JPH11311317A - Hydraulic control device for automatic transmission - Google Patents

Abstract

PROBLEM TO BE SOLVED: To favorably maintain the convergence on a target value and suppress the generation of hunting even when the difference between an actual value and the target value is large. SOLUTION: With this control device, after a specified time tIS for the change in the number of revolutions of an input shaft to be stabilized has elapsed (S44), the deviation E of a target rotative acceleration from an actual rotative acceleration is calculated (S46). At a first cycle (n=0) of operation cycle, proportional control operation mainly functions based on proportional gain K1 and the deviation E (S50). On and after a second cycle (n>=1) of the operating cycle, integral control operation mainly functions based on integrated gain K3 and deviation integrated value (S51).

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 mounted on a motor vehicle, and more particularly to a so-called clutch-to-clutch for engaging one friction engagement element and releasing the other friction engagement element. More particularly, the present invention relates to a hydraulic control device that performs feedback control so that an actual amount of change such as the rotational acceleration of the input shaft speed that changes during the progress of a shift becomes a target value.

[0002]

2. Description of the Related Art Conventionally, a change rate (rotation acceleration) of an input shaft rotation speed is detected, and an engagement side clutch or a release side clutch is operated so as to match the change rate with a target rotation change rate (target rotation acceleration). 2. Description of the Related Art There has been known an apparatus that performs feedback control of a supply pressure to a hydraulic servo.

Further, as disclosed in JP-A-63-266258, the instantaneous value of the input shaft torque of the transmission is detected, and the supply pressure to a hydraulic servo for operating a friction engagement element such as a clutch is detected. Is known in which the torque capacity is adjusted by feedback control. Even in this case, the release-side friction coefficient is controlled so that a change in the rotational speed of the input shaft (actual slip rotational speed change rate) based on the release of the release-side clutch is changed to a predetermined target rotational speed. A drive signal is output to a solenoid valve that adjusts hydraulic pressure to a hydraulic servo that operates the combined element.
Proportional (P) control operation based on the deviation between the actual rotational speed change rate and the target rotational speed change rate and the proportional (P) gain, and integration (I) based on the integral of the deviation (accumulation of the deviation) and the integral (I) gain. Control operation and differentiation of the deviation (deviation change)
And the differential (D) control operation based on the differential (D) gain is taken into account, and the proportional (P) control operation or the integral (I) control operation alone, or the PI control operation in which these proportional control operation and the integral control operation are added Further, the hydraulic pressure correction amount is determined by the PID control operation that also involves the differential control operation, and the hydraulic pressure correction amount is set by appropriately determining the ratio of each gain.

[0004]

In the above-mentioned feedback control, the above-mentioned control operations and the above-mentioned gains are uniquely determined in advance during the shift, so that, for example, the actual rotational speed is determined at the beginning of the shift. When the difference between the change and the target rotation change is large, if the control gain is set to a large value, the convergence to the target will be good, but hunting will easily occur, and as a result, the driver will feel uncomfortable as a shift shock. Will give. Further, if the control gain is set small to prevent hunting, it is not possible to converge, for example, a large difference in rotation change at the beginning of the shift, and the accuracy of feedback control is reduced.

Accordingly, the present invention provides an automatic transmission which solves the above-mentioned problems by maintaining good convergence to the target value and suppressing occurrence of hunting even when the difference between the actual value and the target value is large. It is an object of the present invention to provide a hydraulic control device for a machine.

[0006]

According to a first aspect of the present invention, there is provided an input shaft to which power from an engine output shaft is input;
An output shaft connected to the wheels, a plurality of friction engagement elements for changing a power transmission path between the input shaft and the output shaft, and a hydraulic servo (9, 1) for disconnecting and engaging the friction engagement elements
0), hydraulic pressure control means (SLS, SLU) for controlling the hydraulic pressure of the hydraulic servo, and the hydraulic pressure control means (Sr, SLU) so that the actual change amount (ωr) that changes when the shift progresses matches the target change amount (ω1). A feedback control means (1) for outputting a control signal to the control means, wherein the feedback control means (1)
Is a proportional (P) control operation (1a) based on at least a deviation (E) and a proportional gain (K1) between the target change amount and the actual change amount, and an integral value and an integral gain (K
(I) control operation (1b) based on 3).
In the initial stage of the progression of the shift, the proportional (P) control operation mainly functions, and the proportional control operation is changed so that the integral (I) control operation mainly functions according to the progress of the subsequent shift. A hydraulic control device for an automatic transmission.

According to a second aspect of the present invention, the feedback control means (1) detects the actual change amount in each operation cycle, outputs the control signal, and outputs the control signal for the first time. In the operation cycle (n = 0), the proportional (P) control operation functions, and in the second and subsequent (n ≧ 1) operation cycles, the integral (I) control operation functions (see S48 and S50). ), A hydraulic control device for an automatic transmission according to claim 1.

According to a third aspect of the present invention, the proportional gain (K1) is set smaller as the oil temperature is higher, and the integral gain (K3) is set larger as the oil temperature is higher. Item 1 or 2 is the hydraulic control device for an automatic transmission (see FIG. 9).

According to a fourth aspect of the present invention, the feedback control means (1) determines whether the difference (E) is positive or negative.
The hydraulic control device for an automatic transmission according to any one of claims 1 to 3, wherein the hydraulic pressure correction amount outputs a control signal having a different value (see FIG. 10).

