JP3399302B2 - Hydraulic control device for automatic transmission for vehicles - Google Patents

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a hydraulic control device for an automatic transmission for a vehicle, and more particularly to a technique for highly accurately determining whether clutch-to-clutch gear shifting has been appropriately performed.

[0002]

2. Description of the Related Art There is known an automatic transmission of a type that achieves a desired one of a plurality of shift speeds by combining operations of a plurality of hydraulic friction engagement devices. In such an automatic transmission, gear shifting is often performed by releasing or engaging a single hydraulic friction engagement device by using a one-way clutch, but the structure becomes complicated. On the other hand, in order to make the automatic transmission smaller and lighter, instead of using the one-way clutch,
By releasing the hydraulic friction engagement device (clutch or brake) and engaging the second hydraulic friction engagement device (clutch or brake), a clutch-to-clutch shift for switching from a predetermined shift stage to another shift stage is performed. The adopted automatic transmission has been proposed. JP-A-9-72409
The device described in Japanese Patent Publication is an example.

[0003]

In such clutch-to-clutch shifting, power ON is downshifted in a state in which the driver depresses the accelerator, that is, in a state in which power is transmitted from the power source side to the drive wheel side. During the downshift, the first hydraulic friction engagement device is quickly released, and the output speed of the power source increases the rotational speed of the input member so that the rotational speed of the input member is substantially synchronized with the rotational speed after the downshift. By engaging the second hydraulic frictional engagement device at times, it is considered to smoothly downshift with excellent responsiveness. FIG. 8 shows that in the vehicle automatic transmission shown in FIGS. 1 and 2, the brake B2 as the first hydraulic friction engagement device is released and the brake B3 as the second hydraulic friction engagement device is engaged. C0 rotation speed N at 3 → 2 downshift of power ON
_C0 (= rotational speed N _IN of the input shaft 20 is input member), a time indicating a change in the control pressure P _SLU of the linear solenoid valve SLU for controlling the hydraulic pressure P _B2, P _B3, B3 hydraulic P _B3 of the brake B2, B3 It is an example of a chart.

Here, if the engagement timing of the brake B3 is early, the tie-up causes the C0 rotation speed N _C0 to rise as indicated by the dotted line, and the engine rotation speed is forcibly increased to cause a gear shift shock. If the combined timing is late, the engine speed will overshoot as indicated by the chain line. Therefore, it has been considered to detect such tie-ups and overshoots to learn and correct the engagement timing of the brake B3. The learning correction of the engagement timing is performed, for example, by rewriting the map of the correction value ΔP _SLU of the control pressure P _SLU during the low pressure standby of the brake B3, and the correction value ΔP _SLU.
The larger the value, the earlier the engagement timing becomes, and the correction amount ΔP
_{If the SLU} is reduced, the engagement timing will be delayed.

By the way, the tie-up is detected by C0.
It was performed depending on whether or not the changing speed (change amount per predetermined time) of the rotation speed N _C0 is a predetermined value or more.
Since the changing speed of the 0 rotation speed N _C0 varies depending on the type of engine, individual difference, etc., it is difficult to determine with high accuracy whether or not a tie-up occurs, and it is not always possible to appropriately learn and correct the engagement timing. could not.

The present invention has been made in view of the above circumstances, and an object of the present invention is to determine whether or not clutch-to-clutch shifting is properly performed, specifically whether tie-up is performed. It is to enable highly accurate determination regardless of individual differences in sources.

[0007]

To achieve the above object, the present invention changes the speed of rotation transmitted from a power source to an input member through a fluid type power transmission device and outputs the rotation from an output member. A hydraulic control device for a vehicular automatic transmission that performs clutch-to-clutch shifting, in which a first hydraulic friction engagement device is released and a second hydraulic friction engagement device is engaged to switch gears. A shift adequacy determining means for determining whether or not the clutch-to-clutch shift is properly performed based on the difference between the rotational speed of the power source and the rotational speed of the input member in the inertia phase during gear shifting. Characterize.

A second aspect of the invention corresponds to an embodiment of the first aspect of the invention, in which the rotation transmitted from the power source to the input member via the fluid type power transmission device is shifted and output from the output member. Disengaging the first hydraulic friction engagement device and engaging the second hydraulic friction engagement device During the power-on downshift of the clutch-to-clutch shift, the first hydraulic friction engagement device is released to release the power. A hydraulic control device for an automatic transmission for a vehicle, comprising clutch-to-clutch shift control means for increasing the rotational speed of the input member by the output of a power source and engaging the second hydraulic friction engagement device at a predetermined timing, When the rotational speed of the input member in the inertia phase during the clutch-to-clutch shifting is higher than the rotational speed of the power source by a predetermined value or more, the second hydraulic friction engagement Wherein the engagement timing of the location has a determined shift appropriateness determination means premature.

