JPH11270665A - Hydraulic control device of automatic transmission - Google Patents

Abstract

PROBLEM TO BE SOLVED: To hold precise timing at all times and suppress generation of tieup and engine blow despite dispersion as products and secular change of friction fitting elements in addition to up-shift due to proper timing on the basis of the control of the release side oil pressure related to the rise of the engaging side oil pressure by making feedback control of the release side oil pressure. SOLUTION: The release side oil pressure PB is calculated from the release side sharing torque TB calculated from the engaging side oil pressure PA (S30 and S32). At the time of tieup, the release side oil pressure is feedback controlled with a correction amount obtained by multiplying the correction oil pressure dPw (>0) calculated from the difference between the input shaft rotational acceleration dNT and the reference rotational acceleration WST by the correction gain SdPTie (S37). At the time of engine blow, the release side oil pressure is feedback controlled with a correction amount obtained by adding the oil pressure correction value dPup on the basis of the difference Nup of the input shaft revolving speed from the reference revolving speed to the correction amount based upon the correction oil pressure dPw (<), (S39).

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a hydraulic control system for an automatic transmission mounted on an automobile, and more particularly to a so-called clutch for engaging a predetermined frictional engagement element and releasing another frictional engagement element. The present invention relates to a hydraulic control device that performs pressure regulation control in association with the hydraulic pressure for both friction engagement elements in a two-clutch shift.

[0002]

2. Description of the Related Art Conventionally, as shown in, for example, Japanese Patent Application Laid-Open No. 5-157168, at the time of upshifting to a predetermined gear, for example, at the time of 2 → 3 shift (the brake B-3 is released together with the engagement of the brake B-2). Suppose), the supply pressure (B-2 pressure) to the hydraulic servo and the accumulator on the engaging side acts on the control oil chamber of the 2-3 timing valve, thereby communicating with the hydraulic servo and the accumulator on the releasing side. Release pressure (B-
3), the engagement pressure and the release pressure are adapted to have an inverse ratio relationship, and the release pressure is controlled to decrease linearly from the start of the shift to at least the end of the torque phase. It has been devised.

In the hydraulic servo control method, it is possible to correct the engagement / disengagement timing appropriately in response to a change in the throttle opening during shifting, as compared with a method in which the releasing pressure is performed at once at a predetermined timing. Shock can be suppressed.

[0004]

However, in the above hydraulic servo control method, the release pressure is regulated by the timing valve so that the release pressure is in a predetermined relationship with the supply pressure. In particular, tie-ups and engine blowing may occur due to product variations, aging, and temperature changes of friction engagement elements such as brakes and clutches.

SUMMARY OF THE INVENTION It is an object of the present invention to provide a hydraulic control device for an automatic transmission which solves the above-mentioned problems by performing feedback control of the release hydraulic pressure.

[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), the first frictional engagement element of the plurality of frictional engagement elements is engaged, and the second frictional engagement element is released to perform an upshift to a predetermined gear position. The second frictional engagement is further achieved by the release-side shared torque (T _B ) calculated based on the engagement-side hydraulic pressure (P _A ) acting on the first frictional engagement element hydraulic servo. In a hydraulic control apparatus for an automatic transmission, which calculates a release hydraulic pressure (P _B ) acting on an element hydraulic servo, an engine based on a state in which both the first and second frictional engagement elements are in a disconnected state. Detecting means (U ₂ ) for detecting a tie-up state based on a blowing state and a connection state of both of the friction engagement elements, and a torque phase control for controlling both the engagement side hydraulic pressure and the release side hydraulic pressure, Correction amount based on the signal from the detection means And feedback control means (1c) for correcting said disengagement hydraulic pressure, lying in the hydraulic control device for an automatic transmission, characterized in that comprises a.

According to a second aspect of the present invention, the detecting means (U ₂ ) is means (1b) for detecting the rotational acceleration of the input shaft (dN _T ), and the feedback control means is configured to detect the rotational acceleration. 2. The hydraulic control of the automatic transmission according to claim 1, wherein a difference between the hydraulic pressure and the reference rotational acceleration (W _ST ) is calculated, and the release hydraulic pressure (P _B ) is corrected by a first correction amount based on the difference. In the device.