According to a fifth aspect of the present invention, the feedback control means (1) outputs a feedback control signal after a predetermined time (T _IS ) when the progress of the shift is started (see S44). Item 5. The hydraulic control device for an automatic transmission according to any one of Items 1 to 4.

According to a sixth aspect of the present invention, the feedback control means (1) includes a means (ΔT) for correcting a hydraulic pressure with respect to a change in input torque during a gear shift (see S52). The hydraulic control device for an automatic transmission according to any one of the above.

According to a seventh aspect of the present invention, the feedback control means (1) controls the proportional (P) control operation (1).
The hydraulic control device for an automatic transmission according to any one of claims 1 to 6, wherein a derivative (D) control operation (1c) is added to each of a) and the integral (I) control operation (1b).

According to an eighth aspect of the present invention, the amount of change that changes during the progress of the speed change is the rotational acceleration (ωr) of the input shaft speed (N _T ) that changes based on the gear ratio related to the shift to the predetermined speed. The hydraulic control device for an automatic transmission according to any one of claims 1 to 7, wherein:

According to a ninth aspect of the present invention, a first frictional engagement element of the plurality of frictional engagement elements is engaged,
A shift control to a predetermined shift speed is performed by releasing the second frictional engagement element, and the hydraulic pressure control means controls a hydraulic pressure acting on one of the hydraulic servos of the first and second frictional engagement elements. The hydraulic pressure (P _B ) acting on the other hydraulic servo is controlled mainly by (P _A ), and the feedback control means (1) controls the hydraulic pressure (P _A ) acting on the one hydraulic servo acting as the main body. The hydraulic control apparatus for an automatic transmission according to any one of claims 1 to 8, wherein a control signal is output to control the automatic transmission.

According to a tenth aspect of the present invention, the feedback control means determines that the progress of the shift is a predetermined ratio (a1).
Is performed over the inertia phase control in a state where the predetermined ratio is not exceeded and the end control in a state where the predetermined ratio is exceeded.
10. The automatic transmission according to claim 1, wherein in the end control, the target change amount (ω2) is set in consideration of a delay (offset) of an actual change amount (see S56). In the hydraulic control unit of the machine.

[Operation] Based on the above configuration, the rotational acceleration (ω) of the input shaft rotation speed that changes based on the amount of change that changes as the shift progresses, for example, the gear ratio related to the shift to the predetermined gear position.
By detecting r), the hydraulic pressure of the hydraulic servo (for example, the engagement side hydraulic pressure) (P _A ) is feedback-controlled so that the change amount matches the target change amount (for example, the target rotational acceleration) (ω1). . At this time, in the initial stage of the shift, for example, in the first operation cycle, the deviation (E) between the target change amount (ω1) and the actual change amount (ωr) and the proportional gain (K
The proportional (P) control operation (1a) based on 1) mainly functions, and as the shift progresses thereafter, for example, in the second and subsequent calculation cycles, the integral value of the deviation (E) and the integral gain (K3) are increased. The integration (I) control operation (1b) based on functions mainly.

The reference numbers in parentheses are for comparison with the drawings, but do not limit the configuration of the present invention.

[0018]

According to the first aspect of the present invention, at the beginning of a shift in which the difference between the actual change amount and the target change amount is large, the proportional (P) control operation mainly functions to improve the convergence to the target. After that, by performing mainly the integral (I) control operation, smooth control is performed to prevent occurrence of hunting, and highly accurate feedback control can be performed.

According to the second aspect of the present invention, in the first feedback control of the operation cycle, the target is simultaneously converged by the proportional (P) control operation, and in the feedback control of the second and subsequent operation cycles, the integral control is performed. (I) Hunting is prevented by the control operation, and good feedback control can be performed even when there is a large deviation at the beginning of shifting. Further, in the case where the proportional gain is linearly interpolated and reduced in real time as the shift progresses, 2
In the subsequent calculation cycles, it is necessary to set a longer cycle time until the hydraulic pressure based on the previous proportional gain responds. However, as described above, the control operation is changed in the first and second calculation cycles (P → I). By doing so, the proportional gain becomes 0 in the second and subsequent cycles, and the cycle time can be set without considering the proportional control operation, so that highly accurate feedback control can be performed.

According to the third aspect of the present invention, when the oil temperature is high and the response of the control system is good, hunting is likely to occur when the proportional gain is increased. In this case, the proportional gain is reduced and the integral gain is reduced. Hunting can be prevented from occurring by increasing. In addition, when the oil temperature is low and the response of the control system is poor, if the integral gain is increased, the integral control operation becomes a correction amount obtained by multiplying the integral value of the deviation by the integral gain. The hydraulic pressure correction by the gain cannot be controlled in one cycle time, and it is accumulated in each cycle, and when the hydraulic pressure responds, a large amount of correction is applied and large hunting may occur. In this case, by decreasing the integral gain and conversely increasing the proportional gain,
The occurrence of large hunting can be prevented.

According to the fourth aspect of the present invention, in the feedback control, for example, if the direction of decreasing the oil pressure is a direction in which the shift does not proceed, a large oil pressure correction in the direction of decreasing the oil pressure may hinder the progress of the shift. By changing the correction amount of the hydraulic pressure depending on the sign of the deviation, it is possible to prevent the progress of the shift from being hindered by the feedback control.