[0009]

In such a hydraulic control system for an automatic transmission for a vehicle, whether or not the clutch-to-clutch shift is properly performed based on the difference between the rotational speed of the power source and the rotational speed of the input member. Since the determination is made, the determination can be performed with high accuracy regardless of the type of power source, individual difference, etc., as compared with the case where determination is made based on the change in the rotation speed of the input member as in the related art.

For example, as in the second aspect of the invention, the first hydraulic friction engagement device is released during power-on downshift to increase the rotational speed of the input member by the output of the power source, and the second
When the clutch-to-clutch shift control means for engaging the hydraulic friction engagement device at a predetermined timing is provided, the rotational speed of the power source is usually smaller than the rotational speed of the input member due to the action of the fluid power transmission device. However, if the engagement timing of the second hydraulic friction engagement device is early, the rotational speed of the input member is increased, and the rotational speed of the power source is forcibly increased accordingly. The rotation speed is higher than the rotation speed of the power source. Since such a change in rotation speed due to tie-up occurs solely due to a change in the power transmission direction due to tie-up, regardless of the performance of the power source, the rotation speed of the input member is the same as the rotation speed of the power source. It is possible to determine with high accuracy whether the engagement timing of the second hydraulic frictional engagement device is too early, regardless of the type of power source or individual difference, depending on whether or not the tie-up is greater than the predetermined value.

[0011]

BEST MODE FOR CARRYING OUT THE INVENTION Here, as the power source, an engine operated by combustion of fuel is preferably used, but other power sources such as an electric motor operated by electric energy can also be used. A torque converter is preferably used as the fluid power transmission device, but other means such as a fluid coupling can also be used.

The gear shift adequacy determining means integrates the difference between the rotational speed of the power source and the rotational speed of the input member and divides by time to obtain an average value, thereby making an erroneous determination due to a detection error such as noise. It can be suppressed and higher reliability can be obtained.

According to a second aspect of the present invention, for example, the engagement timing of the second hydraulic friction engagement device by the clutch-to-clutch shift control means is learned so that the engagement timing becomes appropriate according to the determination result of the shift suitability determination means. It is configured to have an engagement timing learning control means for correction.

The gear shift adequacy determining means of the second invention also detects, for example, overshoot of the rotational speed of the input member,
When the overshoot amount is equal to or larger than the predetermined value, it is configured to determine that the engagement timing of the second hydraulic friction engagement device is too late. In that case, when both tie-up and overshoot are detected, it is desirable to prioritize overshoot determination in terms of reliability.

The first hydraulic friction engagement device and the second hydraulic friction engagement device are a multi-plate type or single-plate type friction clutch or brake which is engaged or released by a hydraulic actuator.
Or a band brake or the like.

An embodiment of the present invention will be described below in detail with reference to the drawings. FIG. 1 is a skeleton diagram showing an example of an automatic transmission for a vehicle equipped with a hydraulic control device according to an embodiment of the present invention. In the figure, an engine 10 as a power source
Output of the automatic transmission 1 via the torque converter 12.
4 and is transmitted to the drive wheels via a differential gear unit and an axle (not shown). The torque converter 12 corresponds to a hydraulic power transmission device, and includes a pump impeller 18 connected to a crankshaft 16 of the engine 10, a turbine runner 22 connected to an input shaft 20 of the automatic transmission 14, and the pump impellers. A lock-up clutch 24 that directly connects 18 and the turbine runner 22 and a stator 28 that is prevented from rotating in one direction by a one-way clutch 26 are provided. The input shaft 20 corresponds to an input member.

The automatic transmission 14 is provided with a first transmission 30 for switching between high and low gears and a second transmission 32 for switching between reverse gear and four forward gears. The first transmission 30 includes an HL planetary gear device 34 including a sun gear S0, a ring gear R0, and a planet gear P0 rotatably supported by the carrier K0 and meshed with the sun gear S0 and the ring gear R0.
A clutch C0 and a one-way clutch F0 provided between the sun gear S0 and the carrier K0, and a brake B0 provided between the sun gear S0 and the housing 41 are provided.

The second transmission 32 is a first planetary gear unit 36 including a sun gear S1, a ring gear R1, and a planet gear P1 rotatably supported by the carrier K1 and meshed with the sun gear S1 and the ring gear R1.
, Sun gear S2, ring gear R2, and carrier K
Second planetary gear unit 38, which is rotatably supported by No. 2 and is in mesh with sun gear S2 and ring gear R2, and sun gear S3, ring gear R3, and carrier K3 are rotatably supported by those sun gears. S3 and a third planetary gear set 40 including a planet gear P3 meshed with the ring gear R3.

The sun gear S1 and the sun gear S2 are integrally connected to each other, the ring gear R1, the carrier K2 and the carrier K3 are integrally connected, and the carrier K3 is connected to the output shaft 42 as an output member. Also,
The ring gear R2 is integrally connected to the sun gear S3. A clutch C1 is provided between the ring gear R2 and the sun gear S3 and the intermediate shaft 44, and the sun gear S1
And a clutch C2 between the sun gear S2 and the intermediate shaft 44.
Is provided. Further, a band-type brake B1 for stopping the rotation of the sun gear S1 and the sun gear S2 is provided in the housing 41. Also, the sun gear S1
A one-way clutch F1 and a brake B2 are provided in series between the sun gear S2 and the housing 41. The one-way clutch F1 is configured to be engaged when the sun gear S1 and the sun gear S2 try to rotate in the opposite direction to the input shaft 20.

