JP2002039362A - Control device of automatic transmission for vehicle - Google Patents

Abstract

PROBLEM TO BE SOLVED: To conclude racing at shifting with excellent responsiveness for an automatic transmission shifting by switching an engagement condition with a frictional engagement element for high-speed stage and a frictional engagement element for low-speed stage. SOLUTION: A reference turbine rotating speed is obtained from a gear ratio before shifting and an output shaft rotating speed of a transmission mechanism. When a turbine rotating speed exceeds the reference turbine rotating speed than predetermined, the generation of the racing is detected. When the generation of the racing is detected, a maximum correction oil pressure PHOSMAX set on the basis of an input shaft torque is made to be an initial value and a correction amount PHOS gradually reducing with a differential value of deviation of the turbine rotating speed and the reference turbine rotating speed, in other words, a time constant Ts corresponding to change speed of the racing is calculated. Then, the correction amount PHOS is added to an indicated oil pressure at engagement side.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a control device for an automatic transmission for a vehicle, and more particularly, to shifting by switching an engagement state between a high-speed friction engagement element and a low-speed friction engagement element. TECHNICAL FIELD The present invention relates to a technique for suppressing occurrence of idling at the time of shift control in an automatic transmission.

[0002]

2. Description of the Related Art Conventionally, there has been known an automatic transmission having a structure in which a shift is performed by switching an engagement state between a high speed friction engagement element and a low speed friction engagement element. If the engagement is too fast relative to the release of the friction engagement element, the high speed friction engagement element and the low speed friction engagement element are in an overlapping state (hereinafter referred to as an interlock). Then, the output shaft torque drops (torque reduction), and conversely, if the fastening is too slow relative to the release, the rotation will increase (idle blowing).

Conventionally, a shift characteristic is set so that either the interlock or the idling occurs, and a shift control is switched (a phase is switched) based on the state of the interlock or the idling. At the time of an upshift, the frictional engagement element on the engagement side is engaged after waiting for the rotation to decrease while maintaining the tendency toward idling (see Japanese Patent Application Laid-Open No. 9-264412).

[0004]

However, as described above, in the case of the structure in which the frictional engagement element on the engagement side is engaged during the upshift while waiting for the rotation to be reduced while the tendency toward idling occurs, If the blowing is not converged early, the shift will be delayed, and control to converge the idling is required.

In the technique disclosed in Japanese Patent Application Laid-Open No. 9-264412, when the amount of idling (the amount of slip) is large,
In order to achieve the expected state of minute air blowing (small slip),
The hydraulic pressure of the release-side friction engagement element is feedback-controlled. However, even if feedback correction is performed in accordance with the amount of airflow (slip amount), it takes time for the actual oil pressure to reach the instructed oil pressure due to the response delay of the valve that controls the oil pressure.
There is a possibility that an interlock may occur, and if the feedback gain is suppressed low in order to avoid the occurrence of the overshoot, there is a problem that the convergence of the blowing is delayed.

SUMMARY OF THE INVENTION The present invention has been made in view of the above-mentioned problems, and an object of the present invention is to provide an automatic transmission for a vehicle capable of rapidly changing an oil pressure to quickly converge an idling and avoiding occurrence of overshoot. It is an object to provide a control device.

[0007]

Therefore, according to the first aspect of the present invention, when the occurrence of idling is detected at the time of shifting, it rises stepwise to a predetermined value, and then decreases at a speed corresponding to the changing speed of idling. The changing correction value is added to at least one of the oil pressure of the frictional engagement element on the release side and the oil pressure of the frictional engagement element on the engagement side.

According to this configuration, when an idling occurs, a correction value is added to at least one of the oil pressure (instruction pressure) of the frictional engagement element on the release side and the oil pressure (instruction pressure) of the frictional engagement element on the engagement side. The correction value is set so as to rise stepwise from 0 when the occurrence of the blowing is detected, and is set to decrease with the lapse of time thereafter. Is set according to the speed of change of the idling (acceleration of the input shaft rotation speed).

