JPH09296861A - Controller of automatic transmission - Google Patents

Abstract

PROBLEM TO BE SOLVED: To reduce a shift shock by setting the torque capacity of a friction engager being related to the shift, thereby engaging an overlap control means with it, and lowering the torque capacity of a high speed side friction engager along with a shift progress. SOLUTION: In time of gear shifting, an engine electronic control unit compares a car speed at a point of time when a shift judgment is materialized, with a reference speed. In brief, it is compared with the reference speed with a rotational speed corresponding to the speed of an automatic transmission, and when it is less than the reference speed, low speed time control takes place. That is, a sweep-down condition is set up, and a duty ratio of a linear solenoid valve SLN is made gradually heightened by a fixed value at each prescribed time. With this, hydraulic pressure in a second brake B2 in time of a down-shift ranging from third to second is maintained at high pressure from the start of shifting by way of maintaining the back pressure of an accumulator 121 at the high pressure. Consequently the brakes B2 and B3 have the prescribed torque capacity, whereby overlap shift control is carried out. Thus any shift shock is checked.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a device for controlling a shift of an automatic transmission, and more particularly, a device for controlling a clutch-to-clutch shift for simultaneously engaging and disengaging two friction engagement devices. It is about.

[0002]

2. Description of the Related Art Conventionally, in order to facilitate gear shifting without a shock, a one-way clutch is used, and the one-way clutch is changed by a change in torque accompanying engagement of a friction engagement device such as a multi-plate clutch. The gear shift is executed by engaging or disengaging. With such a configuration, the one-way clutch automatically engages and disengages as the multi-plate clutch engages, so gear shifting is performed without causing shocks due to so-called tie-ups or engine upswing. be able to.

However, there is a case where the one-way clutch is abolished due to a demand for reduction in size and weight of the automatic transmission, and in the automatic transmission having such a structure, a predetermined shift is performed by using two multi-disc clutches and multi-disc brakes at the same time. It is a clutch-to-clutch shift that engages and disengages. When performing this gear shift, one of the friction engagement devices needs to function similarly to the one-way clutch, so that hydraulic pressure control is required according to the progress of the gear shift. An example of such a shift is Japanese Patent Laid-Open No. 6-3.
No. 41525.

[0004]

When performing the downshift in the clutch-to-clutch shift described above, the friction engagement device achieving the high speed stage is released, and
Although the friction engagement device that achieves the low speed stage is engaged, the change range of the input rotation speed differs depending on the vehicle speed and the time allowed for the gear shift differs. It is common to control the engagement device.

That is, in the case of downshifting while traveling at a high vehicle speed, the change range of the input rotational speed is large.
The frictional engagement device that sets the high speed stage is released to bring the torque capacities of the two frictional engagement devices involved in the gear shift to a predetermined value or less, so that the so-called underlap state is set. Therefore, since the load applied to the engine decreases, the input rotation speed increases in the power-on state, and approaches the synchronous rotation speed at the low speed side shift stage. Then, when the input rotational speed approaches the synchronous rotational speed to some extent, the torque capacity of the frictional engagement device on the engaging side, that is, the hydraulic pressure is gradually increased, and finally this is completely engaged to complete the shift. .

In the case of downshifting at a low vehicle speed,
Since the change width of the input speed is small, the time allowed for shifting is short, and the torque capacity of the friction engagement device on the low speed stage side is gradually increased after the release control of both friction engagement devices is once performed as described above. There is no time to do the control. Therefore, in the case of a downshift at a low vehicle speed, while gradually releasing the friction engagement device on the high speed stage side and supplying a hydraulic pressure to the friction engagement device on the low speed side when the torque capacity is maintained to some extent, The torque capacity is gradually increased to a so-called overlap state. Then, as the torque capacity of the low speed side frictional engagement device increases, the input speed is increased toward the synchronous speed at the low speed side gear. In this way, the torque capacities (hydraulic pressures) of the two friction engagement devices involved in gear shifting are coordinated with each other to increase or decrease, so that the input rotational speed finally reaches the synchronous rotational speed of the low-speed gear stage.

When the above-described underlap control is performed due to the downshift at a high vehicle speed, the timing at which the hydraulic pressure of the friction engagement device on the low speed stage side is raised is determined by a predetermined rotational speed such as the input rotational speed. It is performed by predicting the time when the synchronous speed of the low speed side shift speed is reached. This synchronous prediction is
Usually, it is performed by predicting the time to reach the synchronous rotation speed on the assumption that the change gradient of the input rotation speed continues. However, if control is performed to reduce the engine torque during gear shifting in order to reduce the load on the friction engagement device or execute smooth gear shifting, the gradient of change in the input rotational speed changes midway. , The sync prediction goes wrong. In such a case, the timing of the supply of hydraulic pressure to the low-speed-stage frictional engagement device becomes unsuitable for the actual progress of the shift, that is, the low-speed-side frictional engagement device is not engaged. The shift shock could be exacerbated because it was too early or, conversely, too late.

In order to eliminate such an inconvenience, in the case of downshifting underlap, the engine torque reduction control during the shift must be executed at a relatively late time so as not to affect the synchronization prediction. Although it cannot be obtained, if the engine torque is controlled in this way, it becomes impossible to sufficiently reduce the inertial energy immediately before the end of the shift. As a result, the effect of reducing the engine torque cannot be obtained, and a so-called peak torque in which the output torque temporarily increases at the end of the gear shift appears, which causes a problem that the gear shift shock is deteriorated.