According to a third aspect of the present invention, in addition to the detection of the rotational acceleration of the input shaft, the detecting means (U ₂ )
A means for detecting an input shaft rotational speed (1a, 1b), calculates a difference between the feedback control means (1c), the input shaft rotational speed (N _T) and the reference input shaft speed (N _O × gi) and, difference a second correction amount based on the (N _up) (dP _up)
Is added to the first correction amount, and the release hydraulic pressure (P _B )
Is corrected (see S35 and S39).

According to a fourth aspect of the present invention, when the rotational acceleration (dN _T ) is smaller than a reference rotational acceleration (W _ST ) (ie, at the time of tie-up), the feedback control means (1c) performs the rotation control. correcting the first correction amount based on the absolute value of the hydraulic (dP _W) calculated from the difference between the acceleration and the reference rotational acceleration (W _ST -dN _T), to subtract from said disengagement hydraulic pressure (P _B) (S33, S34, S
37), when the rotational acceleration is greater than the reference rotational acceleration (ie, when the engine is blowing), the first value based on the absolute value of the hydraulic pressure (dP _W ) calculated from the difference between the rotational acceleration and the reference rotational acceleration. Is corrected to be added to the release hydraulic pressure (P _B ) (S33, S3
4, S39), the hydraulic control device for an automatic transmission according to claim 2 or 3.

According to a fifth aspect of the present invention, the first correction amount is a difference between the rotational acceleration and a reference rotational acceleration (W _ST −d
N _T ) is calculated by multiplying the hydraulic pressure (dP _W ) calculated based on _NT ) by the correction gain (SdP _Tie ) (dP _W × SdP).
_Tie ), the correction gain is changed by the input torque (T _t ) (see FIG. 10). The hydraulic control apparatus for an automatic transmission according to claim 2, wherein

According to a sixth aspect of the present invention, the second correction amount (dP _up ) is generated only when the input shaft rotation speed (N _T ) is larger than a reference input rotation speed by a predetermined reference value or more. ,
The hydraulic control device for an automatic transmission according to claim 3, wherein the larger the difference ( _Nup ) between the input shaft rotation speed and the reference input shaft rotation speed, the larger the value (see Fig. 11).

[Operation] Based on the above configuration, the engagement-side hydraulic pressure (P _A ) is increased to engage the first frictional engagement element, and the increase in the engagement-side hydraulic pressure (P _A ) is caused. Then, the release-side hydraulic pressure (P _B ) is lowered to release the second frictional engagement element, and upshift to a predetermined gear. At this time, as shown in FIG. 7, when the tie-up input shaft speed (N _T) is reduced, becomes smaller than the input shaft rotational acceleration (dN _T) is the reference rotational acceleration (see dotted line), the With the correction amount calculated based on the difference, the release side hydraulic pressure (P _B ) is feedback-controlled to decrease.

As shown in FIG. 8, when the input shaft rotation speed (N _T ) increases due to engine blowing, the input shaft rotation acceleration (dN _T ) becomes a large value as compared with the reference rotation acceleration. The release side hydraulic pressure (P _B ) is feedback-controlled to the upward side by the correction amount calculated based on the correction amount.

Further, as shown in FIG. 9, when engine blowing occurs during standby control, the input shaft rotational acceleration (dN
_Since the change in _T ) is small, a correction hydraulic pressure (dP _up ) based on the difference (N _up ) of the input shaft rotation speed (N _T ) with respect to the reference input rotation speed is added, and the release hydraulic pressure (P _B ) increases. Is feedback controlled.

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

[0016]

According to the first aspect of the present invention, by detecting the engine blowing state and the tie-up state,
In the torque phase control, the release-side hydraulic pressure is feedback-controlled, so that in addition to the upshift at an appropriate timing based on the control of the release-side hydraulic pressure related to the increase in the engagement-side hydraulic pressure, the product variation and the aging of the friction engagement element Even if there is a change, it is possible to always maintain accurate timing and suppress occurrence of tie-up and engine blowing.

According to the second aspect of the present invention, since the engine blowing state and the tie-up state are detected based on the input shaft rotational acceleration, the automatic transmission can be easily and reliably based on the input shaft rotational speed sensor originally provided. By detecting the engine blowing and the tie-up state, it is possible to effectively suppress the occurrence thereof.

According to the third aspect of the present invention, the release side hydraulic pressure is feedback-controlled based on the input shaft speed, so that even if the engine blowing occurs during the standby control, it is possible to accurately control the engine blowing. Feedback control can be performed.

According to the present invention, the correction amount is calculated based on the difference between the rotational acceleration and the reference rotational acceleration regardless of the tie-up and the engine blowing. Feedback control for suppressing both engine blowing can be performed.

