JP3528712B2 - Automotive gear change control device - Google Patents

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a shift control device for an automobile equipped with an engine and an automatic transmission, and in a power-on state in which power is transmitted from the engine in the wheel direction, a friction side engagement element and a release side engagement element. It is suitable for use in the case of re-grabbing the joint side frictional engagement element, that is, in the case of downshifting by a so-called clutch-to-clutch, and more specifically to control of engine output torque in shift control.

[0002]

2. Description of the Related Art Generally, a power-on downshift is performed when a driver runs out of torque and depresses an accelerator pedal, that is, when more torque is requested.

Conventionally, a control device for adjusting the engine output torque in the power-on downshift is proposed, for example, in Japanese Patent Publication No. 7-59904.
This type detects the turbine (transmission input shaft) rotation speed when it reaches a reference rotation speed that is set lower than the expected convergent rotation speed (rotation speed after gear shift) by a predetermined amount, and from that time, fuel injection is detected. The engine output torque is reduced by correcting the amount in the decreasing direction. And the fuel correction amount in this case,
That is, the engine output torque decrease correction amount is set according to the increase amount (change amount) of the turbine rotation speed, and the decrease of the engine output torque is kept constant until the expected convergence rotation speed of the turbine rotation speed.

[0004]

However, since the engine output torque decrease correction amount is a constant value that is set based on the turbine rotation speed increase amount at a predetermined time point, it does not occur in downshift shifting by clutch-to-clutch. For example, when there is a timing error in the disengagement-side and engagement-side frictional engagement elements, the constant engine output torque reduction correction amount cannot absorb an inappropriate change in output torque due to tie-up or engine blowing. . Specifically, when tie-up of the disengagement-side and engagement-side frictional engagement elements occurs, or when synchronization is shifted to the pre-gear gear stage side due to feedback control failure of the disengagement-side frictional engagement element, Since the output torque drops, a feeling of pulling in the torque is generated, and when the hydraulic servo of the engagement side frictional engagement element causes engine blowing due to insufficient stroke, etc., extra energy is supplied after the engine blowing. A large amount of output torque is output (that is, a peak is generated) at the time of shift output, which causes a large shift shock, which deteriorates the shift feeling in any case.

Therefore, an object of the present invention is to provide a shift control device for an automobile which solves the above-mentioned problems by correcting the engine output torque so as to correspond to the change in the output torque of the automatic transmission. To do.

According to the first aspect of the present invention, the input rotation from the engine output is changed in speed by switching the transmission path by disconnecting and contacting a plurality of friction engagement elements, and the changed rotation is output to the axle. A shift control device for an automobile, comprising an automatic transmission, downshifting the automatic transmission in a power-on state in which power is transmitted from an engine to a wheel , wherein an output torque of the engine during the downshift Engine control means for controlling (T _C ) (1
c), and a torque difference detecting means (1b) for detecting and calculating an excess amount and an insufficient amount of torque supplied from the engine to the automatic transmission during the downshift.
And the torque difference detecting means is provided in the automatic transmission.
A sensor (5) for detecting the input shaft speed (Nt),
The actual input shaft speed (N _U , N
_D ) and target input shaft speed (N
Based on the difference with _TA ) (ΔNd)
And the shortage amount are calculated, and during the downshift
In the case where the torque calculated based on the torque difference detection means is excessive, the excessive torque is reduced with respect to the reference value (T _ca ) of the engine output torque, and the calculated torque is insufficient. , The reference value (T
_(a ) is a shift control device for an automobile, which controls the engine control means so as to apply the insufficient torque to ( _ca ).

[0007]

According to a second aspect of the present invention (see, for example, FIG. 8 and FIG. 9), the reference value of the engine output torque is input based on the sensor (5) at a predetermined progress time (a1) of the downshift. The shift control device for a vehicle according to claim 1, wherein the shift amount is a reduction amount (T _CA ) calculated from an acceleration of a shaft rotation speed (Nt).

According to a third aspect of the present invention (see, for example, FIGS. 3 to 7), the downshift transmission releases one of the friction engagement elements and engages the other, and The hydraulic pressure (PA) of the hydraulic servo (9 or 10) for the frictional engagement element on the release side is converted into the input shaft rotation speed (Nt) by the sensor.
3. The vehicle transmission according to claim 1 or 2, wherein feedback control (S20) is performed based on the change of.