When the downshift is a so-called multiple shift, the hydraulic pressure of the friction engagement device for setting the intermediate gear is not sufficiently lowered, and the output torque at the intermediate gear appears and the output torque is increased. A gradual change may be felt, which may cause a shift shock. That is, the hydraulic pressure of the friction engagement device on the release side is usually controlled by controlling the back pressure of the accumulator attached to the release side friction engagement device. Even if there is a difference, it is controlled uniformly. For this reason, in the case of multiple shifts in which the gear is downshifted to a shift position separated by two or more steps, the back pressure of the accumulator is controlled so as to optimize the hydraulic pressure of the friction engagement device that is most likely to affect the shift shock. In the frictional engagement device that sets another intermediate shift speed, the back pressure becomes too high, which causes the output torque at that intermediate speed to appear, which may cause a shift shock.

On the other hand, when performing the above-described overlap control for downshifting at a low vehicle speed, when the torque capacity (hydraulic pressure) of the engagement side (low speed stage side) friction engagement device rises to some extent. When the torque capacity of the friction engagement device on the disengagement side (high speed stage side) is still high, a tendency similar to the internal lock of the automatic transmission occurs, and as a result, the output torque decreases and the so-called vehicle is called. There was a possibility of a feeling of pulling in.

The present invention has been made in view of the above-mentioned circumstances, and the shift shock can be improved by properly controlling the friction engagement device on the disengagement side in clutch-to-clutch shifting. An object is to provide a control device.

[0012]

In order to achieve the above-mentioned object, the invention described in claim 1 releases the first frictional engagement device which sets the shift stage on the high speed side. A control device for an automatic transmission that executes a clutch-to-clutch shift in which a second frictional engagement device that sets a low-speed gear position is engaged with the friction shifter during the clutch-to-clutch shift. Overlap control means for bringing the coupling device into a state of being engaged together with torque capacity and thereafter reducing the torque capacity of the first friction engagement device;
At least one of the elapsed time from the start of the gear shift, the difference between the rotational speed of the predetermined rotating member at the low speed side shift stage and the current rotational speed, and the time until the low speed side shift stage is achieved. A shift state determination means for determining a progress state of the shift based on a difference between the elapsed time and the number of revolutions and a time until the shift stage on the low speed side is achieved reaches a predetermined value. Sweep down control means for gradually reducing the engagement pressure of the first friction engagement device is provided.

Therefore, in the invention of claim 1, in the case of downshifting in clutch-to-clutch shifting, the overlap control means executes shift control in the overlap based on the running state of the vehicle. That is, the torque capacities of the two friction engagement devices involved in the shift are set to a predetermined value and engaged, and the torque capacities of the first friction engagement devices on the high speed stage side are reduced as the shift progresses. With the execution of this gear shift, the gear shift state determination means determines the elapsed time from the start of gear shift, the difference between the rotation speed of the predetermined rotary member at the low speed side shift speed and the current speed, and the low speed side shift speed. The progress status of the shift is determined based on any one of the times until the rotation speed is reached. Then, based on the determination result of the shift state determination means, that is, based on the fact that any one of the elapsed time, the rotational speed difference, and the arrival time reaches a predetermined value, the sweep down control means causes the first release-side The engagement pressure of the friction engagement device is gradually reduced. Therefore, in the invention of claim 1, since the torque capacity of the first friction engagement device on the disengagement side is reduced in accordance with the progress of the shift, both of the two friction engagement devices have torque capacities above a predetermined level. It is possible to prevent a situation such as a decrease in output torque.

According to a second aspect of the present invention, the first friction engagement device that sets the high-speed gear stage is released and the second friction engagement device that sets the low-speed gear stage is engaged. In a control device for an automatic transmission that executes clutch-to-clutch shift, the friction engagement devices are released during the clutch-to-clutch shift, and the torque of the second friction engagement device is then released. Underlap control means for increasing the capacity, the elapsed time from the start of the clutch-to-clutch shift, the difference between the rotational speed of the predetermined rotating member at the low speed side shift stage and the current rotational speed, and the low speed side The shift state determination means for determining the progress of the shift based on at least one of the times until the shift stage is achieved, and the difference between the elapsed time and the rotation speed and the shift stage on the low speed side. Sweep-up control means for temporarily increasing the engagement pressure of the first frictional engagement device when any of the times until it reaches a predetermined value is provided. To do.

Therefore, in the second aspect of the present invention, in the case of the downshift in the clutch-to-clutch shift, the underlap control means executes the underlap shift control based on the running state of the vehicle. That is, the torque capacities of the two friction engagement devices are both reduced to a substantially released state as the shift is started, and thereafter the torque capacities of the second friction engagement devices on the engagement side are increased. With the execution of this gear shift, the gear shift state determination means determines the elapsed time from the start of gear shift, the difference between the rotation speed of the predetermined rotary member at the low speed side shift speed and the current speed, and the low speed side shift speed. The progress status of the shift is determined based on any one of the times until the rotation speed is reached. Then, based on the determination result of the shift state determination means, that is, on the basis of any of the elapsed time, the rotational speed difference, and the arrival time becoming a predetermined value, the sweep-up control means causes the first release-side The engagement pressure of the friction engagement device is temporarily increased. Therefore, according to the second aspect of the invention, the torque capacity of the friction engagement device on the disengagement side is increased in accordance with the progress of the shift. This suppresses the increase in torque in a situation similar to tie-up.
Therefore, for example, even if the input torque (engine torque) reduction control is executed late after the shift starts, the peak torque at the end of the shift can be reduced due to the torque decrease due to a situation similar to tie-up, and shift shock can be prevented. be able to.