According to the fifth aspect of the present invention, since the correction gain is changed by the input torque, a small correction gain is used for a shift with a low throttle opening, and a shift is performed with a high throttle opening. Has a large correction gain,
Feedback control can always be accurately performed with an appropriate correction amount.

According to the sixth aspect of the present invention, the second correction amount occurs only when the difference between the input shaft rotation speeds is equal to or larger than the reference value and is larger as the difference is larger. Except when the engine is blown from the inside, the feedback control based on the input shaft rotational acceleration does not have a great influence as a disturbance.

[0022]

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) is provided. The input shaft of the automatic transmission mechanism is connected to an engine output shaft via a torque converter, and the output shaft of the automatic transmission mechanism is connected to drive wheels.

FIG. 1 is a block diagram showing an electric control system. U is a control unit (ECU) comprising 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 includes means U ₁ for controlling the engagement side oil pressure, means U ₂ for controlling the release side oil pressure calculated based on the engagement side oil pressure, an engagement side friction engagement element and a release side friction. and means U ₂ for detecting the tie-up state based on the engine racing condition and both of these elements based on the engagement element are both disconnected are both connected state, with a, in particular,
Means 1a for detecting the rotation speed of the input shaft based on the signals from the input shaft rotation speed sensor 5 and the vehicle speed sensor 6, particularly the change in the rotation speed of the input shaft with respect to the output shaft, and calculating the rotational acceleration based on the input shaft rotation speed ( Detecting means 1b, a first correction amount calculated based on a difference between the input shaft rotation acceleration and the reference rotation acceleration, and a second correction amount calculated based on a difference between the input shaft rotation speed and the reference input shaft rotation speed. And a means 1c for feedback-controlling the release-side hydraulic pressure based on the correction amount. The control means outputs a predetermined control signal to the linear solenoid valve SLS or SLU.

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 shown symbolically.
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 operation of the accelerator pedal by the driver determines a shift, for example, an upshift of 2 → 3 shift based on a shift map in the control unit U. 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 ] (hereinafter referred to as sweep down) (S4),
When the engagement pressure P _A becomes a predetermined low pressure P _S2 (S5), the sweep-down is stopped, is held and wait the predetermined low pressure P _S2 (S6). The predetermined low pressure P _S2 is set to pressure not to cause rotation change on a and the input shaft piston stroke pressure or, the predetermined low pressure P _S2 is clocked t is the predetermined time t _SE
It is held until the elapse (S7). The above steps S1 to S1
S7 is the servo activation control that moves the piston 19 of the hydraulic servo 10 to a state immediately before the torque capacity is generated by eliminating the play of the friction plates 22 and 23 of the friction engagement element (reducing the play).

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 ) (S 8), 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 throttle opening and the engine speed from a map on the basis of the running condition of the vehicle to obtain the engine torque. A speed ratio is calculated from the numbers, a torque ratio is determined from a map based on the speed ratio, and the engine ratio is calculated by multiplying the engine torque by the torque ratio.

Then, 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 decreased, 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 speed sensor 5 and the vehicle speed sensor 6.

Then, the engagement-side hydraulic pressure change δP _I is feedback-controlled by the rotation speed change amount ΔN based on the detection of the input shaft speed sensor 5 and is swept up by the gradient of the δP _I (S16). . The sweep-up by the δP _I is continued up to a ₁ [%] of the rotation change amount ΔN from the start of the shift (start of the rotation change) to the completion of the shift, for example, 70 [%] (S17). That, N _TS input shaft speed during the shift start, the rotational variation amount of .DELTA.N, pre-shift to g _i gear ratio, g
_{Assuming that i + 1} is the gear ratio after shifting, [(ΔN × 100) /
{(N _TS / g _i ) × (g _i −g _{i + 1} )}] is a1 [%]
Continue until Steps S16 and S17 above
However, inertia phase control in which the load of the engine torque changes the engine speed is performed.

Furthermore, if it exceeds a1 [%] of the rotational variation amount varies by feedback control based on a smooth input shaft rotational speed change amount ΔN hydraulic change [delta] P _L is set,
The sweep-up is performed by the gradient of δP _L (S1
8). The δP _L generally has a slightly gentler gradient than the δP _I , and the sweep-up is performed by changing the rotation speed change amount a2 from the start of the shift (start of the rotation change) to the vicinity of the completion of the shift.
[%], For example, 90 [%] (S19).
The sweep-up target shift time t _I based on δP _I and δP _L is set based on a plurality of different throttle opening / vehicle speed maps depending on the oil temperature. Step S18 is the end control.