In the present invention according to claim 4 (see, for example, FIGS. 8 and 9), the engine control means (1c) continues correction of the engine output torque based on the torque difference detection means (1b) for a predetermined time. After that (S53, S54), the original engine output torque is restored by the sweep (S5
5), The shift control device for an automobile according to any one of claims 1 to 3.

[Operation] Based on the above configuration, the difference in the output torque from the engine with respect to the target value, that is, the difference in the energy supplied from the engine to the automatic transmission, is calculated based on, for example, the calculated input shaft rotational speed (N _TA ). , The difference (ΔNd) between the actual input speeds (N _U or N _D ) from the sensor 5 is multiplied by the acceleration gain and the engine inertia (S5).
3). In the power-on downshift shift, based on the engine output torque reference value at the time when the shift progresses by a predetermined amount (a1 [%]), for example, the input shaft rotation speed acceleration (dN _T ) at the time when the shift progresses by the predetermined amount. The engine output torque is controlled (S53) by adding or subtracting the correction value based on the energy difference to the reduction amount (T _ca ) that is set (S51).

For example, when the engine blows, the output torque from the engine is excessive, so the torque corresponding to the excess energy is corrected from the reference value (T _ca ) and is corrected. When the up occurs, the output torque from the engine becomes insufficient, so the reference value (T _ca ) is corrected so that the torque corresponding to the insufficient energy is added.

The reference numerals in parentheses are for the purpose of contrasting with the drawings, but are for convenience of understanding and speeding up the understanding, and thus have no influence on the structure of the claims. Not something to give.

[0014]

According to the first aspect of the present invention, in the downshift, in a situation in which excessive engine energy is supplied from the engine to cause engine blowing during downshift, a torque corresponding to the extra energy is used as the engine output torque. Since it is reduced from the reference value, the peak of output torque due to engine blowing can be suppressed to reduce shift shock, and in the situation where energy from the engine is insufficient due to tie-up, etc., the torque equivalent to the insufficient energy Is added to the engine output torque reference value, it is possible to suppress a drop in output torque due to tie-up and the like, reduce a feeling of pulling in (a feeling of braking), and improve a shift feeling.

Further, according to the present invention of claim 1 , the difference between the target value of the engine output torque and the target value of the engine output torque is calculated based on the difference between the target rotational speed of the input shaft and the actual rotational speed. Correction control can be performed.

According to the second aspect of the present invention, the engine torque reference value is a reduction amount set by the input shaft rotation speed acceleration, so that the correction based on the energy difference is performed in a state where engine blowing is suppressed in advance. Therefore, the engine torque control can be performed with high accuracy by quick response.

According to the third aspect of the present invention, in the clutch-to-clutch downshift, feedback control is performed on the hydraulic pressure on the disengagement side. Therefore, normally, the frictional engagement element is highly accurately repositioned. However, even if a feedback control error occurs, the above-described engine torque control can reduce the change in output torque due to poor synchronization timing, engine blowing, or tie-up, thereby reducing shift feeling deterioration. it can.

According to the present invention of claim 4 , since the engine output torque is repeatedly corrected for a predetermined time, it is possible to control so as to gradually reduce the peak of the output torque due to engine blowing and the drop due to tie-up. In addition, the original engine output torque can be restored by a sweep to mitigate the drastic change and improve the shift feeling.

[0019]

BEST MODE FOR CARRYING OUT THE INVENTION The present automatic transmission has a large number of friction engagement elements such as clutches and brakes, and an automatic transmission in which the transmission path of a planetary gear is selected by appropriately connecting and disconnecting these friction engagement elements. A mechanism (not shown) is provided, the input shaft of the automatic transmission is connected to the engine output shaft via a torque converter, and the output shaft is connected to the drive wheels. Specifically, this automatic transmission is applied to the fifth forward speed and the first reverse speed disclosed in Japanese Patent Laid-Open No. 9-21448.