According to the fifth aspect of the present invention, even if the rotation change due to the shift is started, a predetermined time is required from the start of the rotation change until the amount of rotation change is stabilized, and the oil pressure response By starting the feedback control after the amount of change is stabilized in consideration of the delay, it is possible to perform a good feedback control that prevents the occurrence of hunting and a decrease in convergence.

According to the sixth aspect of the present invention, when an input torque change occurs during gear shifting, if the change is corrected by feedback control, the amount of hydraulic pressure correction by feedback control becomes insufficient, and Although there is a possibility that convergence may deteriorate and hunting or the like may occur, with respect to a change in input torque, by providing a means for correcting the change by feedforward control, even if the change in input torque occurs during shifting. , Good feedback control can be performed.

According to the seventh aspect of the present invention, by adding a differential (D) control operation, hunting can be reduced and responsiveness can be improved, and the performance of the feedback control can be improved. .

According to the present invention, the input shaft rotation speed that changes based on the gear ratio related to the shift to the predetermined gear can be accurately and reliably detected by the input shaft rotation sensor and the vehicle speed sensor. In addition, the rotation acceleration (change rate) can be easily calculated, and accurate feedback control can be performed with a simple configuration.

According to the ninth aspect of the present invention, the so-called clutch-to-clutch shift is accurately performed by feedback-controlling the hydraulic pressure of the main body,
It is possible to prevent the driver or the like from giving an uncomfortable feeling such as a shift shock.

According to the tenth aspect of the present invention, in the final control subsequent to the inertia phase control, the target change amount (ω2) takes into account the actual change delay (offset) due to a hydraulic pressure response delay or the like. Therefore, accurate and reliable feedback control based on the target value can be performed even in the final control.

[0028]

DESCRIPTION OF THE PREFERRED EMBODIMENTS The present automatic transmission has a number of frictional engagement elements such as clutches or brakes, and an automatic transmission in which the transmission path of a planetary gear is selected by appropriately connecting and disconnecting these frictional engagement elements. A mechanism (not shown), wherein the input shaft of the automatic transmission is connected to the engine output shaft via a torque converter, and the output shaft of the automatic transmission is connected to the drive wheels. .

FIG. 1 is a block diagram showing an electric control system. U is a control unit (ECU) composed of a microcomputer (microcomputer), which is an engine rotation sensor 2 and a throttle opening for detecting an accelerator pedal depression amount of a driver. Each signal from the degree sensor 3, the sensor 5 for detecting the input shaft rotation speed (= turbine rotation speed) of the transmission (automatic transmission mechanism), the vehicle speed (= automatic transmission output shaft rotation speed) sensor 6, and the oil temperature sensor 7 And is output to the linear solenoid valves SLS and SLU of the hydraulic circuit. The control unit U has feedback control means 1 and the feedback control means 1 controls the progress of a shift such as a rotational acceleration of an input shaft speed that changes based on a gear ratio related to a shift to a predetermined shift speed. A proportional (P) control operation 1a based on the deviation between the actual change amount and the target change amount accompanying the change and a proportional gain, an integral (I) control operation 1b based on the integral value of the deviation and the integral gain, and a differentiation of the deviation And a differential (D) control operation 1c based on the differential value and the differential gain.
Mainly, the proportional control operation 1a functions, and with the progress of the subsequent shift, for example, in the second and subsequent calculation cycles, the integration control operation 1c is changed to mainly function, and the feedback control means performs predetermined control. The signal is the hydraulic control means, the linear solenoid valve SLS or S
Output to LU.

FIG. 2 is a diagram schematically showing a hydraulic circuit.
A plurality of friction engagement elements having the two linear solenoid valves SLS and SLU and switching a transmission path of a planetary gear unit of the automatic transmission mechanism to achieve, for example, a forward fourth speed or a fifth speed and a reverse first speed. A plurality of hydraulic servos 9 for connecting and disconnecting (clutches and brakes)
It has 0. Further, the linear solenoid valve S
Solenoid modulator pressure is supplied to the input ports a ₁ and a ₂ of the LS and SLU, and the control oil pressure from the output ports b ₁ and b _{2 of} these linear solenoid valves is applied to the control oil chambers of the pressure control valves 11 and 12, respectively. 11a and 12a. The pressure control valves 11 and 12 supply the line pressure to the input ports 11b and 12b, respectively, and the pressure control oil pressure from the output ports 11c and 12c adjusted by the control oil pressure is applied to the shift valves 13 and 15 respectively. Are supplied to the hydraulic servos 9 and 10 as appropriate.

This hydraulic circuit is for showing the basic concept, and the hydraulic servos 9 and 10 and the shift valves 13 and 15 are symbolically shown.
A number of hydraulic servos are provided corresponding to the automatic transmission mechanism, and a number of shift valves for switching the hydraulic pressure to these hydraulic servos are also provided. Further, as shown in the hydraulic servo 10, the hydraulic servo has a piston 19 which is fitted to the cylinder 16 in an oil-tight manner by an oil seal 17, and the piston 19 is provided with a pressure control valve 12 acting on a hydraulic chamber 20. The outer friction plate 22 and the inner friction material 23 come into contact with each other based on the pressure-adjusted hydraulic pressure from the first and second springs, and move against the return spring 21. Although the friction plate and the friction material are shown by the clutch, it is needless to say that the friction plate and the friction material also correspond to the brake.