According to the invention of claim 3, in addition to the structure according to claim 1 or 2, there is further provided a vehicle speed detecting means for detecting a vehicle speed, and the clutch / clutch based on the detected vehicle speed.
It is characterized by selectively performing a control for engaging each friction engagement device with a predetermined torque capacity or a control for releasing each friction engagement device at the time of a two-clutch shift.

Therefore, in the third aspect of the present invention, one of the control by the underlap control means and the control by the opposite overlap control means is selected based on the vehicle speed detected by the vehicle speed detection means. Is executed.

Further, in addition to the structure described in claim 2, the invention of claim 4 detects that the clutch-to-clutch shift is a multiple shift to a low-speed side shift step separated by two or more steps. And a low pressure control means for controlling the hydraulic pressure at the start of the shift of the first friction engagement device to a lower pressure than the shifts other than the multiple shift when the multiple shift is detected. It is characterized by having.

Therefore, according to the invention of claim 4, when the clutch-to-clutch downshift to the gear position separated by two or more gears is executed by the underlap control, when the gearshift of the disengagement side first friction engagement device is started. Since the oil pressure of is controlled to be low, the oil pressure of the friction engagement device that sets the intermediate stage is controlled to be low accordingly. As a result, it is prevented that an intermediate shift speed is temporarily set in the middle of the multiple shift and an output torque corresponding to the shift ratio appears, and the shift shock is improved.

[0020]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS Next, an embodiment of the present invention will be described. First, an example of the engine E and the automatic transmission A that are the subject of the present invention will be described. FIG. 5 is an overall control system diagram, and the engine E to which the automatic transmission A is connected has its intake pipe. The path 12 has a main throttle valve 13 and a sub-throttle valve 14 located upstream thereof. The main throttle valve 1
Reference numeral 3 is connected to the accelerator pedal 15 and is opened / closed in accordance with the depression amount of the accelerator pedal 15. The sub throttle valve 14 is adapted to be opened and closed by a throttle actuator 16 such as a servo motor.

The throttle actuator 16 is controlled to adjust the opening of the sub-throttle valve 14,
In addition, an engine electronic control unit (E-ECU) 17 for controlling the fuel injection amount and ignition timing of the engine E
Is provided. The electronic control unit 17 includes a central processing unit (CPU) and storage devices (RAM, ROM).
In addition, the electronic control unit 17 mainly includes an input / output interface, and the electronic control unit 17 has engine (E / G) rotation speed N, intake air amount Q,
Intake air temperature, throttle opening, vehicle speed, engine water temperature,
Various signals such as a signal from a brake switch are input.

In the automatic transmission A, the hydraulic control device 18 controls the gear shift and the lockup clutch, the line pressure, or the engagement pressure of a predetermined friction engagement device. The hydraulic control device 18 is configured to be electrically controlled, has first to third shift solenoid valves S1,..., S3 for performing a shift, and has a first for controlling an engine brake state. 4 solenoid valve S4, linear solenoid valve SLT for controlling line pressure, linear solenoid valve SLN for controlling accumulator back pressure, linear for controlling engagement pressure of a lock-up clutch or a predetermined friction engagement device. A solenoid valve SLU is provided.

An electronic control unit (T-ECU) 19 for an automatic transmission for outputting a signal to these solenoid valves to control a shift, a line pressure or an accumulator back pressure.
Is provided. This electronic control unit for automatic transmission 19
Is a central processing unit (CPU) and a storage device (RA
M, ROM) and an input / output interface, and this electronic control unit 19 indicates throttle opening, vehicle speed, engine water temperature, signals from brake switch, and shift position as data for control. Signal, signal from pattern select switch, signal from overdrive switch, clutch C described later
A signal from a C0 sensor that detects the rotational speed of 0, an oil temperature of the automatic transmission, a signal from a manual shift switch, and the like are input.

The electronic control unit 19 for the automatic transmission and the electronic control unit 17 for the engine are connected to each other so that data communication is possible, and the electronic control unit 17 for the engine transfers to the electronic control unit 19 for the automatic transmission. On the other hand, a signal such as the intake air amount per rotation (Q / N) is transmitted, and an instruction signal for each solenoid valve is sent from the automatic transmission electronic control unit 19 to the engine electronic control unit 17. A signal equivalent to, a signal instructing a shift speed, and the like are transmitted.

That is, the electronic control unit 19 for the automatic transmission
Is based on input data and a map stored in advance, and indicates ON / OF of a gear position and a lock-up clutch.
F, or the pressure regulation level of the line pressure or the engagement pressure is determined, and an instruction signal is output to a predetermined solenoid valve based on the result of the determination, and further, a failure determination and control based on the failure are performed. . In addition to controlling the fuel injection amount, the ignition timing, the opening degree of the sub-throttle valve 14, etc. based on the input data, the electronic control unit 17 for the engine also sets the fuel injection amount during the shift in the automatic transmission A. The output torque is temporarily reduced by reducing the amount, changing the ignition timing, or narrowing the opening of the sub-throttle valve 14.