FIG. 1 is a block diagram showing an electric control system. Reference numeral 1 is a control unit (ECU) composed of a microcomputer, which is an engine rotation sensor 2 and an accelerator pedal for detecting an accelerator pedal depression amount of a driver. Opening sensor 3, sensor 4 for detecting the actual throttle opening in the engine, transmission (automatic transmission mechanism)
Each signal from the sensor 5, which detects the input shaft speed (= turbine speed), the vehicle speed (= automatic transmission output shaft speed) sensor 6, and the oil temperature sensor 7, is input, and the engine throttle It outputs to the electronic throttle system (engine operating means) 8 to be controlled and the linear solenoid valves (pressure adjusting means) SLS and SLU of the hydraulic circuit. The control unit 1 uses the linear solenoid valve SLS or S.
A hydraulic pressure control means 1a for transmitting a pressure adjusting signal to the LU and an engine control means 1c for transmitting a throttle opening command to the electronic throttle system 8 are further provided, and the engine control means further inputs to an automatic transmission (travel system). When the energy supplied from the engine to the automatic transmission is larger than the target value such that the engine blows, the engine output torque is equal to the calculated energy. Correct in a decreasing direction,
Further, in a situation where the energy supplied from the engine to the automatic transmission is insufficient due to tie-up or the like, a signal for correcting the engine output torque in the direction of adding the torque corresponding to the calculated energy is output. It has a torque difference detection means 1b.

FIG. 2 is a diagram showing an outline of the hydraulic circuit.
In addition to having the two linear solenoid valves SLS and SLU which constitute the hydraulic control means 1a, the transmission path of the planetary gear unit of the automatic transmission mechanism is switched to achieve, for example, the fourth or fifth forward speed and the first reverse speed. It has a plurality of hydraulic servos 9 and 10 for engaging and disengaging a plurality of friction engagement elements (clutch and brake). The solenoid modulator pressure is supplied to the input ports a ₁ and a ₂ of the linear solenoid valves SLS and SLU, and the output port b of these linear solenoid valves is supplied.
The control oil pressures from ₁ and b ₂ are supplied to the control oil chambers 11a and 12a of the pressure control valves 11 and 12, respectively. Pressure control valves 11, 12
Are supplied to the input ports 11b and 12b, respectively, and the pressures regulated by the control hydraulic pressures from the output ports 11c and 12c are shifted to the shift valve 1 respectively.
It is appropriately supplied to the respective hydraulic servos 9 and 10 via 3, 15.

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, and actually,
A large number of hydraulic servos are provided corresponding to the automatic transmission mechanism, and a large number of shift valves for switching the hydraulic pressure to these hydraulic servos are also provided.

Next, the power-on / downshift will be described with reference to FIG. 3. First, the control of the disengagement hydraulic pressure PA will be described with reference to FIGS. 4 and 5. In addition, specifically, a downshift (kick down) in which the driver depresses the accelerator pedal to request torque,
2 shows the state of shifting to two gears, and therefore the release side frictional engagement element
It is a C3 clutch, and the hydraulic pressure PA of the hydraulic servo is
(Dedicated for pressure regulation) Pressure regulation is controlled by the linear solenoid valve SLS.

Throttle opening sensor 3 and vehicle speed sensor 6
When the control unit 1 determines the downshift from the map based on the signal from, the timing is started and the shift control is started after a predetermined delay time from the shift determination (S1). At the start time (t = 0), the disengagement hydraulic pressure PA is the engagement pressure, and the disengagement side frictional engagement element is in the engaged state. Then, the disengagement side torque T _A is calculated by a function of the input torque Tt (S2). For the input torque Tt, the engine torque is calculated based on the throttle opening and the engine speed based on a map, the speed ratio is calculated from the input / output speed of the torque converter, and the torque ratio is calculated based on the speed ratio on the map. It is obtained by multiplying the torque by the torque ratio. Further, the disengagement side torque T _A is obtained by the torque sharing rate or the like being involved in the input torque.

The release side standby engagement pressure Pw is calculated from the release side torque T _A (S3), and a control signal is output to the linear solenoid valve so that the release side hydraulic pressure PA becomes the standby engagement pressure Pw ( S4), the control of the hydraulic pressure on the release side based on the input torque or the like is continued until a predetermined time tw has elapsed (S4).
5). The standby control is performed from steps S2 to S4, and the standby control time tw is changed by the input torque Tt.

Then, the predetermined release side hydraulic pressure P _AS and the release side torque T _A are calculated in the same manner as described above (S7, S8), and the target hydraulic pressure P _TA is calculated based on the release torque T _A (S9). Furthermore, the margin rate (tie-up degree) S ₁₁ , S ₂₁
Thus, the release side target hydraulic pressure P _TA is calculated in consideration of the drive feeling (S10). The margin rate is
This is obtained from a large number of throttle opening / vehicle speed maps selected according to the difference in oil temperature. Generally, S ₁₁ > 1.
0, S ₂₁ > 0.0.