Next, a hydraulic control device according to the present invention will be described with reference to FIGS.

A signal from the throttle opening sensor 3 and the vehicle speed sensor 6 based on the driver's operation of the accelerator pedal determines a shift based on a shift map in the control unit U, for example, an upshift of 2 → 3 shift. Then, as shown in FIGS. 3 to 5, after a predetermined time for the pretreatment operation by the predetermined shift valve, the shift control of the engaging pressure P _A and release hydraulic pressure P _B is started (S1). In the shift control, the driver performs an upshift control in a power-on state in which power is transmitted from the engine to the wheels during shifting, while holding the accelerator pedal substantially constant. Then, a predetermined signal is output to the linear solenoid valve SLS (or SLU) so that the hydraulic pressure (engagement hydraulic pressure) P _{A applied} to the hydraulic servo on the engagement side becomes the predetermined pressure P _S1 (S
2). The predetermined pressure (limit pressure) P _S1 is set to a hydraulic pressure required to fill the hydraulic chamber 20 of the hydraulic servo, and is maintained for a predetermined time t _SA . When the predetermined time t _SA has elapsed (S
3), the engagement pressure P _A is the predetermined gradient _{[(P S1 -P S2) /} t
_SB ] to reduce the hydraulic pressure (hereinafter referred to as sweep down) (S
4), when the engagement hydraulic pressure P _A reaches the predetermined low pressure P _S2 (S5),
The sweep down is stopped, and the predetermined low pressure P _S2 is maintained.
It is on standby (S6). The predetermined low pressure P _S2 is set to a pressure that is equal to or higher than the piston stroke pressure and does not cause a change in the rotation of the input shaft, and the predetermined low pressure P _S2 is maintained until the time t reaches a predetermined time t _SE (S7). ). The above steps S1 to S7 are performed by the friction plates 22, 23 of the friction engagement elements.
The servo start control is to move the piston 19 of the hydraulic servo 10 to a state immediately before the torque capacity is generated by eliminating the play (to reduce backlash).

Then, the engagement side shared torque T _A corresponding to the input torque T _t is calculated (for example, if a is the torque sharing ratio, T _A = 1 / a · T _t ) (S8), and The engagement target hydraulic pressure P _TA immediately before the rotation change of the input shaft rotation speed _NT starts (immediately before the start of the inertia phase) is calculated by a predetermined function based on the joint side shared torque T _A (S9). The engagement side oil pressure P _TA immediately before the start of the inertia phase is B _A ; piston stroke pressure (= spring load), A _A ; effective radius of friction plate × piston area × number of friction plates × friction coefficient, dP
_TA : P _TA = (T _A / A
It is calculated by _{A) + B A + dP TA} . Further, the margin (tie-up degree) S _11, S _21, and set in consideration of the drive feeling of the tie-up degree of the release side frictional engagement element, engaging the target pressure P _TA is corrected set Is performed (S10). Then, based on the engagement oil pressure P _TA immediately before the start of the inertia phase calculated according to the input torque T _t ,
A predetermined gradient is calculated based on a predetermined predetermined time t _TA [(P _TA −P _S2 ) / t _TA ], and the engagement-side hydraulic pressure is increased based on the gradient (hereinafter referred to as “sweep up”) (S11).
Due to the first sweep-up having the relatively steep gradient, the engagement torque is increased, and the state immediately before the input shaft speed change starts, that is, the calculated predetermined target engagement hydraulic pressure P _TA
The oil pressure rises until (S12).

The input torque _T.sub.t (= turbine torque) is obtained by linearly interpolating the engine torque based on the throttle opening and the engine speed using a map on the basis of the running condition of the vehicle. A speed ratio is calculated from the numbers, a torque ratio is determined from the speed ratio by a map, and the engine torque is multiplied by the torque ratio.

When the target engagement oil pressure _PTA is reached, that is, when it is predicted that the inertia phase has started in which the input shaft rotation speed starts to change, the oil pressure change δP
Target rotation change _{rate TA} is the target of the rotation change start of the input shaft rotational speed N _T functions corresponding to the (angular acceleration) .omega.a [[delta]
P _TA = fδ _PTA (ωa)] (S1
3). That is, k is a constant, t _aim is the target shift start time, ω
Assuming that a is the target rotation change rate [gradient to the target rotation speed] and I is the inertia amount, the hydraulic pressure change δP _TA = [I · ωa]
/ [K · t _aim ]. Then, the sweep-up with a gradient by the hydraulic change δP _TA (S14).
The second sweep-up is continued until the rotation change ΔN from the input shaft rotation speed N _TS at the start of the rotation change reaches the predetermined shift start determination rotation speed dN _S (S15), and the engagement-side hydraulic pressure P _A is increased. , The inertia phase start hydraulic pressure _PIN becomes substantially the same as the clutch capacity of the engine torque.

In steps S8 to S14, the torque carried by the engagement-side clutch is increased and the torque carried by the disengagement-side clutch is reduced, so that the gear ratio is in the state before the upshift (second speed). This is a torque phase control in which only the torque sharing changes. The start of the rotation change of the input shaft rotation speed _NT refers to the start of the inertia phase, that is, the shift (2 → 3 shift) based on the gear ratio is started, and is related to the gear ratio with respect to the rotation speed of the output shaft. This is a state in which the change in the input shaft speed has been started, and is calculated from the input shaft speed sensor 5 and the vehicle speed sensor 6.