When the target shift time t _I has elapsed, a time t _F is set (S20), and this state substantially corresponds to a state in which the inertia phase and the end control have been 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.

[0035] 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 ; open-side piston stroke pressure).

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.

Next, the release hydraulic pressure P which is the main part of the present invention
First, the feedback control of _B will be described. First, a predetermined time (input shaft rotation acceleration) of an input shaft rotation change (input shaft rotation acceleration) at the time of predetermined shift control (for example, at the time of 2 → 3 shift).
Mean value W _ST and the actual (difference between the input shaft rotational speed change (input shaft rotational speed acceleration) dN _T in) at present from the previous 0ms) (W
_ST -dN _T) was calculated by multiplying the inertia amount I thereto, requires feedback torque amount dT _W at disengagement side pressure control in due to engagement hydraulic pressure based on the above step S32 is calculated (dP _W = I × (W _ST −d
_NT ) (S33). Note that the average value W _ST becomes a reference rotational acceleration, and the required feedback torque amount dT _W
In the calculation of the average value _WST and the actual acceleration dN _T
Based on this difference, the influence of disturbance due to unevenness of the road surface or the like is eliminated.

The required feedback amount dT _W
A feedback correction hydraulic pressure dP _W [= df _PB (dT _W )] of the release hydraulic pressure is calculated by a predetermined function based on
34). Further, from the input shaft rotation speed _NT to the output shaft rotation speed N _O
Multiplied by the pre-shift gear ratio gi ((N _T −N _O ×
gi), the change amount of the input shaft rotation speed excluding the change in the vehicle speed and the gear ratio, that is, the engine blowing amount N _up is calculated (S35).

It is determined whether the feedback correction hydraulic pressure dP _W calculated in step S34 is in a negative state (S3).
6). Small current input shaft rotational acceleration dN _T with respect to the reference rotational acceleration W _ST, the correction when hydraulic dP _W is positive, that is, when the tie-up, the release-side hydraulic pressure P _B, step S3
At 7, the correction is made in the lowering direction. That is, as shown in FIG. 10, the feedback control gain SdP _Tie is set to be larger in proportion to the input torque T _t, the release-side torque distributed T disengagement side pressure P _B calculated at _B From [= fP _B (T _B )] (see step S32), the control gain S is added to the feedback oil pressure dP _W.
The value obtained by subtracting the oil pressure multiplied by dP _Tie becomes the release oil pressure P _B corrected by the feedback control [P _B = f
P _B (T _B ) −dP _W × SdP _Tie ] (S37).

[0042] Thus, as shown in FIG. 7, the tie-up, when the input shaft rotational speed N _T is lower than the normal value shown by a dotted line, the input shaft rotational acceleration dN _T becomes even lower than the reference rotation acceleration indicated by the dotted line , release-side hydraulic pressure P _B is corrected to decrease by feedback control based on the input shaft rotational acceleration (see dotted line). At this time, in a shift at a low throttle opening (small input torque), the shift is performed smoothly even after the tie-up is detected, but when the feedback control (correction) gain SdP _Tie is large, the hydraulic pressure is rapidly released by the feedback control. As a result, a shift shock may occur. Also, in a shift at a high throttle opening (a large input torque), if the correction gain SdP _Tie is small, the oil pressure drop due to the feedback control is slow, and the input shaft rotation change hardly occurs. Accordingly, in step 37, the correction gain SdP
By changing _{Tie based} on the input torque _Tt , the correction gain is reduced at a low throttle opening to prevent a shift shock due to a sudden drop in hydraulic pressure, and the correction gain is increased at a high throttle opening. Thus, a proper shift is performed while preventing a delay in the release-side hydraulic pressure drop.