Furthermore, the preset time t _TA, the gradient to the target hydraulic pressure _{P TA, [(P AS -P} TA) /
t _TA ], and sweep down is performed according to the gradient (S11). That is, in the power-on state, the sweep down with a relatively steep gradient is performed until the release side hydraulic pressure PA reaches the target hydraulic pressure P _TA just before the start of the inertia phase (S12). Then, the release side hydraulic pressure change δP _TA is calculated by the function [δP _TA = fδ _PTA (ωa)]
Is calculated based on (S13). Note that ωa is a target target input shaft rotation change rate (target rotation acceleration) at the start of rotation change of the input shaft rotation speed (gear ratio) N with respect to the output shaft rotation speed. Then, the (second) sweep down is performed with a gradient according to the hydraulic pressure change δP _TA (S1
4) In the power-on state, the sweep down is continued from the input shaft rotation speed N _TS before the shift start to the shift start determination rotation speed at which the rotation change amount ΔN is detected with a predetermined accuracy (S15). ). The steps S7 to S14 are the initial shift control, and the disengagement side frictional engagement element reduces its torque capacity, but the shift is not proceeding.

Then, the sweep down is performed at a gradient according to a predetermined hydraulic pressure change δP _I which is a preset relatively low gradient (S16). In the sweep down, when the release side hydraulic pressure PA is larger than the load pressure of the return spring in the power-on state, that is, when the torque capacity of the release side hydraulic servo does not become 0, the shift is started (rotational change start). The process is performed up to aF [%] of the total rotational speed change amount until completion, that is, a predetermined shift progress degree (S18). Incidentally, the shift progress degree, an input shaft speed during rotation change start N _TS,
The amount of change in the input shaft speed (the amount of change in the input shaft speed with respect to the output shaft speed due to constant rotation) based on the gear ratio from the start of rotation change to the present is ΔN, and the gear ratio before shifting is g _i ,
If the gear ratio after shifting is g _{i + 1} , then [(ΔN × 100) / (N _TS / g _i ) · (g _{i + 1} −g
_i )]. The sweep down with the gradient δP ₁ becomes inertia phase control, and the change of the input shaft speed N _T based on the gear ratio is started.

Then, when a predetermined shift progress rate aF [%], for example, 20 [%] at which the change of the input shaft rotation speed N _T is stabilized, downshift feedback control (S20) is performed. The feedback control is controlled at each shift progress stage so that the difference between the actual input shaft (turbine) rotational speed change rate (acceleration) and the target input shaft rotational speed change rate is minimized. . At this time, the gain set at each stage of the control may be corrected based on the speed ratio of the torque converter (Japanese Patent Application No. 10-3).
16621). The feedback control is continued until the gear shift progress degree reaches a2 [%], for example, 90 [%] in the vicinity of the full rotation change rotation speed of the gear ratio at which the downshift is completed (S21). It should be noted that until the end of the servo start control time t _{SE due} to the relationship with the control of the engagement side hydraulic pressure described later (S23),
And until the engagement side pressure PB is greater than the target pressure P _TB (S24), the feedback control (S20) is continued. The step S20 is feedback control.

When the shift to the above a2 [%] is completed, a predetermined hydraulic pressure change δP _FA having a relatively steep gradient is set, and sweep down is performed at the gradient (S25).
When the release side hydraulic pressure PA becomes 0, the release side hydraulic pressure control at the downshift is completed (S26). Step S above
25 is the completion control.

Next, the control of the engagement side hydraulic pressure PB in the downshift will be described with reference to the flowcharts of FIGS. 6 and 7 and the time chart of FIG. In addition, specifically, as described above, the 4-2 downshift is performed. Therefore, the engagement side frictional engagement element is the B5 brake, and the hydraulic pressure PB of the hydraulic servo is linear (for lockup control). Pressure regulation is controlled by the solenoid valve SLU.