Next, feedback control according to the main part of the present invention, which will be described later, is performed (S16). The feedback control corresponds to the inertia phase and the end control in FIGS. In the feedback control, the change in hydraulic pressure is a predetermined gradient δ
This is performed with reference to P _I , and the sweep-up by δP _I is continued until a2 [%] of the rotation change amount ΔN from the start of the shift (start of rotation change) to the completion of the shift, for example, 90 [%] ( S19). That is, the input shaft speed during the shift start to N _TS, rotational variation amount of the rotation change start the .DELTA.N, g _i
Is the gear ratio before shifting and g _{i + 1} is the gear ratio after shifting.
[(ΔN × 100) / ｛(N _TS / g _i ) × (g _i −g
_{i + 1} )｝] becomes a2 [%].

Further, when the rotation change amount exceeds a2 [%], a time t _F is set (S20), and this state substantially corresponds to a state in which the inertia phase and the end control are completed. Further, a relatively steep oil pressure change δP _F is set, and the oil pressure suddenly sweeps up due to the oil pressure change (S2
1) Then, after a predetermined time t _FE, which is set to a time sufficient to increase to the engagement pressure from the time t _F , has elapsed (S22), the hydraulic control on the engagement side is completed. Steps S20 and S21 are the completion control.

[0040] Then, along the Figures 3 and 6, a description will be given of the control of the disengagement side pressure P _B in the upshift mentioned above.

First, in response to a shift command from the control unit U, timing of the disengagement side hydraulic control is started simultaneously with the engagement side (S2).
5). Then, the calculated allotted torque T _B of the disengagement side frictional engagement element based on the input torque T _t (S26), the engagement pressure P _W is calculated further for the disengagement side torque distributed T _B (S27). The release hydraulic pressure P _B is supplied with the hydraulic pressure P _W consisting of the engagement pressure (S28). The supply of the hydraulic pressure P _W is continued until the engagement hydraulic pressure P _A starts the first sweep-up (start of the torque phase). ) (T _SE ) wait and hold (S29). Therefore, steps S26 to S28 are standby control.

Then, the engagement oil pressure P _A and the input torque T _t
[T _B = f _TB (P _A , T _T )], the release side torque T _B is calculated (S30), and the margins S ₁₁ and S ₂₁ are taken into consideration (T _B = S ₁₁ × T _B). + S _21), the disengagement side allotted torque T _B is calculated (S31). Then, released from the release side torque distributed T _B pressure P _B is calculated [P _B = f
_PB (T _B )] (S32). That is, first, the torque T _A shared by the engagement-side frictional engagement elements is [T _A = A _A × (P _A −B
_A )] (A _A ; effective radius x piston = area x number of sheets x friction coefficient, B _B : piston stroke pressure), and thereby the torque T shared by the disengagement side frictional engagement element
_B is calculated by [T _B = (1 / b) T _t − (a / b) T _A ]. Here, b is the torque share on the release side,
a is the torque share on the engagement side, and _Tt is the input shaft torque.
The margin by (tie-up degree) S _11, S _21,
The tie-up degree of the engagement side frictional engagement element, and set in consideration of the drive feeling, the disengagement side torque T _B [T
_B = S ₁₁ × T _B + S ₂₁ ] (S 31). The allowance ratios S ₁₁ and S ₂₁ are arbitrarily set so as to match the driver's feeling in a number of throttle opening / vehicle speed maps selected according to the difference in oil temperature. ₁₁ > 1.0 and S ₂₁ > 0.0. Further, the release hydraulic pressure P _B is calculated from [P _B = (T _B / A _B ) + B _B ] from the release side torque T _B taking the margin into consideration (A _B ; Release side frictional engagement element). Effective radius × piston area × number of sheets × friction coefficient, B _B ; release side piston stroke pressure) (S32).

The release hydraulic pressure P _B calculated as described above
Is dependent on the engagement oil pressure P _A , so the two-step gradient that bends at the start of the inertia phase (t _TA ) at which the input shaft rotation speed starts to change, ie, the first engagement side A relatively steep sweepdown corresponding to a sweepup and a relatively gentle sweepdown corresponding to a second sweepup on the engagement side. Then, the sweep-down is performed in the same manner as the engagement side, so that the input shaft rotation change amount ΔN
Continues until the predetermined rotation change start determination rotational speed dN _S is reached (S33). Next, the change δP _{E of the} release oil pressure is set, and the sweep down is performed with the gradient due to the change of the oil pressure (S3
4), the sweep-down is continued until the release-side hydraulic pressure P _B becomes 0 (S35), thereby the hydraulic pressure control of the disengagement side is completed. Step S34 is release control.

Next, a feedback control which is a subroutine of step 16 in the flowchart of the engagement side hydraulic control (see FIG. 5) will be described with reference to FIGS. 7, 8 and 9.

When the feedback control is started, that is, when the variation ΔN of the input shaft speed that changes based on the gear ratio related to the shift to the predetermined shift speed becomes equal to or more than the measurable predetermined value dN _S (the input shaft speed at that time). N _TS ) and timing starts (t =
0) (S40), and the control unit (ECU) U of the operational cycle is set to n = 0 (S41), is further stored engagement hydraulic pressure P _A in this case is as inertia phase start oil pressure P _IN (S42 ). Then, the engagement side oil pressure sweeps up from the inertia phase start oil pressure P _IN at a predetermined gradient δP ₁ set in advance in the map (S43), and the sweep up is maintained until a predetermined time t _IS elapses. . That is, immediately after the start of the change in the input shaft rotation, the rate of change in the number of rotations is not stabilized. _Therefore , the feedback control is not performed for the predetermined time t _IS until the rotation becomes stable in consideration of the response delay of the hydraulic pressure.