[0043] Also, large current input shaft rotational acceleration dN _T with respect to the reference rotational acceleration W _ST, the feedback correction oil pressure dP _W is the amount of change N _{Stay up-negative} cases positive, calculated in the step S35 It is determined whether the change amount N _up of the input shaft rotation speed with respect to the output rotation speed is positive (S3).
8). If the correction hydraulic pressure dP _W is negative, that is, if engine blowing occurs, the release hydraulic pressure P _B is corrected in a direction to increase in step S39. That is, step S39
Is the same correction gain SdP _Tie as in step S37.
In addition to the correction by, as shown in FIG. 11, a hydraulic compensation amount dP _{Stay up-varying} by blowing amount N _{Stay up-the} input shaft speed change amount that is, the engine is added [P _B = fP _B (T
_{B) -dP W × SdP Tie +} dP up]. In the equation, the correction gain Sd is added to the feedback correction hydraulic pressure dP _W.
Although the oil pressure multiplied by P _Tie (dP _W × SdP _Tie ) is subtracted from the oil pressure [fP _B (T _B )] based on the release-side shared torque T _B , because the correction oil pressure dP _W becomes negative,
The release pressure P _B is corrected to increase side.

[0044] That is, the case where the rotational acceleration dN _T is smaller than the reference rotation accelerations W _ST (during tie-up), the correction oil pressure (dP _W × SdP _Tie) of the absolute value disengagement hydraulic pressure P
Subtracted from _B, also the case where the rotation acceleration dN _T is larger than the reference rotational acceleration W _ST (when engine racing), the correction oil pressure (dP _W × SdP _Tie) of the absolute value of the release-side hydraulic pressure P
Correct to add to _B. The above description is corrected positively hydraulic dP _W, is provided with the negative, the correction gain may be changed in time racing during tie-up the engine.

[0045] Thus, as shown in FIG. 8, on the basis of the sweep-down from the engagement pressure P _W of the disengagement side pressure P _B (with increasing engagement hydraulic pressure P _A), the input shaft to the output shaft speed When the rotation speed _NT rises above the normal value indicated by the dotted line,
Input shaft rotation acceleration d with respect to reference rotation acceleration indicated by dotted line
The change in _NT can be calculated, and the release-side hydraulic pressure P _B is mainly based on the feedback correction hydraulic pressure (dP _W × SdP _Tie ) based on the input shaft rotation acceleration as in the case of the tie-up. The hydraulic pressure correction amount (dP _up ) is added and corrected in the upward direction (see the dotted line).

As shown in FIG. 9, when the input shaft speed gradually increases from the stand-by control with respect to the reference input speed indicated by the dotted line, that is, when engine blowing occurs during the stand-by control, small change in axial rotation acceleration dN _T is unable to detect better the engine racing accuracy. Therefore, in such a case, the release side hydraulic pressure P _B is corrected in the ascending direction mainly by the hydraulic pressure correction amount dP _up based on the engine blowing amount N _up . At this time, the hydraulic pressure correction amount dP _up is
As shown in FIG. 1, when the engine blowing amount N _up exceeds a reference value, for example, 50 [rpm], it rises with a relatively gentle gradient, and when the engine blowing amount exceeds a predetermined value, it becomes relatively steep. It rises with a gentle gradient.

[0047] Therefore, as shown in FIG. 8, if the engine racing is caused by the sweep-down on the release side hydraulic pressure, the absolute value of the engine racing amount N _{Stay up-is} relatively small, the hydraulic pressure for the disengagement side pressure P _B in the feedback control Correction amount d
Since the ratio of P _up is small and the change amount of the input shaft rotational acceleration is relatively large, it is corrected mainly by the correction value (dP _W × SdP _Tie ) based on the input shaft rotational acceleration, and as shown in FIG. , if the engine racing from standby control occurs, the absolute value of the engine racing amount N _{Stay up-during} the initial control is relatively large, the ratio of the hydraulic compensation amount dP _{Stay up-for} disengagement hydraulic pressure P _B is larger, and the input shaft rotation Since the amount of change in acceleration is relatively small, it is mainly corrected by a correction value (dP _up ) based on the engine blowing amount.

If the feedback correction hydraulic pressure is negative in step S36 and the engine blowing amount is negative in step S38, the feedback control is not performed, and the release hydraulic pressure P _B is determined by the hydraulic pressure calculated in step S32. Descend.

The above steps S30 to S39 are the initial control by the feedback control.

The sweep-down of the disengagement side hydraulic pressure P _B by the feedback control is similar to the case of the engagement side, in which the input shaft rotation change amount ΔN is reduced to a predetermined rotation change start rotation speed d.
It continues until the N _S (S40). Then, the change δP _{E of the} release hydraulic pressure is set, and the sweep down is performed with the gradient due to the change in the hydraulic pressure (S41). The sweep down continues until the release hydraulic pressure P _B becomes 0 (S42). Is completed. Step S41 is release control.