FIG. 6 is a diagram showing an example of a gear train of the above-described automatic transmission A, and in the configuration shown here, it is configured to set five forward gears and one reverse gear. That is, the automatic transmission A shown here is
, A sub transmission unit 21 and a main transmission unit 22. The torque converter 20 includes a lock-up clutch 23
The lock-up clutch 23 has a front cover 25 integrated with the pump impeller 24.
(Hub) 2 integrally mounted with the turbine runner 26
7 is provided. An engine crankshaft (each not shown) is connected to a front cover 25 and an input shaft 28 to which a turbine runner 26 is connected.
Is connected to the carrier 30 of the overdrive planetary gear mechanism 29 that constitutes the subtransmission portion 21.

The carrier 3 in this planetary gear mechanism 29
A multi-plate clutch C0 and a one-way clutch F0 are provided between the first gear 0 and the sun gear 31. The one-way clutch F0 is engaged when the sun gear 31 rotates forward relative to the carrier 30 (rotation in the rotation direction of the input shaft 28). A multi-disc brake B0 for selectively stopping the rotation of the sun gear 31 is provided.
The ring gear 3 which is an output element of the subtransmission portion 21
2 is connected to an intermediate shaft 33 which is an input element of the main transmission unit 22.

Therefore, in the auxiliary transmission portion 21, the entire planetary gear mechanism 29 rotates integrally with the multi-plate clutch C0 or the one-way clutch F0 in the engaged state, so that the intermediate shaft 33 has the same speed as the input shaft 28. At low speed. When the brake B0 is engaged and the rotation of the sun gear 31 is stopped, the speed of the ring gear 32 is increased with respect to the input shaft 28 and the ring gear 32 is rotated forward, so that a high gear is established.

On the other hand, the main transmission section 22 has three sets of planetary gear mechanisms 40, 50, and 60, and their rotating elements are connected as follows. That is, the sun gear 41 of the first planetary gear mechanism 40 and the sun gear 51 of the second planetary gear mechanism 50 are integrally connected to each other, and the ring gear 43 of the first planetary gear mechanism 40 and the carrier 52 of the second planetary gear mechanism 50 are connected to each other. The three members of the third planetary gear mechanism 60 and the carrier 62 are connected, and the output shaft 65 is connected to the carrier 62. Further, the ring gear 53 of the second planetary gear mechanism 50 is connected to the sun gear 61 of the third planetary gear mechanism 60.

In the gear train of the main transmission section 22, a reverse gear and four forward gears can be set, and clutches and brakes for this are provided as follows. First, the clutch will be described. The ring gear 53 of the second planetary gear mechanism 50 and the third gear
A first clutch C1 is provided between the sun gear 61 of the planetary gear mechanism 60 and the intermediate shaft 33, and the sun gear 41 of the first planetary gear mechanism 40 and the sun gear 51 of the second planetary gear mechanism 50 and the intermediate shaft are connected to each other. A second clutch C2 is provided between the clutch 33 and the second clutch C2.

Next, the brake will be described. The first brake B1 is a band brake, and the sun gears 41 and 5 of the first planetary gear mechanism 40 and the second planetary gear mechanism 50.
It is arranged to stop the rotation of 1. A first one-way clutch F1 and a second brake B2, which is a multi-plate brake, are arranged in series between the sun gears 41 and 51 (that is, the common sun gear shaft) and the casing 66. One-way clutch F1 has sun gears 41 and 5
1 is engaged when it is about to rotate in the reverse direction (rotation in the direction opposite to the rotation direction of the input shaft 28).

A third brake B3, which is a multi-disc brake, is provided between the carrier 42 of the first planetary gear mechanism 40 and the casing 66. As a brake for stopping the rotation of the ring gear 63 of the third planetary gear mechanism 60, a fourth brake B4, which is a multi-plate brake, and a second one-way clutch F2.
Are arranged in parallel with the casing 66. The second one-way clutch F2 is adapted to be engaged when the ring gear 63 is about to rotate in the reverse direction.

In the above-described automatic transmission A, it is possible to set five forward speeds and one reverse speed by engaging and releasing each clutch and brake as shown in the operation table of FIG. In FIG. 7, a circle indicates an engaged state, a circle indicates an engaged state during engine braking, a triangle indicates either engaged or disengaged, and a blank indicates a disengaged state.

As shown in FIG. 7, 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 vehicle speed, the power on / off state, and the shift up / down state. , It is necessary to control the hydraulic pressures of the second brake B2 and the third brake B3 based on the input torque and the progress of the shift. Therefore, the above hydraulic control device 18 incorporates the circuit shown in FIG. 8 in order to smoothly and quickly execute this shift, and the configuration thereof will be briefly described below.

In FIG. 8, 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, 7
2 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 B3
A damper valve 77 is connected between and. 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 denotes a B-3 control valve which directly controls the engagement pressure of the third brake B3 by the B-3 control valve 78. That is, the B-3 control valve 78 includes a spool 79, a plunger 80, and a spring 81 interposed therebetween, and an oil passage 75 is connected to an input port 82 opened and closed by the spool 79. Output port 8 that is selectively communicated with input port 82
3 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. A lockup clutch linear solenoid valve SLU is connected to a control port 88 formed on the end side of the plunger 80.