[0009] In the invention according to claim 2, the predetermined value is:
The configuration is determined in accordance with the input shaft torque of the transmission mechanism. According to this configuration, according to the input shaft torque of the speed change mechanism indicating the transmission torque capacity required for the friction engagement element,
A predetermined value that is an initial value of the correction value is determined. According to the third aspect of the invention, the predetermined value is a value obtained by subtracting the reference oil pressure of the friction engagement element from the maximum oil pressure set in accordance with the input shaft torque of the transmission mechanism.

According to this configuration, the result obtained by subtracting the reference oil pressure, which is the oil pressure (instruction pressure) set by the normal control, from the maximum oil pressure according to the input shaft torque at that time is determined by the predetermined value that is the initial value of the correction value. By setting this value, the command oil pressure gradually decreases from the maximum oil pressure and returns to the reference oil pressure. According to a fourth aspect of the present invention, the predetermined value is determined in accordance with the speed of change of the idling and / or the engine speed.

According to this configuration, the initial value of the correction value is determined from the speed of change in the speed of the blown air (the rising speed of the input shaft rotation speed) and / or the speed of the engine at the time when the blown air is detected. According to a fifth aspect of the present invention, the correction value is decreased and changed with a first-order delay from the predetermined value, and the time constant of the first-order delay is set in accordance with a change speed of the idling.

According to this configuration, the correction value is changed to a predetermined value in a stepwise manner at the time of detection of the idling, and then gradually attenuated with a first-order lag. It is set according to the speed of change of the blowing (the rising speed of the input shaft rotation speed). In the invention according to claim 6, the differential value of the deviation between the reference input shaft rotation speed calculated from the gear ratio before the shift and the output shaft rotation speed of the transmission mechanism and the input shaft rotation speed of the transmission mechanism is calculated as the differential value. The calculation was performed as a parameter indicating the change speed of the blowing.

According to this configuration, the input shaft rotational speed (turbine rotational speed) when the disengagement side frictional engagement element holds the engaged state is determined from the gear ratio before the gear shift and the output shaft rotational speed of the transmission mechanism. And the difference between this reference value and the input shaft rotation speed increased by the occurrence of the idling is determined as the increased rotation, and the differential value of the increased rotation is defined as the changing speed of the idling.

According to the present invention, the input shaft rotation speed of the transmission mechanism is higher than a reference input shaft rotation speed which is calculated from the gear ratio before the gear shift and the output shaft rotation speed of the transmission mechanism. Sometimes, it is configured to determine the occurrence of an idling. According to this configuration, the input shaft rotation speed (turbine rotation speed) when the disengagement side frictional engagement element holds the engaged state is determined as a reference value, and the actual input shaft rotation speed is higher than a predetermined value by more than the reference value. When it becomes higher, it is determined that an air blow has occurred.

[0015]

According to the first aspect of the present invention, when the occurrence of idling is detected, the indicated hydraulic pressure is raised in a stepwise manner, so that the valve for controlling the hydraulic pressure is driven with the maximum response and quickly. The hydraulic pressure (transmission torque capacity) can be increased, and the hydraulic pressure increase correction amount is attenuated at a speed corresponding to the speed of change of the idling. There is an effect that the hydraulic pressure can be increased and corrected for a longer time to converge the idling.

According to the second and third aspects of the invention, the maximum value of the hydraulic pressure to be increased and corrected can be limited based on the input shaft torque (required transmission torque capacity). There is an effect that can be avoided. According to the fourth aspect of the present invention, when a sudden idling occurs, the hydraulic pressure is increased and corrected to be larger, so that the idling can be surely converged while avoiding excessive correction.

According to the fifth aspect of the invention, it is possible to make a correction corresponding to the characteristic of the actual oil pressure which rises with a delay with respect to the step change of the command oil pressure, and it is longer when the speed of change of the idling is large. There is an effect that the time hydraulic pressure can be increased and corrected, and the idling can be surely converged.
According to the sixth and seventh aspects of the present invention, it is possible to accurately determine the occurrence of the blowing by using the input shaft rotation speed in a state where the blowing does not occur as a reference, and to determine the changing speed of the blowing. This has the effect of making accurate judgments.