After the lapse of the predetermined time t _IS , the oil pressure change δ
To P _1, along said engagement side hydraulic pressure P _A as a control amount, the target value of the input shaft rotational speed of the rotational acceleration of the detected amount (amount of change) (the target rotational acceleration) .omega.1 and the upper limit ω1max , The lower limit ω1min is set (S4
5). The target rotation acceleration ω1 is input shaft larger the rotational speed N _TS to start the rotation change is large, is set such that the absolute value omega _A increases. Then, a deviation E between the actual rotation acceleration (extreme value) ωr calculated based on the input shaft rotation speed (gear ratio) _NT and the target rotation acceleration ω1.
(= Ω1−ωr) is calculated (S46).

Further, the deviation E is compared with the upper limit value ω1max and the lower limit value ω1min of the target rotational acceleration (S47), and it is determined whether the operation cycle is the first (n = 0) (S48). . The deviation E is within a predetermined range of the target rotational acceleration (ω1max <E <ω1min).
, The feedback amount (command value) P _FB is set to 0 (S49). If the deviation E is out of the predetermined range of the target rotational acceleration and the first calculation cycle (n = 0), the process proceeds to step S50,
If the operation cycle is the second or later (n ≧ 1), the process proceeds to step S51.

Here, K1 is a proportional gain, K2 and K4
Is a differential gain, K3 is an integral gain, and Δω is a differential value of the deviation E. In step S50, that is, the feedback command value P _FB l in the first calculation cycle
(0) is P _FB l (0) = K1 × E + K2 × Δω. In step S51, that is, the feedback command value P _FB l (n) in the second and subsequent calculation cycles.
Is P _FB l (n) = l (n−1) + K3 × E + K4 × Δω. In addition, ln is the accumulated value of the n-th operation cycle, and l (n-1) is [K3 ×
E + K4 × Δω].

That is, the general formula is: P _FB l (n) = K1 × E (n) + K2 × Δω (n) +1
(N) ln = l (n-1) + K3 × E (n) + K4 × Δω
(N) However, when n = 0, K3 = K4 = 0, that is, l (0) =
K1 = K2 = 0 when 0 and n ≧ 1.

Therefore, the first operation cycle (n = 0)
Then, a proportional (P) control operation (K1 * E) based on the proportional gain K1 and the deviation E, and a differential (D) control operation (K2 * Δω) based on the differential gain K2 and the differential value (Δω) of the deviation. Is carried out by the PD control operation to which is added, and the responsiveness is mainly improved by the proportional (P) control operation, thereby improving the convergence to the target value for a large input shaft rotation change amount at the beginning of the shift. Feedback control can be performed. In the second and subsequent operation cycles (n ≧ 1), the integral (I) control operation (l (n−1) + K3 *) based on the integral value of the integral gain K3 and the deviation E is performed.
E) and the differential gain K4 and the differential (D) control operation (K4 * Δω) based on the differential value of the deviation E are added.
Performed in the D control operation, and by performing smooth control mainly by the integral (I) control operation, feedback control in which hunting is prevented can be performed.

The feedback command value P _FB (n) based on each gain is added to the predetermined gradient δP ₁ from the hydraulic pressure P _IN at the start of inertia, and a correction oil pressure ΔT by feed forward control is added. P _A [= P
_IN + P _FB (n) + δP ₁ + ΔT] is set (S5
2). The correction hydraulic pressure ΔT is used to correct a change in the input torque when the input torque changes due to an accelerator being turned on during a shift, for example, when the input torque changes during the shift. If the change is compensated for by the feedback control, the amount of correction of the hydraulic pressure by the feedback control becomes insufficient, and there is a possibility that convergence to the target value is reduced and hunting occurs.

Next, while incrementing the operation cycle n (n = n + 1) (S53), the feedback control (S46 to S52) is performed for each operation cycle.
The feedback control is based on the input shaft speed (gear ratio).
The change amount ΔN from the start of the shift (start of the rotation change) is a1
[%], For example, to 70 [%] (S54).
Step S40~S53 becomes inertia phase control, engagement hydraulic pressure P _A is feedback-controlled along the swept up by the hydraulic change [delta] P _1.

When the rotation change amount ΔN exceeds a1 [%], the process proceeds to the end control shown in FIG. In the end control, first, the target rotational acceleration ω2 and its upper and lower limits are set again (S56). In this case, since the rotational acceleration used as the target value of the feedback control includes a delay due to a hydraulic response delay or the like, when calculating the target rotational acceleration ω2 used for the end control, the delay (offset) is considered in advance. Is done. That is, assuming that the actual delay time (experimental value) of the rotational acceleration ωr is t _D and the target shift time from the start of shifting to the end control is t _E , the delay ratio offset is given by offset =
It is calculated by (t _D / t _E ) * 100 [%]. Further, the current progress ratio based on the gear ratio to the predetermined gear is represented by R
= [(ΔN * 100) / ｛(N _TS / g _i ) * (g _{i + 1} −
g)｝], the target rotational acceleration ω2 is: ω2 = min ｛(100−R + offset) / (10
0−a1), 1｝ * ω1.