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. 8, reference numeral 89 is a 2-3 timing valve. The 2-3 timing valve 89 comprises a spool 90 having a small-diameter land and two large-diameter lands and a first plunger. 91, a spring 92 arranged between them, and a second plunger 93 arranged on the opposite side of the first plunger 91 with the spool 90 interposed therebetween. 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 / discharging hydraulic pressure to / 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. This port 114 outputs the signal pressure of the third solenoid valve S3 at the shift speed of the third speed or lower, and the fourth speed.
It is a port for outputting the signal pressure of the fourth solenoid valve S4 at a shift speed higher than the high speed. 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.

The port 11 for outputting the D range pressure at the shift speed of the second speed or lower 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. 8, reference numeral 121 designates an accumulator for the second brake B2, the back pressure chamber of which is supplied with accumulator control pressure adjusted according to the hydraulic pressure output by the linear solenoid valve SLN. There is. The accumulator control pressure is controlled according to the input torque, and the linear solenoid valve SLN
The lower the output pressure of, 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 SLN, the engaging pressure of the second brake B2 can be temporarily raised.

Reference numeral 122 indicates a C-0 exhaust valve, and reference numeral 123 indicates 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.

In the case of the downshift to the second speed which is executed by engaging and releasing the second brake B2 and the third brake B3, the shock due to the decrease in the output torque due to the tie-up and the large peak torque are generated. In order to prevent the above, and to prevent the engine from overshooting, the hydraulic pressure of the friction engagement device involved in gear shifting is controlled as follows. FIG. 1 is a schematic flowchart for explaining the control of downshifting from the third speed to the second speed as an example, and FIG. 2 is a time chart showing changes in the control value and the output speed.

The shift from the third speed to the second speed is judged by the change of the running state of the vehicle such as the increase of the throttle opening (time t0 in FIG. 2). The shift control is started at time t1 when a predetermined time T1 has elapsed from time t0.
A predetermined duty ratio D32 is output to the linear solenoid valve SLN for controlling the back pressure of the accumulator (step 1). This duty ratio D32 is a value read from a predetermined map, and as shown in FIG. 3, the higher the vehicle speed V, the greater the value set.
As described above, the back pressure of the accumulator is set by the control valve (not shown) that regulates the output pressure of the linear solenoid valve SLN as the signal pressure. The higher the signal pressure, the lower the pressure is set. The higher the vehicle speed, the lower the back pressure of the accumulator.

Therefore, when downshifting from the third speed to the second speed, the second brake B2 which sets the third speed is set.
If the vehicle oil pressure is high, the accumulator 12
By controlling the back pressure of 1 to be low, it is controlled to be low, and conversely, if the vehicle speed is low, it is controlled to be higher than that at high vehicle speed. Therefore, in the case of a high vehicle speed, the hydraulic pressure of the second brake B2 decreases before the hydraulic pressure (torque capacity) of the third brake B3 that sets the second speed increases, so that a so-called underlap control is performed. On the other hand, when the vehicle speed is low, the second brake B2 that sets the third speed is set.
Since the oil pressure is maintained at a certain torque capacity, it is controlled to overlap. Therefore, this step 1 corresponds to the overlap control means of claim 1 and the underlap control means of claim 2.

After the shift control is started as described above,
It is determined whether or not the vehicle speed V at the time when the shift determination is established is greater than or equal to each of two predetermined reference vehicle speeds α and β (α <β) (step 2). It should be noted that this determination is made based on the number of revolutions corresponding to the vehicle speed such as the output number of revolutions of the automatic transmission. And if an affirmative decision is made in step 2,
The vehicle speed is considerably high, and the difference between the engine speed at the time of gear shift determination and the engine speed after gear shift is large.
The control at high vehicle speed is executed. The time chart in this case is shown in FIG.

That is, first, it is judged whether the sweep-up condition is satisfied (step 3). This sweep-up condition is a start condition for gradually increasing the hydraulic pressure of the second brake B2, which is the disengagement side frictional engagement device, and can be specifically determined by the satisfaction of any of the following conditions. One of them is that the difference between the current input speed NC0 and the input speed at the second speed (NO x second speed gear ratio ρ2: NO is the output shaft speed) is smaller than a predetermined value ΔN32D. That is (NC0> NO.rho.2-.DELTA.N32D). Second,
The time between reaching the synchronous speed of the second speed, that is, the second
It means that the time until the speed is achieved is less than the predetermined time. This is determined by the establishment of the following equation.

{(NO .rho.2-NC0) /. DELTA.NC0} .ltoreq..DELTA.NTD Here, .DELTA.NC0 is the amount of change in the input speed during one execution of the routine shown in FIG. 1, and .DELTA.NTD is a predetermined value Is. Furthermore, the third is the elapsed time TD from the start of shifting.
Has reached a predetermined value. In addition, this elapsed time is
Even if it is replaced with the time from when the shift determination is established, the control is substantially the same. Therefore, this step 3 corresponds to the shift state determination means of claim 2.