A brake B3 is provided between the carrier K1 and the housing 41, and a brake B4 and a one-way clutch F2 are provided in parallel between the ring gear R3 and the housing 41. This one-way clutch F
2 is configured to be engaged when the ring gear R3 tries to rotate in the reverse direction.

In the automatic transmission 14 configured as described above, for example, according to the operation table shown in FIG. 2, one of the reverse shift speed and the forward shift speed of 5 speed ratios are sequentially changed. In FIG. 2, “◯” indicates an engaged state, blank indicates a released state, “●” indicates an engaged state at the time of engine braking, and “Δ” indicates an engagement not involved in power transmission. . As is clear from FIG. 2, in the downshift from the third shift speed (3rd) to the second shift speed (2nd), clutch-to-clutch shift is performed in which the brake B2 is released and the brake B3 is engaged. In the power-on downshift of accelerator ON, the underlap shift control is performed in which the brake B2 is released and then the brake B3 is engaged. Other gear changes are performed only by engaging or releasing one clutch or brake. In the present embodiment, the brake B2 is a first hydraulic friction engagement device and the brake B3 is a second hydraulic friction engagement device, both of which are wet multi-plate brakes that are engaged by a hydraulic actuator.

As shown in FIG. 3, the intake pipe of the engine 10 of the vehicle has a first throttle valve 52 and a throttle actuator 54 operated by an accelerator pedal 50.
And a second throttle valve 56 operated by. Further, an engine rotation speed sensor 58 for detecting the rotation speed N _E of the engine 10, an intake air amount sensor 60 for detecting the intake air amount Q of the engine 10, an intake air temperature sensor 62 for detecting the intake air temperature T _A , Throttle sensor 6 for detecting the opening θ _TH of the first throttle valve 52
4, the rotational speed N _OUT namely a vehicle speed sensor 66 for detecting the vehicle speed V of the output shaft 42, a coolant temperature sensor 68 for detecting the cooling water temperature T _W of the engine 10, a brake switch 70 for detecting the operation of the brake, the shift lever 72 The operation position sensor 74 for detecting the operation position P _SH , the rotation speed N _IN of the input shaft 20, that is, the rotation speed N _C0 (= turbine rotation speed N _T ) of the clutch C0, the input shaft rotation speed sensor 7
3. An oil temperature sensor 75 for detecting the hydraulic oil temperature T _OIL of the hydraulic control circuit 84 and the like are provided. From these sensors, the engine speed N _E , the intake air amount Q, the intake air temperature T _A , the first Signals representing throttle valve opening θ _TH , vehicle speed V, engine cooling water temperature T _W , brake operating state BK, shift lever 72 operating position P _SH , input shaft rotational speed N _C0 , hydraulic oil temperature T _OIL are for the engine. It is supplied to the electronic control unit 76 or the electronic control unit 78 for shifting.

The shift lever 72, as shown in FIG. 4, is located in the front-rear direction of the vehicle.
The range, the D and 4 ranges, the 3 range, the 2 range and the L range are operated, and the D range and the 4 range and the 2 range and the L range are operated in the left-right direction of the vehicle. ing.

The engine electronic control unit 76 of FIG.
A so-called microcomputer including a PU, a RAM, a ROM, and an input / output interface, and a CPU is a RAM
The input signal is processed according to a program stored in advance in the ROM while utilizing the temporary storage function of, and various engine controls are executed. For example, the fuel injection valve 80 is controlled to control the fuel injection amount, and the igniter 8 is controlled to control the ignition timing.
2 to control a bypass valve (not shown) for idle speed control, control a second throttle valve 56 by a throttle actuator 54 for traction control, and set an engine speed N _E in an over-rotation range in which the engine speed N _E is set in advance. When it enters (for example, the red zone), the fuel injection valve 80 is shut off to prevent further increase in the engine speed N _E. The engine electronic control unit 76 is communicably connected to a shift electronic control unit 78, and a signal required for one is appropriately transmitted from the other.