In the above embodiment, the rotational acceleration of the input shaft and the rotational speed of the input shaft are used as means for detecting the engine blowing and the tie-up state. However, the present invention is not limited to this. Means for detecting a value, that is, another index such as acceleration, may be used.

[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 (gear ratio), a release hydraulic pressure and an engagement hydraulic pressure with respect to an output shaft rotation speed 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 time chart at the time of tie-up.

FIG. 8 is a time chart at the time of engine blowing.

FIG. 9 is a time chart at the time of engine blowing during standby control.

FIG. 10 is a diagram showing a correction gain changed by an input torque.

FIG. 11 is a diagram showing a hydraulic pressure correction amount that changes according to an engine blowing amount.

[Explanation of symbols]

U controller U ₁ the engagement side hydraulic control means U ₂ disengagement side hydraulic control means 1a input shaft rotational speed detecting means 1b input shaft rotational acceleration detecting means 1c feedback control means 9, 10 hydraulic servo P _A engagement hydraulic pressure P _B released Side hydraulic pressure N _T input shaft rotation speed dN _T input shaft rotation acceleration W _ST reference rotation acceleration N _up difference (engine blowing amount) dP _W (× SdP _Tie ) First correction amount SdP _Tie correction gain dP _up Second correction amount

 ──────────────────────────────────────────────────続 き Continuing from the front page (72) Inventor Masao Saito 10 Takane, Fujii-machi, Anjo City, Aichi Prefecture Inside Aisin AW Co., Ltd. (72) Masaaki Nishida 10 Takane, Fujii-machi, Anjo City, Aichi Prefecture Aisin・ AW Co., Ltd.

Claims (6)

[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. And a hydraulic servo that disconnects / engages the friction engagement elements. The hydraulic servo engages a first friction engagement element of the plurality of friction engagement elements and a second friction engagement element. , The upshift to the predetermined gear position is achieved, and the disengagement side shared torque calculated based on the engagement side oil pressure acting on the first frictional engagement element hydraulic servo,
In the hydraulic control device for an automatic transmission, which calculates a release hydraulic pressure acting on the hydraulic servo for the second frictional engagement element, both the first and second frictional engagement elements are in a disconnected state. Detecting means for detecting a tie-up state based on the engine blowing state and the friction engagement elements being both connected to each other, and torque phase control for controlling both the engagement side hydraulic pressure and the release side hydraulic pressure. And a feedback control unit that corrects the release hydraulic pressure with a correction amount based on a signal from the detection unit. 2. The method according to claim 1, wherein the detecting unit detects a rotational acceleration of the input shaft, and the feedback control unit calculates a difference between the rotational acceleration and a reference rotational acceleration, and performs a first correction based on the difference. The hydraulic control device for an automatic transmission according to claim 1, wherein the release-side hydraulic pressure is corrected by an amount. 3. The input control device according to claim 2, wherein the detecting unit detects an input shaft rotation speed in addition to the rotation acceleration of the input shaft, and the feedback control unit detects the input shaft rotation speed and a reference input shaft rotation speed. 3. The hydraulic control apparatus for an automatic transmission according to claim 2, wherein the difference is calculated, and a release amount hydraulic pressure is corrected by adding a second correction amount based on the difference to the first correction amount. 4. The method according to claim 1, wherein when the rotational acceleration is smaller than a reference rotational acceleration, the feedback control unit determines the first correction amount based on an absolute value of a hydraulic pressure calculated from a difference between the rotational acceleration and the reference rotational acceleration. Is corrected to be subtracted from the release-side hydraulic pressure, and when the rotational acceleration is greater than a reference rotational acceleration, the first value based on the absolute value of the hydraulic pressure calculated from the difference between the rotational acceleration and the reference rotational acceleration The hydraulic control apparatus for an automatic transmission according to claim 2, wherein the correction amount is corrected so as to be added to the release-side hydraulic pressure. 5. The first correction amount is calculated by multiplying a hydraulic pressure calculated based on a difference between the rotational acceleration and a reference rotational acceleration by a correction gain, and the correction gain is changed by an input torque. The hydraulic control device for an automatic transmission according to any one of claims 2 to 4. 6. The second correction amount is generated only when the input shaft rotation speed is larger than a reference input rotation speed by a reference value or more, and a difference between the input shaft rotation speed and a reference input shaft rotation speed is determined. The hydraulic control device for an automatic transmission according to claim 3, wherein the larger the value, the larger the value.