[0018]

Embodiments of the present invention will be described below. FIG. 1 shows a transmission mechanism of an automatic transmission according to an embodiment, in which an output of an engine is transmitted to a transmission mechanism 2 via a torque converter 1.

The transmission mechanism 2 includes two sets of planetary gears G1,
G2, 3 sets of multiple disc clutches H / C, R / C, L / C, 1
A set of brake bands 2 & 4 / B, a set of multiple disc brakes L & R / B, and a set of one-way clutch L / OWC. The two sets of planetary gears G1 and G2 are simple planetary gears including sun gears S1 and S2, ring gears r1 and r2, and carriers c1 and c2, respectively.

The sun gear S1 of the planetary gear set G1 is configured to be connectable to the input shaft IN by a reverse clutch R / C, and is configured to be fixed by a brake band 2 & 4 / B. Sun gear S2 of the planetary gear set G2
Are directly connected to the input shaft IN. The carrier c1 of the planetary gear set G1 is configured to be connectable to the input shaft I by a high clutch H / C, while the ring gear r2 of the planetary gear set G2 is connected to the planetary gear set G1 by a low clutch L / C.
And the carrier c1 of the planetary gear set G1 can be fixed by the low & reverse brake L & R / B.

A ring gear r1 of the planetary gear set G1 and a carrier c2 of the planetary gear set G2 are directly connected to the output shaft OUT. In the transmission mechanism 2 having the above-described configuration, the first to fourth speeds and the reverse are realized by a combination of engagement states of the clutches and brakes, as shown in FIG. In FIG. 2, a circle indicates the engaged state, and a part without a symbol indicates the released state. In particular, the engaged state indicated by a black circle of the low & reverse brake L & R / B at the first speed. Indicates the fastening only in one range.

According to the combination of the engaged states of the clutches and brakes shown in FIG. 2, for example, at the time of the downshift from the fourth speed to the third speed, the release of the brake band 2 & 4 / B (high-speed gear friction engagement element) is performed. Do both low clutch L
/ C (friction engagement element for low speed gear) is engaged, and the brake band 2 & 4 /
B (low speed gear friction engagement element) is released and the high clutch H / C (high speed gear friction engagement element) is engaged. A configuration in which the shift is performed by switching the engagement state with the low-speed gear friction engagement element is referred to as a shift change.

Each of the clutches and brakes (friction engagement elements) is operated by a supply hydraulic pressure.
The supply hydraulic pressure for each clutch / brake is adjusted by various solenoid valves included in the solenoid valve unit 11 shown in FIG. An A / T controller 12, which controls various solenoid valves of the solenoid valve unit 11, includes an A / T oil temperature sensor 13, an accelerator opening sensor 14, a vehicle speed sensor 15, a turbine rotation sensor 16, an engine rotation sensor 17, an air flow meter. 18
And the like, and the engagement hydraulic pressure in each friction engagement element is controlled based on the detection results.

In FIG. 3, reference numeral 20 denotes an engine combined with the automatic transmission. Here, A /
The state of the shift change by the T controller 12 will be described below with reference to a time chart of FIG. 4 by taking an example of an upshift (hereinafter, referred to as a power-on upshift) in a state where the driving torque of the engine is applied. explain.

The flowchart of FIG. 5 shows a main routine of hydraulic pressure (transmission torque capacity) control common to the engagement-side friction engagement element and the release-side friction engagement element. In step S1, a shift determination of a power-on upshift is performed. A / T
The controller 12 previously stores a shift map in which a shift stage is set according to the vehicle speed VSP and an accelerator opening (throttle opening). For example, a search is made from the current (pre-shift) shift stage and the shift map. If the shift speed is different from the set shift speed, the shift speed is in the upshift direction, and the accelerator is not fully closed, it is determined as a power-on upshift.