The electronic shift control device 78 is also a microcomputer similar to that described above, and the CPU uses the temporary storage function of the RAM while processing the input signal in accordance with the program stored in the ROM in advance, and the hydraulic control circuit 84. Each solenoid valve or linear solenoid valve is driven. For example, the electronic shift control device 78 supplies a command value D _SLT for generating the throttle pressure P _TH having a magnitude corresponding to the opening degree θ _TH of the first throttle valve 52 to the linear solenoid valve SLT, and the command The control pressure P _SLT corresponding to the value D _SLT is output. Further, by supplying a command value D _SLN for controlling the accumulator back pressure P _ACC to the linear solenoid valves SLN, to output a control pressure P _SLN corresponding to the command value D _SLN. Further, in order to control the engagement and release of the lockup clutch 24, the slip amount, the direct control of the brake B3, and the oil pressure P _B3 of the brake B3 during the 2 → 3 shift and the 3 → 2 shift, which are clutch-to-clutch shifts. the command value D _SLU is supplied to the linear solenoid valve SLU, and outputs the control pressure P _SLU corresponding to the command value D _SLU. Electronic control unit 7 for speed change
Reference numeral 8 also determines the gear position of the automatic transmission 14 and the engagement state of the lockup clutch 24 based on the actual throttle valve opening θ _TH and the vehicle speed V from a previously stored gear shift diagram. The solenoid valves S1, S2, S3 are driven so that the shift speed and the engaged state are obtained, and the solenoid valve S4 is driven when the engine brake is generated.

FIG. 5 is a view showing a main part of the hydraulic control circuit 84, which is a 1-2 shift valve 90, a 2-3 shift valve 92, 3
-4 shift valve 94, B2 release valve 96, B3 control valve 98, relay valve 100, and B2 accumulator 102 are arranged and controlled by the solenoid valves S1 to S4 and linear solenoid valves SLU, SLN, SLT, etc. It

The B3 control valve 98 is connected to the brake B3.
Of the hydraulic pressure P _B3 is applied upward in the figure, and the linear solenoid control pressure P _SLU is applied downward in the opposite direction, and the spool 104 that adjusts the hydraulic pressure P _{B3 of the} brake B3 in accordance with these pressures, It is arranged coaxially with the spool 104, and the hydraulic pressure P _{B2 of the} brake B2 is applied in the upward direction in the drawing during the 2 → 3 shift to engage the brake B2 and release the brake B3, and at least during the 2 → 3 shift, a linear solenoid is applied. The plunger 106 to which the control pressure P _SLU is applied downward
The plunger 106 is brought into contact with the spool 104 by the application of the hydraulic pressure P _B2 of the brake B2 so as to interlock with the spool 104. B3 control valve 98
Is supplied with the D range pressure P _D through the 1-2 shift valve 90 which is not switched during the 2 → 3 shift, and the hydraulic pressure P _B3 is regulated using this as the original pressure. A relay valve 100 controlled by the hydraulic pressure P _B2 from the brake B2 is arranged between the B3 control valve 98 and the brake B3.

D connected to a manual shift valve (not shown) that is mechanically switched by the shift lever 72
The range pressure oil passage 108 branches via the 1-2 shift valve 90, and one oil passage 108a is connected to the relay valve 100 via the 2-3 shift valve 92, and the brake B3 via the relay valve 100. Is connected to the oil passage 110. The other branched oil passage 108b is connected to the 3-4 shift valve 94, B2.
It is connected to the import 112 of the B3 control valve 98 via the release valve 96 and the oil passage 108c.
Relay valve 10 from control valve 98 via oil passage 114
It is connected to 0.

Another D range pressure oil passage 116 connected to the manual shift valve branches via the 2-3 shift valve 92, and one oil passage 116a is connected to the oil passage 118 of the brake B2 via the orifice. ing. This oil passage 118
Is connected to the oil passage 116a via the B2 release valve 96, the bypass oil passage 124, and the check valve, and
It is connected to the B2 accumulator 102 via the orifice.

The 3-4 shift valve 94 is connected to the oil passage 108.
In addition to connecting and disconnecting b and 116b, in order to apply the signal pressure P _S3 of the solenoid valve S3 to the spool end of the B2 release valve 96, it is connected to the B2 release valve 96 via the oil passage 120.

The B2 release valve 96 is provided to form a bypass circuit that speeds up the drainage of the hydraulic pressure of the B2 accumulator 102 at the final release of the brake B2.
The spring loaded spool 122,
The signal pressure P _S3 of the solenoid valve S3 via the 4-shift valve 94 is applied to the end of the spool 122, and the bypass oil passage 124 and the oil passage 1
18, communication with and disconnection from the D range pressure oil passage 108b
From the oil passage 108c to the oil passage 108c or to the oil passage 108d connected to the signal port at the end of the plunger 126, and between the oil passage 108e and the oil passage 108c branched from the other D-range pressure oil passage 108a. Shut off.
Therefore, import 11 of B3 control valve 98
2 to 1-2 shift valve 90, 2-3 shift valve 92, B
Oil passages 108a, 108e passing through the two-release valve 96,
And 108c, and 1-2 shift valves 90, 3
-4 shift valve 94, B2 release valve 96 through two passages of oil passages 108b and 108c D range pressure P
_D is supplied.