First, timing is started based on a downshift command from the control unit 1 (S30), and a predetermined signal is output to the linear solenoid valve SLU so that the engagement side hydraulic pressure PB becomes the predetermined pressure P _S1 (S31). . The predetermined 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 (S32), the engagement side hydraulic pressure PB sweeps down at a predetermined gradient [(P _S1 −P _S2 ) / t _SB ] (S3).
3) When the engagement side hydraulic pressure PB becomes the predetermined low pressure P _S2 (S3
4) The sweep down is stopped and the predetermined low pressure P _S2 is maintained (S35). 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 rotation change of the input shaft, and the predetermined low pressure P _S2 is held until the time t reaches a predetermined time t _SE (S36). ). Servo activation control is performed from steps S31 to S36.

Next, the engagement side torque T _B is changed to the release side hydraulic pressure P.
Function of A and an input torque _{Tt [T B = f TB (} PA, T
t)] (S37), and in consideration of the margin rate, the engagement side torque T _B is [T _B = S _1D × T _B + S
_2D ] is calculated (S38). Then, the engagement side hydraulic pressure PB is calculated from the engagement side torque T _B [PB = f
_{PB (T B)] (S39} ). Steps S37 to S39
Is the engagement control. Then, the control by the engagement-side hydraulic pressure PB based on the engagement-side input torque T _B (depending on the disengagement-side hydraulic pressure PA and the input torque Tt) at step S39 is a1 [%] of the total shift progress degree of the downshift, For example, it continues up to 70% (S40). That is, _letting N _{TS be} the input shaft rotation speed at the start of gear shifting, ΔN being the amount of rotation change, g _i being the gear ratio before gear shifting, and g _{i + 1} being the gear ratio after gear shifting, [(ΔN × 10
0) / (N _TS / g _i ) · (g _{i + 1} −g _i )] is a1
Continue until [%] is reached.

In step S40, if the a1 [%] of the total shift progress is exceeded, the final control is started. First, the engagement side target pressure P _TB is calculated from the engagement side input torque T _B (S4
1) Further, the engagement side oil pressure PB at the time point of the rotation change amount a1 [%] is stored as P _LTB (S42). As a result, the predetermined slope [(P _TB −P _LSB ) / t _LE ] is calculated by the predetermined time t _LE set in advance, and the sweep up is performed at the relatively gentle slope (S43). , Is continued until the engagement side hydraulic pressure reaches the target hydraulic pressure P _TB (S44). Further, a predetermined gradient δP _LB is set, and sweeping up is performed at this gradient (S46). The sweep-up has a shift progress degree of a2 [%], for example, 90%.
Continue to [%] (S47). The final control is performed from steps S41 to S46.

Furthermore, the end time t _F of the final control is set (S48), a relatively steep gradient δP _FB is set, and sweep up is performed at this gradient (S49).
The completion control time t _{FE is} continued (S50). The gradient δP _FB
If the power is on, the sweep up of step S2
The steep slope is set according to the release side hydraulic pressure δP _FA . The steps S48 and S49 are the completion control.

Next, the engine torque control, which is the main part of the present invention, will be described with reference to FIGS. As described above, feedback control of the release hydraulic pressure PA (S
20), the input speed N _T rises, and the amount ΔN of change of the input speed from the start of control (N _TS ) is the predetermined value a1 [%], for example 70 [%] (see S40). ), That is, when the progress of the shift by the disengagement hydraulic pressure PA is almost completed and the engagement hydraulic pressure PB is in the vicinity of the transition from the engagement control to the final control, the engine torque down control is activated. (S50). If the torque down timing is always constant, the amount of heat generated by the disengagement side frictional engagement element also increases when the number of rotations at the start of gear shifting is large, that is, when the amount of rotation change during gear shifting is large.
If the torque down timing is delayed, the durability of the friction material may be impaired. However, in the present engine torque control, the start time of the torque down is the input rotational speed N _TS (ΔN = Since it is set based on 0) (ΔN ≧ a1), it is possible to prevent deterioration of the durability of the disengagement side frictional engagement element from high vehicle speed to low vehicle speed.