If a negative determination is made in step 3, the process returns. That is, the output of the duty ratio and the counting of time are continued. On the other hand, if the affirmative judgment is made in step 3 (at time t2 in FIG. 2), the sweep-up control is executed (step 4). Specifically, this is a control for gradually lowering the duty ratio of the linear solenoid valve SLN for controlling the back pressure (accumulation back pressure) of the accumulator, and a constant value Δ
The duty ratio is reduced by D32d. This constant value ΔD
32d is a value prepared in advance in the form of a map, and is set to a larger value as the vehicle speed increases. This step 4 corresponds to the sweep-up control means of claim 2.

This control is continued until the end condition is satisfied (until an affirmative decision is made in step 5). This termination condition is
The state immediately before the input speed NC0 reaches the synchronous speed of the second speed is determined by a known method based on the input speed and the output speed, the speed ratio of the second speed and a predetermined constant,
Second, a certain period of time has passed since the decision was made,
It is possible to adopt that it is determined that a shift to a shift stage other than the high speed has been made, that the elapsed time from the start of shift has reached a specified value (that the guard timer has ended), and the like.

Therefore, when the duty ratio of the linear solenoid valve SLN is controlled as described above, the hydraulic pressure of the second brake B2, which is the friction engagement device on the release side at the time of downshifting from the third speed to the second speed, becomes Due to the increase in the back pressure of the accumulator 121, the pressure is maintained at a certain level at the end of the shift, and then the orifice control valve 105 shown in FIG. 8 is switched to rapidly discharge the pressure from the second brake B2. As a result, when the downshift is performed from the third speed to the second speed, the vehicle speed is high to some extent at the start of the gear shift, so that the second brake B2 on the release side is released.
The hydraulic pressure is set low so that the second brake B2 and the third brake B3 both have a low torque capacity, that is, the so-called underlap state, and the third brake B3 has a certain torque capacity. Brake B2
Since the hydraulic pressure of is increased, the torque capacities of the two brakes involved in the gear shift are both increased, so that the rapid increase of the output torque is suppressed and the gear shift shock is improved.

On the other hand, if the vehicle speed V is lower than the reference vehicle speeds α and β and a negative determination is made in step 2, the process proceeds to step 6 to determine whether the vehicle speed V is lower than one of the reference vehicle speeds α. Is judged. If an affirmative determination is made here, the vehicle speed is considerably low and the difference from the synchronous rotation speed (rotation speed after gear shift) is small, and the control is performed as shown in the time chart of FIG. Done. That is, it is determined whether or not the sweep down condition is satisfied (step 7). This sweep-down condition is a condition for gradually reducing the hydraulic pressure of the second brake B2, which is the disengagement side frictional engagement device, and can be specifically determined by the satisfaction of any of the following conditions. One of them is that the difference between the current input rotational speed NC0 and the input rotational speed at the second speed (NO x second speed gear ratio ρ2: NO is the output shaft rotational speed) is smaller than a predetermined value ΔN32U. (NC0> NO · ρ2 −Δ
N32U). Secondly, the time required to reach the synchronous speed of the second speed, that is, the time until the second speed is achieved is less than a predetermined time. This is determined by the establishment of the following equation.

{(NO .rho.2-NC0) /. DELTA.NC0} .ltoreq..DELTA.NTU where .DELTA.NC0 is the amount of change in the input speed during one execution of the routine shown in FIG. 1, and .DELTA.NTU is a predetermined value. Is. Furthermore, the third is the elapsed time TU from the start of shifting.
Has reached a predetermined value. In addition, this elapsed time is
Even if it is replaced with the time from when the shift determination is established, the control is substantially the same.

Steps 2 and 6 correspond to the vehicle speed determining means in claim 3, and step 7 corresponds to the shift state determining means in claim 1.

If a negative determination is made in step 7, the process returns. That is, the output of the duty ratio and the counting of time are continued. On the other hand, if an affirmative decision is made in step 7 (time t12 in FIG. 2), the sweep down control is executed (step 8). Specifically, this is a control in which the duty ratio of the linear solenoid valve SLN that controls the back pressure of the accumulator is gradually increased, and the duty ratio is increased by a constant value ΔD32u every predetermined time. This constant value ΔD32u is a value prepared in advance in the form of a map according to the vehicle speed. The step 8 corresponds to the sweep-down control means in claim 1.

This control is continued until the end condition is satisfied (until an affirmative decision is made in step 9). This termination condition is
This is the same as the ending condition determined in step 5 described above, and the state immediately before the input speed NC0 reaches the synchronous speed of the second speed is set to the input speed and the output speed, the speed ratio of the second speed, and the predetermined speed. Judgment by a known method based on a constant and that a certain time has elapsed from the time when the judgment was established,
It is possible to adopt that it is determined that a shift to a shift stage other than the second speed, that the elapsed time from the start of shift has reached a specified value (that the guard timer has ended), or the like.

If a negative decision is made in step 6, the same end condition as in step 9 is decided (step 1
0), and step 1 until the end condition is satisfied.
Control continues.

Therefore, if the duty ratio of the linear solenoid valve SLN is controlled as described above, the hydraulic pressure of the second brake B2, which is the disengagement side friction engagement device at the time of downshifting from the third speed to the second speed, becomes By maintaining the back pressure of the accumulator 121 at a high pressure, a high hydraulic pressure is maintained from the beginning of the gear shift, and as a result, the third brake B3, which is the engagement side friction engagement device, has a certain torque capacity. However, the brakes B2 and B3 of both brakes perform a so-called overlap shift control having a predetermined torque capacity. When the above-mentioned sweep-down condition is satisfied, that is, the hydraulic pressure of the second brake B2, which is the disengagement side frictional engagement device, is gradually reduced according to the progress of the shift. Therefore, the two brakes B2 and B3 involved in the gear shift are both released to cause an engine overshoot, or conversely, the two brakes B2 and B3 are engaged together and the output torque sharply decreases. This can be prevented in advance.