The B3 control valve 98 opens and closes the import 112 on one side of the two lands provided on the spool 104 by the feedback pressure applied to the end of the spool 104 via the feedback signal pressure import 128, and on the other side the drain port EX. Is opened and closed to adjust the hydraulic pressure P _B3 of the oil passage 114 connected to the outport 130. Further, the plunger 106 disposed coaxially with the spool 104 has a differential piston shape, and has a linear solenoid control pressure P _SLU at the diameter difference portion and 2 at the end face.
The hydraulic pressure P _B2 of the oil passage 118a connected to the oil passage 118 of the brake B2 is applied via the -3 shift valve 92 to have a stroke range in which the oil pressure P _B2 can come into contact with and separate from the spool 104. The B3 control valve 98 is further provided with a plunger 126 for changing the spring load on the spool 104 on the side opposite to the plunger 106.
On the end face of the plunger 126, the oil passages 108d, B
The D range pressure P _D can be applied and released via the 2 release valve 96 and the oil passage 108b.

The relay valve 100 is a spring-loaded spool type switching valve in which the oil pressure P _{B2 of the} oil passage 118 faces the spool end on the spring loaded side and the line pressure P _L faces the other spool end. Is applied to switch the communication between the oil passage 110 of the brake B3 and the oil passages 108a and 114.

Here, the pressure receiving area of the linear solenoid control pressure P _SLU of the B3 control valve 98 is 2 → 2 in comparison with 1 → 2 shift, 2 → 1 shift, and 3 → 2 shift.
Since it increases during the third gear shift, description will be made with reference to the B3 control valve 98 'in FIG. The B3 control valve 98 'is different from the B3 control valve 98 in the way of applying a spring load, but the relationship between the spool 104 and the plunger 106 is substantially the same. As the pressure regulating function of the B3 control valve 98 ', 1 → 2, 2 →
Linear solenoid control pressure P when shifting 1 and 3 → 2
_Adjust the hydraulic pressure P _B3 within the _SLU hydraulic pressure control range, and brake B 3
It is enough to secure the torque capacity of. Therefore, assuming that the pressure receiving area of the end face of the plunger 106 is A ₁ , the pressure receiving area of the diameter difference part is A ₂ , the pressure receiving area of the diameter difference part of the spool 104 is A ₃ , and the pressure receiving area of the land end face is A ₄ , the brake B 3 The hydraulic pressure P _B3 is expressed by the following equation (1). P _B3 = A ₄ × P _SLU / A ₃ (1)

On the other hand, during the 2 → 3 shift, the hydraulic pressure until the hydraulic pressure P _{B2 of the} brake B2 overcomes the return spring force acts on the B3 control valve 98 '.
Since the hydraulic pressure P _B3 must ensure torque capacity of the brake B3, the hydraulic pressure P _B3 of the brake B3 at the time '
Must maintain the torque capacity of the brake B3 by maintaining the relationship of the following equation (2). The pressure receiving area A ₄ ′ of the linear solenoid control pressure P _{SLU in} the equation (2) is (A ₄ + A ₂ ). Further, it is considered that the hydraulic pressure P _B3 of the brake B3 may temporarily drop due to the switching of the circuit during the 2 → 3 shift. Therefore, in order to correct this, it is necessary to set the hydraulic pressure P _B3 high. The pressure receiving area of each part is determined so that the hydraulic pressure P _B3 satisfies the following expression (3). P _B3 ′ = (A ₄ ′ × P _SLU −A ₁ × P _B2 ) / A ₃・ ・ ・ (2) P _B3 ′> P _B3・ ・ ・ (3)

FIG. 7 is a functional block diagram for explaining the main part of the control function of the electronic shift control device 78. In the figure, the automatic gear shift determination means 158 determines the gear shift based on the running state of the vehicle, for example, the actual throttle valve opening θ _TH and the vehicle speed V from a well-known gear shift diagram as a premise for executing the automatic gear shift. To execute.

Clutch-to-clutch shift control means 160
Is determined by the automatic shift determination means 158 from the third shift stage to the third shift stage.
Shift judgment from any of the 5 gears to the 2nd gear
If done, release brake B2 and break
The key B3 is engaged to establish the second gear. Sanawa
The 2-3 shift valve 92 moves from the third gear side to the second gear stage.
When switched to the D side, the D range pressure P_DBrake B2
Supply to the brake B2 and hydraulic oil in the brake B2
Is drained and B2 hydraulic pressure P_B2Is lowered by the break
While B2 is released, the B3 control valve 98
B3 hydraulic pressure P adjusted according to the above equation (1)_B3Is a relay
The brake B3 is supplied from the valve 100 to the brake B3.
3 are engaged. Transient hydraulic pressure P of brake B2_B2Is
Is gradually lowered by the B2 accumulator 102
However, the power that the accelerator pedal 50 is being depressed.
-Accum back pressure P in ON downshift_ACCIs said Lini
A solenoid valve SLN controls the pressure to a low pressure
B2 is released promptly. As a result, the engine 1
C0 rotation speed N together with the input shaft 20 based on the output of 0 _C0
Is raised. Also, the hydraulic pressure P of the brake B3_B3Is
As shown in FIG. 8, C0 rotation speed N_C0Shifts to the second gear
Linear solenoid valve S until just before shifting to the next gear
The low pressure standby state is maintained by the LU, and the C0 rotation speed N_C0
Is the rotational speed of the second gear (N_OUT× i₂)
For example, C0 times so that the brake B3 is engaged at substantially the same time.
Rolling speed N_C0Has reached the specified value or the specified elapsed time
It will be soared after. Figure 8 shows power O
It is a 3 → 2 downshift of N, i₂Is the second gear
Gear ratio, i₃Is the gear ratio of the third gear. C0 above
Rotation speed N_C0Indicates that the clutch C0 is engaged
Downshift from 3rd and 4th gear to 2nd gear
The rotation speed N of the input shaft 20_INMatches