Then, the rate of change of the input rotational speed at the predetermined value a1 [%], that is, the acceleration dN ₁ is calculated, and based on the rate of change, the torque reduction amount T _ca is calculated in accordance with a predetermined relationship such as a direct proportional function. S51). Furthermore, after the control amount T _c of the engine torque at that time is virtually set to 0 (S52), the control amount is controlled. The control amount T _c of the engine torque is set to a predetermined value a1 of the input shaft speed as a correction amount by calculating energy supplied excessively by engine blowing or the like or insufficient energy due to tie-up (output torque). The correction amount calculated above is subtracted from the torque reduction amount T _ca set by the calculation (S53). That is, the actual rotation speed of the input shaft rotation speed Nt is detected by the input shaft rotation speed sensor 5, and the actual rotation speed (in FIG. 9, the rising side due to blowing is represented as N _U, and the falling side due to tie-up is N _D ), the target input shaft rotation speed calculated based on the vehicle speed sensor 6, the gear ratio (4-2 gear ratio), the input torque from the throttle opening sensor 3, etc. (the rotation based on the gear position). A straight line from the start to the end of change) Difference from N _TA Δ
Calculate Nd. The difference ΔNd ₁ between the torque reduction amount and the rising side N _U acts as a plus side,
The difference ΔNd _{2 from} the descending side N _D acts as the minus side. Then, the rotation difference ΔNd [= N _TA −N _U (or N
_D )] is multiplied by the acceleration gain and the amount of engine inertia to be converted into energy (that is, output torque by a coefficient), which corresponds to the difference between the target energy supplied from the engine at that time and the energy actually supplied. The calculated torque difference is calculated, and this value is subtracted from the set torque reduction amount T _ca to calculate the engine control amount T _c .

Then, the torque reduction amount T _ca based on the energy difference from the engine in step S53 described above.
The above correction is repeatedly performed until the elapsed time t _E from the time when the shift progress becomes a1 [%] has passed a preset predetermined time t ₁ . The predetermined time t ₁ is at the end of the feedback control of the disengagement side hydraulic pressure PA (step S20) and the final control of the engagement side hydraulic pressure PA (step S46), that is, when the shift progress degree ΔN of the input shaft rotational speed is a2.
[%] (Steps S21 and S27, eg 90
[%]), The downshift is almost achieved, and the input shaft speed reaches the low speed side (2nd speed) gear stage, or the input shaft speed becomes stable in consideration of the increase in the input shaft speed due to acceleration. The time t _E from the start of the engine control control is set to correspond to the predetermined time t.
_{When 1} is exceeded (T _E > t ₁ ), the above-described correction control based on the energy supplied to the engine is stopped, and the release side hydraulic pressure PA (and the engagement side hydraulic pressure PA) is also subjected to completion control (step S24). , S48) is started (S5
4).

Then, the engine control T _c is swept up at the preset time t ₂ with the reduction amount corrected in step S53 (S5).
5) While avoiding a drastic change in engine output, the engine output returns to the normal value based on the accelerator pedal opening sensor 3 of the driver (S56), and the engine torque control ends.

The output torque T _O of the automatic transmission is held at a substantially constant value by the feedback control of the hydraulic pressure PA on the release side until the downshift (for example, 4-2 shift) is almost completed. When the rotation speed N _T increases and approaches the rotation speed to the low speed side gear stage (for example, the second speed) due to the downshift, the rotation speed increases relatively rapidly and increases excessively until it converges to the torque value after the downshift. Peak torque tends to occur. Therefore, when the predetermined shift progresses before the downshift is completed (ΔN = a1 [%]), the output from the engine is reduced by the reduction amount T _ca calculated by the input shaft acceleration dN ₁ at that time. As a result, generation of peak torque due to the output torque is suppressed.

However, the final control (S41 to S46) of the engagement-side hydraulic pressure PB is delayed, or the feedback timing (S20) of the release-side hydraulic pressure PA is missed so that the synchronization timing of the release side and the engagement side is delayed. Or
Engine blowing may occur, but in such cases,
The above-mentioned constant value of torque reduction amount T _ca does not make it in time, and the output torque T _O causes a large blow and a large shift shock, as indicated by a thin dashed line. Therefore, in the present invention, the torque corresponding to the extra energy calculated based on the difference ΔNd ₁ between the target rotation speed N _TA of the input shaft and the actual rotation speed N _U is relative to the reduction amount T _ca. The engine torque T _c is added as shown in FIG.
_{The ca} is corrected to be large, and the energy supply from the engine that causes engine blowing is reduced. As a result, the output torque T _O has a low peak amount on the rising side, as indicated by the thick chain line. Become.