Although FIG. 1 does not show the control when the vehicle speed V is between the two reference vehicle speeds α and β, when the vehicle speed is in this region, the control is performed as follows. That is, when the vehicle speed is between these reference vehicle speeds α and β, the two brakes B2 and B3 involved in gear shifting are controlled to be so-called underlaps, but the difference between the synchronous rotational speeds and the rotational speeds shown in FIG. It is not as large as in the case shown in (A), and it is unlikely that the peak torque during synchronization will be extremely large. Therefore, in this case, the shock can be avoided by reducing the peak torque at the time of synchronization by the engine torque reduction control. Therefore, when the vehicle speed V is between the two reference vehicle speeds α and β (α ≦ V <β), the duty ratio D32 of the linear solenoid valve SLN is maintained at a constant value from the start to the end of the shift control. . Its duty ratio D
32 is an intermediate value between the value shown in FIG. 2A and the value shown in FIG. 2B, and is set to a value that does not cause shock due to engine overshoot or tie-up. At the end of the shift, engine torque reduction control is also executed.

Thus, by dividing the vehicle speed into three and controlling the hydraulic pressure in accordance with each, the downshift to the second speed, which is clutch-to-clutch shift, can be performed without causing shock or engine overshoot. Although it can be implemented while avoiding it, in the case where accurate control is possible without causing a response delay in control, in practice, control at the so-called intermediate vehicle speed (α ≦ V <β) is performed. Without doing, the vehicle speed V
May be divided into two regions with one of the above-described reference vehicle speeds α, and any of the controls shown in FIGS. 2A and 2B may be performed.

As shown in FIG. 7, the second brake B2 is the third
Since the gears are engaged at a shift speed higher than the first speed, the clutch / clutch can be engaged even when downshifting from the fourth speed or the fifth speed to the second speed.
Two-to-clutch speed change. In that case, the hydraulic pressure of the frictional engagement device on the disengagement side is set to be lower than the hydraulic pressure after that until synchronizing with the rotation speed of the third speed. Explaining this concretely, FIG. 4 is a flowchart showing a hydraulic pressure control routine in a multiple shift from the fifth speed to the second speed. First, a downshift from the fifth speed to the second speed is determined. (Step 2
0), the process returns if a negative determination is made, and D52 is output as the duty ratio of the linear solenoid valve SLN if a positive determination is made (step 21). This duty ratio D52 is a value read from a predetermined map, and as shown in FIG. 3, the vehicle speed V is set to a higher value as the vehicle speed becomes higher, and the downshift from the third speed to the second speed. It is set to a value larger than the above-mentioned duty ratio D32 output at the time of shifting. Then, as described above, the back pressure of the accumulator is set by the control valve (not shown) that regulates the output pressure of the linear solenoid valve SLN as the signal pressure. The higher the signal pressure, the lower the pressure. The back pressure of the accumulator becomes lower as the vehicle speed increases, and the hydraulic pressure is controlled to be lower than that in the case of downshifting from the third speed. Therefore, step 20 corresponds to the multiple shift determination means of claim 4, and step 21 corresponds to the low pressure control means.

By starting the downshift from the 5th speed to the 2nd speed in this way, the torque capacity of the second brake B2 changes so that other frictions such as the second clutch C2 and the brake B0 of the auxiliary transmission portion 21 are generated. It becomes a value adapted to the torque capacity of the engagement device, that is, it can be in the underlap state as a whole, and an output torque corresponding to the gear ratio appears when synchronizing the fourth speed and the third speed which are intermediate stages. There is no.

After the duty ratio is set as described above and the gear shift is started, it is judged whether or not the third speed is synchronized (step 22). This can be determined by comparing the value obtained by multiplying the output speed by the gear ratio of the third speed and the input speed. When a negative determination is made in step 22, the process returns, and when an affirmative determination is made, on the contrary, the duty ratio is reduced to the duty ratio D32 in the case of downshifting from the third speed to the second speed. (Step 2
3). That is, the back pressure of the accumulator is made higher than before. The control thereafter is as described with reference to FIG. 1, and the sweep up control or the sweep down control is executed based on this duty ratio.

Therefore, even when the second brake B2 for setting the third speed is constructed to have a relatively large torque capacity, the multiple speed change (jumping over) including the third speed as an intermediate stage. In this case, the torque capacity of the second brake B2 can be maintained at a low level during the gear shift), so that an output torque corresponding to the gear ratio of the intermediate gear appears during the gear shift to perform the second gear or the third gear. It is possible to prevent such inconvenience.

The present invention is not limited to the specific examples described above, and can be applied to the control of hydraulic pressure during clutch-to-clutch shifting other than downshifting to the second speed. Further, the control device of the present invention,
The present invention can be applied to an automatic transmission having a gear train other than the gear train shown in FIG. 6 or an automatic transmission having a hydraulic circuit other than the hydraulic circuit shown in FIG.
Therefore, the device for executing the sweep-up control or the sweep-down control may use the linear solenoid valve SLT for line pressure control in addition to the above-mentioned linear solenoid valve SLN, or the other device may control the linear solenoid valve SLT. You may comprise.