The engagement timing learning control means 162 has a tie-up in which the engagement of the brake B3 is early and the C0 rotation speed N _C0 rises as shown by the dotted line, and the engine rotation speed N _E is forcibly increased. The engagement timing is delayed so that the overshoot (blowing) that rises above the synchronous rotation speed (N _OUT × i ₂ ) along with the engine rotation speed N _E is within a predetermined allowable range as shown by the one-dot chain line.
Engagement timing of the brake B3, i.e. a correction value [Delta] P _{SL U} of the linear solenoid control pressure P _SLU of low-pressure standby, specifically the correction value [Delta] D _SLU command value D _SLU for the linear solenoid valve SLU map etc. rewrite.
If so the correction value [Delta] P _SLU increases engagement timing faster, engagement timing if such correction value [Delta] P _SLU is reduced becomes slow. B3 oil pressure P _B3 is the same as B
The control valve 98 controls the linear solenoid control pressure P _SLU according to the equation (1). However, since there is a predetermined response delay with respect to the change of the linear solenoid control pressure P _SLU , tie-up or overshoot occurs. To prevent this, learning control is necessary. This learning correction is executed when a predetermined learning permission condition is satisfied. In addition,
The timing for raising the linear solenoid control pressure P _SLU from the low pressure standby state may be changed by learning correction.

The shift suitability determination means 164 determines whether or not the tie-up and the overshoot are within a predetermined allowable range, and the C0 rotation speed N _C0 in the inertia phase.
Difference between the engine speed N _E and the engine speed N E32 (= N _C0
-N _E ) is detected, and when the average value NCE32AV is larger than a predetermined determination value LNCE32, it is determined that the tie-up requires the learning correction by the engagement timing learning control means 162. That is, in the power-on downshift, when the clutch B-clutch shift control means 160 releases the brake B2 to increase the C0 rotation speed N _C0 by the output of the engine 10 and the brake B3 is engaged at a predetermined timing, In the inertia phase, the presence of the torque converter 12 normally causes the engine rotational speed N _E to be slightly faster than the C0 rotational speed N _C0 , as shown by the solid line in FIG. 9, but if the engagement timing of the brake B3 is early, the Since the C0 rotation speed N _C0 is increased by the increase and the engine rotation speed N _E is forcibly increased accordingly, the C0 rotation speed N _C0 is the engine rotation speed N _{E as} indicated by the one-dot chain line in FIG. 9. Will be higher than. Also, the amount of overshoot (N _CO −N _OUT ×
i ₂ ) is detected and its maximum value SNFK32 = (N _CO −N
_{When OUT} × i ₂ ) max is larger than a predetermined determination value LFKC032, it is determined that the overshoot requires learning correction by the engagement timing learning control means 162.

On the other hand, in the present embodiment, the C0 rotation speed N _C0 is used instead of the rotation speed N _IN of the input shaft 20, so that the downshift from the fifth shift stage in which the clutch C0 is released to the second shift stage. During the shift, if the engagement timing of the brake B3 is early, the C0 rotation speed N _C0 is decreased as shown by the dotted line in FIG. 9 as the rotation speed of the intermediate shaft 44 increases.
Therefore, in the present embodiment, the rotation difference NEC3 between the engine rotation speed N _E and the C0 rotation speed N _C0 in the inertia phase is generated.
2 (= N _E −N _C0 ) is detected and its average value NEC32A
Even when V is larger than the determination value LNCE32, it is determined that the tie-up requires the learning correction by the engagement timing learning control means 162. Although the same determination value LNCE32 as the determination of the average value NCE32AV is used in the determination of the average value NEC32AV in the present embodiment, different determination values may be set.

FIG. 10 shows determination of tie-up and overshoot during power-on downshift from the third to fifth speeds to the second speed, and brake B3.
In the flowchart for explaining the learning control of the engagement timing of step S10, steps SA2 to SA13 are gear shift suitability determination means 16.
4 and the steps SA4 to SA
7, SA10, SA11, and SA13 are portions related to tie-up determination, and steps SA9 and SA12 are portions related to overshoot determination. Also, step SA
14 and SA15 are engagement timing learning control means 1
Performed by 62. The flowchart of FIG. 10 is repeatedly executed with a predetermined cycle time.