When the shift of the power-on upshift is determined, the process proceeds to step S2, and the occurrence of idling is detected based on the input shaft rotation speed (turbine rotation speed) Nt of the transmission mechanism. Specifically, the output shaft rotation speed No of the transmission mechanism
[Rpm] is the gear ratio before shifting (gear ratio = turbine rotation Nt)
/ Output shaft rotation speed No) to obtain a reference turbine rotation speed, and the input shaft rotation speed (turbine rotation speed) Nt [rpm] of the speed change mechanism exceeds the reference turbine rotation speed + hysteresis value HYS. At times, it is determined whether or not the wind blows.

In a preparation phase described later, the oil pressure of the disengagement side frictional engagement element is gradually decreased and the oil pressure of the engagement side frictional engagement element is gradually increased.
By setting the release control to be performed relatively quickly with respect to the engagement control, an idling is generated. Based on the occurrence of the idling, switching from the preparation phase to the torque phase is determined. It is.

In step S2, the turbine rotation speed Nt
[Rpm] ≦ reference turbine speed + hysteresis value HY
When it is determined to be S, a preparation phase process of step S3 is executed. The preparation phase process in step S3 is divided into a release-side process and a fastening-side process.

The flowchart of FIG. 6 shows the main routine of the preparation phase process for the disengagement-side friction engagement element. In step S31, the type of the shift, the type of the friction engagement element to be disengaged, and the oil temperature are determined. Then, it is determined whether or not a predetermined time TIMER1 stored in advance has elapsed from the shift determination. If within the predetermined time TIMER1,
Proceeding to step S32, the calculation of the release initial hydraulic pressure is performed. The release initial hydraulic pressure is an initial pressure for performing release control, and is a predetermined time T from the hydraulic pressure during non-shifting to the release initial hydraulic pressure.
It should be reduced within IMER1.

The calculation of the initial disengagement hydraulic pressure in step S32 is shown in detail in the flowchart of FIG. 7. In step S321, the non-shifting oil pressure Po0 (instruction pressure) of the friction engagement element for which the present release control is performed is calculated. I do. The non-shift hydraulic pressure Po0 is expressed as Po0 = K1 × (Tt × Tr-o) × room allowance initial value + Prtn−
Calculated as o.

Here, K1 is a coefficient for converting the transmission torque capacity of the disengagement side frictional engagement element into hydraulic pressure, and is stored in advance in accordance with the type of shift and the type of frictional engagement element for release control. ing. Tt is an estimated value of the input shaft torque of the speed change mechanism, and is obtained from the intake air flow rate / rotation speed of the engine and the speed ratio of the torque converter as described above. Tr-o is a release critical torque ratio for obtaining a critical transmission torque capacity at which the release-side friction engagement element causes slippage with respect to the input shaft torque Tt. The margin allowance initial value is an initial value of a margin allowance which is a correction coefficient for adding a marginal transmission torque capacity to the critical transmission torque capacity, and is stored in advance, for example, as a value of about 3.0. Prt
"no" is the release-side standby pressure (release-side return spring pressure), which is stored in advance for each friction engagement element.

In step S322, the allowance is calculated. The allowance is calculated to decrease from the initial allowance (= 3.0) to the target value (the allowance (1)) after a lapse of a predetermined time TIMER1, and more specifically, the allowance corresponding to the elapsed time t. It is assumed that the allowance is obtained as a margin allowance = initial value × (1−gain α × t ^1/2 ).

Here, if the target value of the allowance (the allowance (1)) after the elapse of the predetermined time TIMER1 is 1.2, the predetermined time TIMER1 is substituted into the above-mentioned t, and if the allowance is substituted by 1.2, the gain becomes α is determined, and a margin for each elapsed time t is obtained by using the gain α. It should be noted that the target value of the allowance after the lapse of the predetermined time TIMER1 is set as a value by which the disengagement-side friction engagement element can maintain the engaged state even if the estimation error of the input shaft torque occurs within the expected range. .

In step S323, using the margin for each elapsed time t obtained as described above, the predetermined time TI
The release-side hydraulic pressure Po1 in the MER1 is calculated according to the following equation. Po1 = K1 × (Tt × Tr-o) × Margin + Prtn-o After the release-side hydraulic pressure is gradually reduced within the predetermined time TIMER1 as described above, the turbine rotation speed Nt [rpm] is determined in step S33. The process proceeds to step S34 until it is determined that the transition of the torque phase has been determined based on the comparison between the torque phase and the reference turbine rotation speed.