That is, the remaining ratio (100-R + offse) until the shift is completed based on the current actual input shaft rotation state.
t) and the ratio of the remaining ratio (100-a1) at the start of the final control, which is set in advance, is compared with 1, and the smaller one is multiplied by the target rotational acceleration ω1 set in step S45. Can be For example, the actual shift progress ratio R is 76 [%], and the delay ratio off
t is 5%, and the ratio a1 at the start of the final control is 70%.
In the case of ω2 = min ｛(100−76 + 5) / (100−70), 1｝ * ω1 = (29/30) * ω1 = 0.966ω1.

Then, similarly to step S46, the deviation E between the actual rotational acceleration ωr and the target rotational acceleration ω2 is calculated (S57), and the upper limit ω2max and the lower limit ω2min of the deviation E and the target rotational acceleration are calculated. (S58), and when the deviation E is within the upper and lower limit range of the target rotational acceleration (ω2max <E <ω2min), the feedback command value P _FB is set to 0 (S5).
9). Further, when the deviation E is out of the range of the upper and lower limits of the target rotational acceleration, as in the case of the step S51, the ID is calculated by a calculation cycle following the inertia phase control.
Feedback control based on the control operation is performed, and smooth control is mainly performed by the integral (I) control operation (S6).
0). That is, the feedback command value P _FB l (n) is P _FB l (n) = l (n−1) + K3 * E + K4 × Δω.

Further, in the final control, the sweep-up reference value (the amount of change in hydraulic pressure) δP _L of the engagement-side hydraulic pressure P _A is
It is set to be lower than the reference value δP _{1 in} the inertia phase control, and similarly to the step S52, the correction hydraulic pressure ΔT by the feedback control is added, and the engagement side hydraulic pressure P
_A is set by P _A = P _IN + P _FB (n) + δP _L + δP ₁ + ΔT (S61). Further, while the operation cycle n is incremented (n = n + 1) (S62), the target rotational acceleration ω2 is set for each operation cycle (S5).
The above feedback control from 6) is repeated, and feedback control is performed so that the engagement side oil pressure P _A follows a reference value δP _L having a sweep gradient that is gentler than the reference value δP _I of the inertia phase control. The feedback control by the end control is continued until the change amount ΔN of the input shaft rotation speed (gear ratio) from the start of the shift (start of rotation change) to a2 [%], for example, 90 [%] (S63). When the variation exceeds a2 [%], the process proceeds to the completion control in step 20 in FIG.

In the final control, since the change in the input shaft rotation is also stable, the feedback control need not be performed. Even in the final control, the target rotational acceleration ω1 in the inertia phase control is required. The feedback control may be performed as it is.

FIG. 10 is a diagram showing changes in the proportional (P) gain K1 and the integral (I) gain K3 with respect to the oil temperature.
Proportion (P) that functions in the first operation cycle (n = 0)
The gain K1 is set to decrease as the oil temperature increases, as shown in FIG. When the oil temperature is high, the responsiveness of the control system is improved, but in this case, if the proportional gain is increased, even if the convergence to the target value is increased, there is a risk that hunting will be exceeded thereafter, and as described above, Hunting can be prevented by setting the proportional gain.

The second and subsequent operation cycles (n ≧ n)
As shown in FIG. 10B, the integral (I) gain K3 that functions in 1) is set to increase as the oil temperature increases. The integral (I) control operation based on the integral gain is accumulated for each operation cycle, but if there is a hydraulic pressure response delay,
Hydraulic pressure correction by the integral gain cannot be controlled in one operation cycle time, and it is accumulated for each cycle, and a considerable amount of correction is required before the hydraulic pressure responds. Therefore, when the oil temperature is low, that is, when the response of the control system is poor, if the integral (I) gain K3 is increased, large hunting will occur. Therefore, in this case, large hunting can be prevented by reducing the integral gain and increasing the proportional gain.

FIG. 10 is a diagram showing the setting of the feedback control correction amount based on the target rotational acceleration deviation E. As the absolute value of the deviation E increases, the feedback control correction amount increases, that is, the gains K1, K2, K3, K4
However, the correction amount (oil pressure correction amount) is changed depending on the sign of the deviation. In the hydraulic control by the clutch-to-clutch, the side on which the clutch is engaged, that is, the side on which the hydraulic pressure is increased is the safe side in the direction in which the shift proceeds.
When the deviation E is positive, that is, when the target rotational acceleration ω is larger than the actual rotational acceleration, feedback control is performed to increase the engagement-side hydraulic pressure. The value is set to be larger than the negative side of the deviation E for the feedback control. Thus, by performing large feedback control in the direction of decreasing the oil pressure, it is possible to prevent the progress of the shift from being hindered or slowed down.

In the above embodiment, the proportional (P) control operation based on the proportional gain K1 is performed only in the first operation cycle, and the feedback control is performed by the integral (I) control operation based on the integral gain K3 in the second and subsequent operations. But
However, the present invention is not limited to this.
Of course, the integral control operation may be performed after the first time, and the proportional gain may be gradually decreased and the integral gain may be gradually increased with the increase of the operation cycle. In the early stage, if the proportional (P) control operation mainly functions, and if the change of the gear ratio is changed so that the integral (I) control operation mainly functions from the proportional control operation according to the progress of the subsequent gear shift, the technique of the present invention may be used. Included in the target range. Further, in the above embodiment, the up-shift speed change mainly based on the engagement-side hydraulic pressure has been described. Of course, the same can be applied to the feedback control of the release hydraulic pressure in Japanese Patent Application Laid-Open No. 9-170654.