At step SA1 in FIG. 10, the power is turned on.
Whether or not the downshift to the second speed stage at the time is determined based on, for example, the throttle valve opening θ _TH or switching of the solenoid valve for switching the hydraulic pressure control circuit 84, and if it is the 2nd downshift at power ON. Step SA2 is executed. At Step SA2, it is judged whether or not the predetermined monitor period in the latter half of the inertia phase is reached according to the following equation (4),
When the monitoring period is reached, steps SA3 and thereafter are executed.
In the present embodiment, the monitoring period is a period from an intermediate point between the third shift stage and the second shift stage to the end of the shift, of which C0
The rotation speed N _C0 is the synchronous rotation speed (N _OUT × i
_The tie-up monitoring period is until the point ₂ ) is reached, and the overshoot monitoring period is thereafter. N _C0 > N _OUT × K1 (4) where K1 = (i ₂ + i ₃ ) / 2

In step SA3, it is judged whether or not the C0 rotational speed N _C0 has reached the synchronous rotational speed (N _OUT × i ₂ ) of the second shift stage, and step S3 is reached until the synchronous rotational speed is reached.
Execute A4 to SA7. In step SA4, it is determined whether the engine rotation speed N _E is higher than the C0 rotation speed N _C0 . If N _E > N _C0 , the rotation speed difference N is calculated in step SA5.
EC32 = N _E −N _C0 and the rotation difference NCE 32 = 0, while if N _E <N _C0 , the rotation difference NCE is executed in step SA6.
32 = N _C0 −N _E and the rotation difference NEC 32 = 0. Then, in the next step SA7, the previous integrated value SNEC3
2 _i-1 , SNCE32 _i-1, and the current rotation difference NEC32,
NCE32 and the integrated value SNEC32 _i , SNC
E32 _i is calculated. As a result, the integral value S of the rotational difference NEC32 and NCE32 until the C0 rotational speed N _C0 reaches the synchronous rotational speed (N _OUT × i ₂ ) of the second shift stage.
NEC32 and SNCE32 are required.

C0 rotation speed N_C0Is the synchronous rotation of the second gear
Speed (N_OUT× i₂) Is reached and the judgment in step SA3
If YES, step SA8 is executed and the shift is completed.
Whether or not the rotation speed N within a certain time is determined_C0And N_OUT
Value, hydraulic pressure P of brake B3 _B3Judge by the following. So
Then, when it is judged that the shift is completed, step SA10 and below
But if not, ie C0 rotation speed
N_C0Is the synchronous rotation speed (N_OUT× i₂) Over
If it rises, step SA9 is executed and the
Bachute amount (N_CO-N_OUT× i₂) Is detected
The maximum value is stored as SNFK32.

At step SA10, the integrated value SNE is obtained.
By dividing C32 and SNCE32 by the elapsed time until the determination of step SA3 becomes first YES after the determination of step SA2 becomes first, that is, by dividing them by the tie-up monitor period TCE32, respectively. The average values NEC32AV and NCE32AV are calculated. The monitor period TCE32 is obtained by a timer or a counter. In step SA11, the larger one of these average values NEC32AV and NCE32AV is selected as NHK32.

At step SA12, the maximum value SNFK32 of the overshoot amount is set to a predetermined judgment value LF.
Judge whether it is larger than KC032, SNFK32
In the case of> LFKC032, it is determined that the overshoot exceeds the predetermined allowable range, so that the overshoot is reduced in step SA14, in other words, the engagement timing of the brake B3 is advanced, so that the linear solenoid in the low pressure standby state is set. The correction value ΔP _SLU (ΔD _SLU ) is rewritten in the direction in which the control pressure P _SLU increases. SNF
If K32 ≦ LFKC032, the overshoot is within the allowable range, so step SA13 is executed next, and it is determined whether or not the selection value NHK32 is larger than a predetermined determination value LNCE32. And
If NHK32> LNCE32, it is determined that the tie-up exceeds the predetermined allowable range, so that the tie-up is suppressed in step SA15, in other words, the engagement timing of the brake B3 is delayed, so that the low-pressure standby state is achieved. The correction value ΔP _SLU (ΔD _SLU ) is rewritten in the direction in which the linear solenoid control pressure P _SLU decreases.

As described above, in the present embodiment, the rotational speed differences NEC32, NC between the engine rotational speed N _E and the C0 rotational speed N _C0.
Although the tie-up is determined based on E32, the rotation difference due to the tie-up is NEC32 or NCE.
Since the change of 32 is caused solely by the change in the power transmission direction due to the tie-up regardless of the performance of the engine 10, the tie-up is determined with high accuracy regardless of the type or individual difference of the engine 10. it can. This prevents erroneous learning of the engagement timing of the brake B3.

Further, in this embodiment, the rotation difference NEC32,
NCE32 is integrated and its integrated value SNEC3
2. The SNCE32 is divided by the monitor period TCE32 to obtain the average values NEC32AV and NCE32AV, and the tie-up is determined based on the average values NEC32AV and NCE32AV. It can be suppressed and higher reliability can be obtained.