In step S34, a sharing ratio ramp control is performed. The details of the sharing ratio ramp control in step S34 are shown in the flowchart of FIG.
At 41, a predetermined time TIMER2 stored in advance according to the type of shift and the type of friction engagement element to be released is controlled.
Within the predetermined time TIME, the speed is reduced at a constant speed from the allowance (1) to the allowance (2) (for example, 0.8).
A margin in R2 is determined (see FIG. 9).

In step S342, the hydraulic pressure Po2 on the release side is calculated according to the following equation, using the allowance determined in step S341. Po2 = K1 × (Tt × Tr-o) × room allowance + Prtn-o On the other hand, the preparation phase process on the fastening side is shown in the flowchart of FIG. In step S41, it is determined whether or not a shift to the torque phase has been determined.

If the transition to the torque phase has not been determined, the process proceeds to step S42 as the preparation phase. In step S42, the precharge pressure (standby pressure) of the engagement-side friction engagement element is set according to the type of the friction engagement element. In step S43, a transient response compensation process is performed on the precharge pressure (standby pressure), and the result is output as the final engagement side hydraulic pressure Po0.

The transient response compensation processing means that the command oil pressure P is processed by an inverse filter (transient oil pressure compensation filter). The inverse filter reduces the damping rate of the hydraulic control system by ζre
al, the target value of the damping rate is ζtgt, the natural frequency of the hydraulic control system is ωreal, and the target value of the natural frequency is ωtgt,
Using the Laplace transform, the transfer function (transfer function) is calculated as (s <sup>2
+ 2ζrealωreals + ωreal 2) / (</sup> s 2 + 2ζtgtωtgt
s + ωtgt ² ), and the filter gain GAINatf is G
This is a filter in which AINatf = ω ² tgt / ω ² real. The damping rate ζreal and the natural frequency ωreal of the hydraulic control system are
The configuration is set according to the ATF temperature (oil temperature) at that time.

In general, the dynamic characteristic of the actual hydraulic pressure with respect to the command hydraulic pressure has a dead time and a secondary delay, and the secondary delay is approximated by a transfer function using the natural frequency and the damping rate as parameters. The hydraulic response will be degraded due to the resonance at. Therefore, in order to cancel the resonance point, an inverse filter is constructed by multiplying a model identified by the system (actual transfer characteristic) and a reference model (target transfer characteristic) not exhibiting resonance in a transient response. The hydraulic response is improved by processing the command value of the hydraulic pressure to control the solenoid valve.

When the ATF temperature (oil temperature) increases, the damping rate increases by ζreal and the natural frequency ωreal.
The damping rate is changed to ζreal and the natural frequency ωreal according to the TF temperature (oil temperature) so that an accurate inverse filter can be set. In the precharge, since air is mixed in the oil path, the natural frequency ωreal is lower than in the torque phase and the like, and the natural frequency ωreal changes depending on the elapsed time from the start of the precharge. For this reason, the damping rate ζreal and the natural frequency ωreal in the precharge are provided in separate maps corresponding to the ATF temperature (oil temperature) and the elapsed time t from the start of the precharge, which correlates with the change in the amount of air entrapment. By using the damping rate ζreal and the natural frequency ωrea retrieved from this map at the time of charging, a hydraulic response in precharging can be ensured.

In step S44, it is determined whether or not the elapsed time from the shift start determination has exceeded the predetermined time TIMER1, and if it has exceeded the predetermined time TIMER1, the flow proceeds to the sharing ratio ramp control in step S45. Details of the sharing ratio ramp control in step S45 are shown in the flowchart of FIG. 11, and in step S451, the predetermined time TIMER
Within 2, allowance (1) (for example, 0.8) to allowance (2)
(Eg 1.2) at a constant rate,
A margin within a predetermined time TIMER2 is determined (see FIG. 12).