[0070]

As described above, according to the invention described in claim 1, the clutch-to-clutch is in the overlap state in which the torque capacities of the plurality of friction engagement devices involved in gear shifting are maintained at a predetermined value or more. When performing a gear shift, the torque capacity of the disengagement side frictional engagement device, that is, the hydraulic pressure, is gradually reduced according to the progress of the gear shift, so that a plurality of frictional engagement devices become engaged and the output torque is greatly reduced. However, inconveniences such as an increase in shock caused by that can be eliminated in advance.

According to the second aspect of the invention, when the clutch-to-clutch shift is executed in the so-called underlap state, the torque capacity of the disengagement side frictional engagement device is determined at the end of the shift based on the progress of the shift. That is, since the hydraulic pressure is temporarily increased, and the torque is lowered in a state close to what is called tie-up, the control of reducing the input torque such as the engine torque is insufficient or the control of reducing the input torque is not particularly performed. However, it is possible to suppress the temporary protrusion of the output torque at the end of the gear shift, that is, the increase of the peak torque, and, finally, it is possible to effectively prevent the gear shift shock as in the case where the input torque reduction control is performed.

Further, according to the third aspect of the invention, the overlap control and the underlap control are selected according to the vehicle speed. Therefore, the control more suitable for the running state of the vehicle becomes possible.

According to the invention of claim 4, in the case of clutch-to-clutch shift in so-called multiple shift,
Since the control characteristic of the hydraulic pressure in the shifting process is made different, the hydraulic pressure can be controlled according to the characteristic of the friction engagement device that sets the intermediate stage, and as a result, the output torque of the intermediate stage appears and It is possible to eliminate inconveniences such as shifting.

[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 change in duty ratio and a change in input rotation speed when the control is executed.

FIG. 3 is a diagram conceptually showing a map of a duty ratio at the start of gear shift.

FIG. 4 is a flowchart for explaining a hydraulic pressure control routine in the case of downshifting to a second speed in multiple shift.

FIG. 5 is a diagram schematically showing an overall control system according to the present invention.

FIG. 6 is a skeleton diagram showing an example of a gear train of an automatic transmission targeted by the present invention.

FIG. 7 is a diagram showing an engagement operation table of a friction engagement device for setting each shift speed in the automatic transmission.

FIG. 8 is a partial hydraulic circuit diagram showing a configuration for mainly controlling hydraulic pressures of a second brake and a third brake.

[Explanation of symbols]

 17 Electronic Control Device for Engine 18 Hydraulic Control Device 19 Electronic Control Device for Automatic Transmission A Automatic Transmission B2 Second Brake B3 Third Brake E Engine

Front Page Continuation (72) Inventor Yasushi Nakamura 1 Toyota-cho, Toyota-shi, Aichi Toyota Motor Co., Ltd.

Claims (4)

[Claims] 1. A clutch-to-clutch that disengages a first friction engagement device that sets a high-speed gear stage and engages a second friction engagement device that sets a low-speed gear stage. In a control device for an automatic transmission that executes a gear shift, in the clutch-to-clutch gear shift, each friction engagement device is brought into a state of being engaged together with a predetermined torque capacity, and then the first friction engagement device. Control means for reducing the torque capacity of the gear, the difference between the rotational speed of the predetermined rotating member at the low speed side shift speed and the current speed, and the low speed side shift speed. A shift state determination means for determining the progress of the shift based on at least one of the times until the above is achieved, and the time until the difference between the elapsed time and the rotational speed and the low-speed gear stage are achieved. And a sweep down control means for gradually lowering the engagement pressure of the first friction engagement device when any one of them becomes a predetermined value. 2. A clutch-to-clutch that disengages a first friction engagement device that sets a high-speed gear stage and engages a second friction engagement device that sets a low-speed gear stage. In an automatic transmission control device for executing a gear shift, an underlap for releasing each friction engagement device during the clutch-to-clutch gear shift and thereafter increasing a torque capacity of the second friction engagement device. The control means, the elapsed time from the start of the clutch-to-clutch shift, the difference between the rotational speed of the predetermined rotary member at the low speed side shift speed and the current rotational speed, and the low speed side shift speed are achieved. A shift state determination means for determining a progress state of the shift based on at least one of the time until the shift, and the time until the difference between the elapsed time and the rotational speed and the low-speed gear stage And a sweep-up control means for temporarily increasing the engagement pressure of the first frictional engagement device when any one of them reaches a predetermined value. apparatus. 3. A control for engaging each friction engagement device with a predetermined torque capacity during the clutch-to-clutch shift based on the detected vehicle speed. The control device for the automatic transmission according to claim 1 or 2, wherein control for releasing each friction engagement device is selectively performed. 4. The clutch-to-clutch shift is 2
A multiplex shift detecting means for detecting a multiplex shift to a low-speed side shift step separated by more than one step, and a multiplex hydraulic pressure at the start of shift of the first friction engagement device when the multiple shift is detected. 3. The control device for an automatic transmission according to claim 2, further comprising: a low pressure control means for controlling the pressure to be lower than that in the case of a shift other than the shift.