Further, the shift suitability determination means 164 of this embodiment is also provided.
Determines whether or not the overshoot of the C0 rotation speed N _C0 is within the allowable range. This overshoot determination is performed prior to the tie-up determination, and the overshoot exceeds the allowable range. If so, the engagement timing of the brake B3 is learned and corrected based on the overshoot regardless of whether or not the tie-up is performed, so that high reliability is obtained. That is, although the tie-up determination accuracy is improved as compared with the conventional one, its reliability is still lower than that of the overshoot determination, and it is desirable to give priority to the overshoot determination.

Although one embodiment of the present invention has been described in detail with reference to the drawings, the present invention can be implemented in other modes.

For example, in the above embodiment, the power-on downshift to the second speed stage, which is achieved by releasing the brake B2 and engaging the brake B3, has been described. The type of shift is appropriately determined according to the configuration of the automatic transmission.

Further, the flow chart of FIG.
There is no problem even if steps are added or the step contents are changed within the range of achieving the same control function.

Further, in the above-mentioned embodiment, the engine electronic control unit 76 and the shift electronic control unit 78 are constructed independently of each other, but they may be constructed by a common arithmetic control unit.

Although not illustrated one by one, the present invention can be implemented in various modified and improved modes based on the knowledge of those skilled in the art.

[Brief description of drawings]

FIG. 1 is a skeleton diagram illustrating a configuration of an automatic transmission for a vehicle including a hydraulic control device that is an embodiment of the present invention.

FIG. 2 is a diagram showing a relationship between a combination of operations of a plurality of friction engagement devices in the automatic transmission of FIG. 1 and a shift speed established thereby.

3 is a block diagram including a hydraulic control circuit and an electric control circuit for controlling the automatic transmission of FIG.

FIG. 4 is a diagram illustrating an operating position of a shift lever of FIG.

5 is a diagram illustrating a main part of the hydraulic control circuit of FIG.

FIG. 6 is a diagram illustrating a B3 control valve of a hydraulic control circuit.

FIG. 7 is a functional block diagram illustrating a main part of a control function of the electronic shift control device of FIG.

8 is a time chart explaining an example of a control operation by the clutch-to-clutch shift control means of FIG.

FIG. 9 is a time chart illustrating an example of changes in engine rotation speed N _E and C0 rotation speed N _C0 during a power-on downshift of clutch-to-clutch shifting.

10 is a flowchart illustrating a specific control operation of the shift suitability determination means of FIG. 7.

[Explanation of symbols]

10: Engine (Power Source) 12: Torque Converter (Fluid Power Transmission Device) 14: Automatic Transmission 20: Input Shaft (Input Member) 42: Output Shaft (Output Member) 160: Clutch-to-Clutch Shift Control Means 164: Speed Change Suitability determination means B2: Brake (first hydraulic friction engagement device) B3: Brake (second hydraulic friction engagement device) N _E : Engine rotation speed (rotation speed of power source) N _CO : C0 rotation speed (input (Rotation speed of the member)

─────────────────────────────────────────────────── ─── Continuation of the front page (72) Inventor Kagenori Fukumura 1 Toyota-cho, Toyota-shi, Aichi Toyota Motor Corporation (56) References JP-A-8-338517 (JP, A) JP-A-9- 269060 (JP, A) (58) Fields investigated (Int.Cl. ⁷ , DB name) F16H 59/00-61/12 F16H 61/16-61/24 F16H 63/40-63/48

Claims (2)

(57) [Claims] 1. A first hydraulic friction engagement device is released and a second hydraulic system is released while the rotation transmitted from a power source to an input member via a fluid power transmission device is changed in speed and output from an output member. A hydraulic control device for an automatic transmission for a vehicle, which performs a clutch-to-clutch shift in which a friction engagement device is engaged to switch a shift speed, wherein a rotational speed of the power source and an input member in an inertia phase during the clutch-to-clutch shift. A hydraulic control device for an automatic transmission for a vehicle, comprising a shift adequacy determining means for determining whether or not the clutch-to-clutch shift is properly performed based on a difference from the rotational speed of the. 2. The rotation transmitted from the power source to the input member via the fluid power transmission device is shifted and output from the output member, while the first hydraulic friction engagement device is released and the second hydraulic type During the power-on downshift of the clutch-to-clutch shift in which the friction engagement device is engaged, the first hydraulic friction engagement device is released to increase the rotation speed of the input member by the output of the power source and (2) A hydraulic control device for an automatic transmission for a vehicle, comprising clutch-to-clutch shift control means for engaging a hydraulic friction engagement device at a predetermined timing, wherein a rotational speed of the input member in an inertia phase during the clutch-to-clutch shift. Is greater than the rotation speed of the power source by a predetermined value or more, it is determined that the engagement timing of the second hydraulic friction engagement device is too early, and the gear shift adequacy determination hand. Hydraulic control apparatus for a vehicular automatic transmission, characterized in that it comprises a.