed by learning, D3 may be changed in synchronization at a predetermined rate. Therefore, the third brake B3 is
Even if the input rotation speed is increased, the state immediately before the sufficient engagement corresponding thereto is quickly set, and the low pressure standby state is set. After that, the duty ratio is reduced to D2, the third brake pressure PB3 is kept at a low pressure, and the throttle opening is increased, so that the engine speed Ne becomes
At a time point t8 when the rotation speed reaches a predetermined rotation speed Δβ (> Δα) lower than the high speed synchronous rotation speed, the duty ratio is increased stepwise to increase the third brake pressure PB3. That is, this step 10 corresponds to the initial hydraulic pressure control means in this invention.

Therefore, even in the case of a constant speed shift in which the engine speed is increased during gear shifting, the initial hydraulic pressure control is completed earlier than usual, so a delay may occur in the sweep up of the third brake pressure PB3. Not as a result,
Even if the sweep-up is carried out at an early stage or there is almost no low pressure standby period, the shift control can be made satisfactory.

Since the initial hydraulic pressure control is a control in which the friction plate of the friction engagement device is moved to the state immediately before engagement as described above, the input rotational speed is increased by increasing the initial hydraulic pressure supply time. It is also possible to deal with the rising condition of. An example thereof is shown by a broken line in FIG. 4, and is an example in which the state of the predetermined duty ratio D1 continues for T3 (> T2) seconds. Here, T3 is determined on the basis of T2, and when T2 is changed by learning control or the like, it may be changed at a predetermined ratio correspondingly.

When changing the oil pressure of the initial oil pressure control and the duration thereof in accordance with the constant speed shift as described above, the duty ratio D3 and time T3 thereof can be determined as shown in FIG. That is, these values are used as the engine speed N
It is set as a function of the rate of change of e (Ne dot), and the respective coefficients k1 and k2 and constants a and b are corrected (learned after shifting) based on the learning value during low pressure standby. FIG. 5C shows a general tendency with respect to the learning value of the low pressure standby pressure of the constants a and b. That is, as the rate of change of the engine speed increases, the initial hydraulic pressure control time is set longer or the hydraulic pressure is set higher.

By the way, the above-mentioned correction of the initial hydraulic pressure control is
Since the constant speed shift is performed, step 3 in FIG. 1 may be replaced with a step (step 3 ') of determining whether or not the downshift is performed by the constant speed shift switch. Therefore, this step 3'corresponds to the constant velocity shift determination means of the present invention.

In the above embodiment, the third speed to the second speed are used.
Although the description has been given by taking the downshift to the high speed as an example, the present invention is not limited to the above embodiment, and the hydraulic pressure of the device for performing the downshift control to another gear or the friction engagement device is linear. It can also be applied to devices directly controlled by solenoid valves or the like. Therefore, the friction engagement device that is the target of the engagement pressure control may be a friction engagement device other than the third brake described above. Furthermore, the present invention is characterized in that the control content of the engagement pressure of the friction engagement device involved in the gear shift is provided with the control content peculiar to the constant speed shift, and therefore the control content is the same as the above-mentioned figure. The control pattern is not limited to that shown in FIG. 4 and may be appropriately changed as necessary. The present invention can be implemented for an automatic transmission or a control device therefor having a gear train or hydraulic circuit different from the gear train or hydraulic circuit shown in FIGS. 7 and 10. It should be noted that another power output device such as an electric motor may be used as the driving force source instead of the engine.

[0058]

As described above, according to the invention of claim 1, the downshift of clutch-to-clutch shift or direct pressure control is carried out as a constant speed shift for increasing the rotational speed of the driving force source. In this case, the initial hydraulic pressure control of the frictional engagement device on the engagement side is changed to be suitable for the constant speed shift by making the initial hydraulic pressure control different from the normal downshift that does not accompany the increase in the rotation speed of the internal combustion engine. The standby state of the side frictional engagement device is achieved at an early stage, and as a result, the shift control becomes appropriate, and the shift can be performed in the constant speed shift.

Further, particularly in the invention described in claim 2, in the case of downshifting of the constant speed shift, either the pressure value of the initial hydraulic pressure control of the friction engagement device on the engagement side or the duration thereof is in the normal case. Even if the input speed is increased during gear shifting, the frictional engagement device on the engagement side can be brought into the standby state at an early stage, and as a result of the increase in the input speed, a constant speed shift can be achieved. The gear shift of can be made appropriate.

[Brief description of drawings]

FIG. 1 is a flow chart for explaining control contents executed by a control device of the present invention.

FIG. 2 is a time chart showing a control pattern of a third brake pressure at the time of downshifting from the third speed to the second speed in a power-on state.

FIG. 3 is a time chart showing a control pattern of a third brake pressure at the time of downshifting from the third speed to the second speed in a power off state.

FIG. 4 is a time chart showing a control pattern of the third brake pressure at the time of downshifting from the third speed to the second speed in the constant speed shift.

FIG. 5A is a diagram showing the relationship between the duty ratio that determines the initial hydraulic control pressure and the rate of change of engine speed;
(B) is a diagram showing the relationship between the initial hydraulic control time and the rate of change of the engine speed, and (C) shows the change tendency of the duty ratio or the constant used for correction of the initial hydraulic control time with respect to the low pressure standby learning value. It is a diagram showing.

FIG. 6 is a diagram showing an overall control system according to the present invention.

FIG. 7 is a skeleton diagram showing an example of a gear train of an automatic transmission targeted by the present invention.

FIG. 8 is a diagram showing an engagement operation table of a friction engagement device for setting each shift speed in the automatic transmission.

FIG. 9 is a diagram showing an array at each range position in the shift device.

FIG. 10 is a diagram showing a part of a hydraulic circuit for controlling an engagement pressure when executing a shift between the second speed and the third speed.

[Explanation of symbols]

 1 Engine 3 Automatic Transmission 21 Electronic Control Device for Engine 33 Electronic Control Device for Automatic Transmission 38 Hydraulic Control Device

Claims (2)

[Claims] 1. A hydraulic pressure to be supplied to an engagement side friction engagement device at the time of a gear shift in which a predetermined friction engagement device is released and another friction engagement device is engaged, is predetermined after a gear shift output. In an automatic transmission control device connected to a driving force source that performs initial hydraulic pressure control to raise the pressure to a predetermined value and temporarily increases the rotation speed during a predetermined downshift, the engagement of the two friction engagement devices And a constant speed shift determining means for determining whether or not the shift determined by the shift determining means is a shift for temporarily increasing the rotational speed of the internal combustion engine. And an initial hydraulic control means for changing the control content of the initial hydraulic control according to the determination result of the constant speed shift determination means. 2. The constant speed shift determining means sets the maintenance time of the hydraulic pressure set at substantially the same time as the shift output or the hydraulic pressure set at substantially the same time as the shift output, in the initial hydraulic pressure control means, The control device for the automatic transmission according to claim 1, wherein, when a shift for temporarily increasing the rotation speed is determined, the rotation speed is increased compared to a case where a shift other than the shift is determined.