In step S452, the hydraulic pressure Pc2 on the engagement side is calculated according to the following equation, using the allowance determined in step S451. Pc2 = K2 × (Tt × Tr−c) × Margin + Prtn−c Here, K2 is a coefficient for converting the transmission torque capacity (required transmission torque capacity) of the engagement side frictional engagement element into hydraulic pressure. Are stored in advance according to the type of shift and the type of friction engagement element to be released. Tr-c is a critical engagement torque ratio for obtaining a critical transmission torque capacity at which the engagement-side frictional engagement element starts engaging with respect to the input shaft torque Tt. Prtn-c is the standby pressure on the engagement side (return spring pressure on the engagement side) and is stored in advance for each friction engagement element.

Returning to the flow chart of FIG. 5, if the transition to the torque phase is determined in step S2 (turbine rotational speed Nt [rpm]> reference turbine rotational speed + hysteresis value HYS), it is determined. The process proceeds to step S4, where the gear ratio (gear ratio = turbine rotation speed N)
t / output shaft rotation speed No) exceeds a predetermined F / B (feedback) start gear ratio and changes in the upshift direction. Then, the torque phase process of step S5 is performed until the gear ratio exceeds the F / B start gear ratio and changes in the upshift direction.

In the torque phase processing of the release-side friction engagement element, the release-side hydraulic pressure Po2 obtained by continuing the control for reducing the margin in the preparation phase at the same speed.
Is added to the correction oil pressure Po3 for compensating for the shortage of the transmission torque capacity and suppressing the idling, and the final release oil pressure P
Obtain o4. Specifically, as shown in the flowchart of FIG. 13, first, the release correction hydraulic pressure Po3 is calculated according to the differential value ΔNt of the turbine rotation speed Nt. The release correction oil pressure Po3 is set to a larger value as the differential value ΔNt is larger.

In step S52, the release-side hydraulic pressure Po2 calculated based on the allowance set by continuing the reduction control of the allowance in the preparation phase at the same speed is set as follows.
The release correction hydraulic pressure Po3 is added, and the result is set as a final release-side hydraulic pressure Po4 (Po4 = Po2 + Po).
3). Note that the final release hydraulic pressure Po4 is equal to the release hydraulic pressure P4.
A limit is added so as not to fall below o2. That is, the release correction hydraulic pressure Po3 is equal to the release-side hydraulic pressure Po2.
Is used only for the correction of

On the other hand, the state of the torque phase processing of the engagement side frictional engagement element is shown in the flowchart of FIG. In the flowchart of FIG.
When it is determined that the shift to the torque phase has been determined, the process proceeds to step S62 to determine whether or not the gear ratio has exceeded the F / B start gear ratio and has changed in the upshift direction. And if it does not exceed the F / B start gear ratio,
Proceed to step S63.

In step S63, the engagement side oil pressure Pc2 (reference oil pressure) is obtained based on the margin set by continuing the increase control of the margin in the preparation phase at the same speed. In addition, in step S64, the air blow convergence control is executed. This blowing convergence control is shown in detail in the flowchart of FIG.

First, in step S641, the maximum value PHOSMAX of the hydraulic pressure after correction based on the input shaft torque Tt at that time.
Is retrieved from a table as shown in FIG.
Note that the maximum correction hydraulic pressure PHOSMAX is set to a larger value as the input shaft torque Tt is larger. In step S642, a correction oil pressure PHOS is calculated which increases in a stepwise manner to a value corresponding to the maximum correction oil pressure PHOSMAX and then gradually decreases to increase the reference command oil pressure in a so-called triangular manner.

More specifically, as shown in FIG. 17, it is determined that turbine rotation speed Nt [rpm]> reference turbine rotation speed + hysteresis value HYS, and when the occurrence of idling is detected, the engagement side hydraulic pressure is determined. Is increased stepwise from the command pressure (reference oil pressure) at that time to the maximum correction oil pressure PHOSMAX, the transfer function of the first-order lag system = G / (1 + Ts)
(G is a gain and Ts is a time constant)
Calculate the primary delay oil pressure that increases and changes to OSMAX, and calculates the deviation between the maximum correction oil pressure PHOSMAX and the primary delay oil pressure,
It is calculated as the corrected hydraulic pressure PHOS.