On the other hand, the torque capacity of the disengagement side frictional engagement element increases before the torque capacity of the disengagement side frictional engagement element is reduced, and tie-up occurs, or feedback control of the disengagement side hydraulic pressure (S20). ) If the synchronization timing of the disengagement side and the engagement side shifts to the pre-shift gear stage side (the side where the gear ratio is not established) due to a mistake, and the energy supplied from the engine is excessively consumed, the above constant value At the torque reduction amount T _ca , the energy is insufficiently supplied, and the output torque T _o is excessively lowered as shown by a thin dotted line, which gives the driver a discomfort such as braking. Therefore, in the present invention,
Difference ΔN between the target speed N _TA of the input shaft and the actual speed N _D
The torque corresponding to the insufficient energy calculated based on d ₂ is subtracted from the reduction amount T _ca , and the engine torque T _c is corrected so that the reduction amount becomes smaller as shown by the dotted line, and the insufficient energy is reduced. compensates for which the output (output) torque T _O, as indicated by a thick dotted line, to reduce the drop in torque.

Incidentally, the releasing-side and engaging-side hydraulic pressure control is not limited to the above-mentioned embodiment, and needless to say, may be other control. Further, in the engine torque control, the reduction amount based on the input shaft acceleration is used as a reference value and correction is made for this, but the engine based on the normal driver accelerator pedal opening is set without setting the reduction amount. The torque may be used as a reference value and the torque may be corrected.

[Brief description of 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, a disengagement side hydraulic pressure, and an engagement side hydraulic pressure command based on a gear ratio showing a power-on downshift.

FIG. 4 is a flowchart showing control of downshift release hydraulic pressure.

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

FIG. 6 is a flowchart showing control of engagement side hydraulic pressure of downshift.

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

FIG. 8 is a flowchart showing engine torque control according to the present invention.

FIG. 9 is a time chart relating to engine torque control.

[Explanation of symbols]

1 control unit 1a hydraulic control unit 1b torque difference detecting means 1c engine control unit 5 input shaft rotation sensor 9 hydraulic servo Nt input shaft rotational speed N _TA target value (input shaft rotational speed target value) N _D, N _U actual Input shaft speed ΔNd Difference T _c Torque control amount T _ca Reference value (reduction amount) a1 Predetermined progress T _o Output (output) torque PA Release hydraulic pressure PB Engage hydraulic pressure

Continuation of the front page (56) Reference JP 10-278635 (JP, A) JP 61-119432 (JP, A) JP 2-3545 (JP, A) (58) Fields investigated (Int .Cl. ⁷ , DB name) F16H 59/00-63/48 B60K 41/06

Claims (4)

(57) [Claims] Claim: What is claimed is: 1. An engine, comprising: an automatic transmission that shifts input rotation from an engine output by switching a transmission path by disconnecting and contacting a plurality of friction engagement elements and outputting the shifted rotation to an axle. In a power-on state in which power is transmitted from the vehicle to the wheels, downshifting the automatic transmission, a shift control device for an automobile, comprising: engine control means for controlling the output torque of the engine during the downshifting A torque difference detecting means for detecting and calculating an excess amount and an insufficient amount of torque supplied from the engine to the automatic transmission during the downshift , and the torque difference detecting means is configured to Transmission input shaft speed
Is equipped with a sensor that detects
Target input shaft that is determined by the input shaft speed of the
Based on the difference from the rotation speed, the excess and shortage of the torque
The result by computing, during the the downshift, when the calculated torque based on the torque difference detecting means extra, so as to reduce the該余amount of torque to the reference value of the engine output torque and said A shift control device for a vehicle, wherein the engine control means is controlled so as to add the insufficient torque to the reference value when the calculated torque is insufficient. Wherein the reference value of the engine output torque, said a reduction amount calculated by the acceleration of the input shaft rotational speed based on the sensor at a predetermined progress of the downshift, according to claim 1 car of Shift control device. 3. The downshift shift releases one of the friction engagement elements and engages another one of the friction engagement elements, and the hydraulic pressure of a hydraulic servo for the friction engagement element on the release side is set to the sensor. The shift control device for an automobile according to claim 1 or 2, wherein feedback control is performed based on a change in input shaft rotation speed due to. 4. The engine control means restores the original engine output torque by sweeping after continuing correction of the engine output torque based on the torque difference detection means for a predetermined time. Or a vehicle speed change control device as described above.