[Brief description of the drawings]

FIG. 1 is a block diagram showing an electronic control unit according to the present invention.

FIG. 2 is a diagram schematically showing a hydraulic circuit according to the present invention.

FIG. 3 is a time chart of an input shaft rotation speed, an engagement hydraulic pressure, and a release hydraulic pressure based on a gear ratio in a power-on upshift.

FIG. 4 is a flowchart showing control of an engagement side hydraulic pressure of an upshift.

FIG. 5 is a flowchart showing a continuation of FIG. 4;

FIG. 6 is a flowchart showing control of an upshift release hydraulic pressure.

FIG. 7 is a flowchart showing feedback control according to the present invention.

FIG. 8 is a flowchart showing a continuation of FIG. 7;

FIG. 9 shows the input shaft rotation speed by the feedback control,
Time chart of acceleration and feedback command value.

FIG. 10 is a diagram showing setting of a proportional gain and an integral gain.

FIG. 11 is a diagram showing setting of a feedback control correction amount.

[Explanation of symbols]

U control unit 1 feedback control means 1a proportional (P) control operation 1b integration (I) control operation 1c differentiation (D) control operation 2-7 sensor 9,10 hydraulic servo K1 proportional gain K2, K4 differential gain K3 integral gain n operation Cycle ω1 Target change amount (target rotation acceleration) ω2 Target change amount of end control (rotation acceleration) ωr Actual change amount (rotational acceleration of input shaft rotation speed that changes based on gear ratio) t _IS predetermined time ΔT Corrects hydraulic pressure means

──────────────────────────────────────────────────の Continued on the front page (51) Int.Cl. ⁶ Identification code FIF16H 59:72 (72) Inventor Masao Saito 10 Takane, Fujiimachi, Anjo-shi, Aichi Prefecture Inside Aisin AW Co., Ltd. (72) Inventor Masaaki Nishida 10 Takane, Fujii-machi, Anjo, Aichi Prefecture Inside Aisin AW Co., Ltd.

Claims (10)

[Claims] An input shaft to which power from an engine output shaft is input, an output shaft connected to wheels, and a plurality of friction engagement elements for changing a power transmission path between the input shaft and the output shaft. A hydraulic servo for disconnecting and engaging these friction engagement elements, hydraulic control means for controlling the hydraulic pressure of the hydraulic servo,
A feedback control unit that outputs a control signal to the hydraulic control unit so that an actual change amount that changes when the shift progresses matches the target change amount. The control means has at least a proportional control operation based on a deviation and a proportional gain between the target change amount and the actual change amount, and an integral control operation based on an integral value and an integral gain of the deviation, and In the initial stage of the automatic transmission, the proportional control operation mainly functions, and after the shift progresses, the proportional control operation is changed so that the integral control operation mainly functions. Hydraulic control device. 2. The feedback control means detects the actual change amount in each operation cycle and outputs the control signal. In a first operation cycle in which the control signal is output, the feedback control means performs the proportional control operation. The hydraulic control device for an automatic transmission according to claim 1, wherein the integral control operation is functioned in the second and subsequent calculation cycles. 3. The hydraulic control device for an automatic transmission according to claim 1, wherein the proportional gain is set smaller as the oil temperature is higher, and the integral gain is set larger as the oil temperature is higher. . 4. The hydraulic control device for an automatic transmission according to claim 1, wherein said feedback control means outputs a control signal having a different hydraulic pressure correction amount depending on whether the difference is positive or negative. 5. The hydraulic control device for an automatic transmission according to claim 1, wherein the feedback control means outputs a feedback control signal a predetermined time after the progress of the shift is started. 6. The hydraulic pressure control apparatus for an automatic transmission according to claim 1, wherein said feedback control means includes means for correcting a hydraulic pressure with respect to a change in input torque during a gear shift. 7. The hydraulic control device for an automatic transmission according to claim 1, wherein said feedback control means is configured by adding a differential control operation to each of said proportional control operation and said integral control operation. 8. The amount of change that changes when the shift progresses,
The hydraulic control device for an automatic transmission according to any one of claims 1 to 7, wherein the rotational speed is a rotational acceleration of an input shaft speed that changes based on a gear ratio related to a shift to a predetermined shift speed. 9. A shift control to a predetermined shift speed is performed by engaging a first frictional engagement element of the plurality of frictional engagement elements and releasing a second frictional engagement element. The hydraulic control means controls mainly a hydraulic pressure acting on one hydraulic servo of the first and second friction engagement elements, and controls a hydraulic pressure acting on the other hydraulic servo. The hydraulic control device for an automatic transmission according to any one of claims 1 to 8, wherein a control signal is output to control a hydraulic pressure acting on one of the main hydraulic servos. 10. The feedback control means includes: an inertia phase control in a state where the progress of the shift does not exceed a predetermined ratio; and an end control in a state where the progress of the shift exceeds the predetermined ratio. The hydraulic control device for an automatic transmission according to any one of claims 1 to 9, wherein the target change amount is set in consideration of a delay of an actual change amount.