Here, as shown in FIG. 18, the time constant Ts is set in accordance with the turbine rotational speed Nt [rpm] -differential value of the reference turbine rotational speed (change speed of idling). Is larger, the time constant Ts is larger. Therefore, when the speed of change of the idling at which the turbine rotation speed rises rapidly is high, the first-order lag hydraulic pressure is calculated with a large time constant, so that the correction hydraulic pressure PHOS becomes larger and the time increase correction becomes longer. Is set to a value that performs

Step S65 in the flowchart of FIG.
Then, in the same manner as in step S51, the engagement correction hydraulic pressure Pc3 is calculated according to the differential value ΔNt of the turbine rotation speed Nt, and the final engagement-side hydraulic pressure Pc4 is set as Pc4 = Pc2 + PHOS + Pc3.

Although there is a response delay in the valve for controlling the hydraulic pressure, and there is a time delay until the actual pressure matches the command pressure, the actual pressure is increased stepwise by the correction hydraulic pressure PHOS. Hydraulic pressure can be started up with good response. Also, the maximum corrected hydraulic pressure corresponding to the input shaft torque Tt
With PHOSMAX as the initial value, the correction hydraulic pressure PHOS is gradually attenuated with a characteristic corresponding to the speed of change of the blowing, so that the correction amount necessary for suppressing the blowing can be added over the necessary period, Can be effectively suppressed.

In the present embodiment, the correction according to the differential value of the turbine rotational speed is performed in parallel with the correction by the correction hydraulic pressure PHOS. However, the correction by the correction hydraulic pressure PHOS can approximately converge the idling. The correction load is small, and even if there is a slight air blow after the end of the step correction, it can be converged with good response. However, the correction according to the differential value of the turbine rotation speed may be omitted.

In the above embodiment, the command pressure of the engagement-side friction engagement element is corrected by the correction hydraulic pressure PHOS. However, the command pressure of the release-side friction engagement element may be similarly corrected. Then, the correction hydraulic pressure PHOS may be divided into the engagement side and the release side, and the hydraulic pressure may be corrected on both the engagement side and the release side. Although the initial value of the corrected hydraulic pressure PHOS is set to a value equivalent to the maximum corrected hydraulic pressure PHOSMAX in the above description, the maximum corrected hydraulic pressure PHOSMAX is changed to a change speed of the idling (turbine rotation speed Nt).
The smaller the [rpm] -the differential value of the reference turbine rotation speed), the smaller the correction, and the maximum corrected hydraulic pressure PH after the correction.
The configuration may be such that the correction amount PHOS is calculated based on OSMAX, or the initial value of the correction hydraulic pressure PHOS is set only from the speed of change of the idling (turbine rotation speed Nt [rpm]-the differential value of the reference turbine rotation speed). The configuration may be such that a larger initial value is given as the speed of change of the blowing when the occurrence of the blowing is detected is higher. Further, the initial value of the correction hydraulic pressure PHOS may be set to a larger value as the engine rotation speed is lower.

In step S4 of the flowchart of FIG.
If it is determined that the gear ratio has exceeded the predetermined F / B start gear ratio, the process proceeds to step S6, where it is determined whether the gear ratio has exceeded a predetermined F / B end gear ratio (<F / B start gear ratio). Determine. When the gear ratio is between the F / B start gear ratio and the F / B end gear ratio, an inertia phase process of step S7 is performed.

For the disengagement side frictional engagement element, the oil pressure at the end of the torque phase process (normally, the oil pressure at the end = 0) is held during the inertia phase. On the other hand, the inertia phase processing of the engagement side frictional engagement element is performed as shown in FIG.
Is shown in the flowchart of FIG. In the flowchart of FIG. 19, in step S81, the basic control shown in the flowchart of FIG. 20 is performed.

In the basic control, first, in step S811, the target inertia torque Tinr [Nm] is calculated as follows.
It is calculated according to the following equation. Tinr = Inertia INS × Target turbine angular acceleration [rad
/ sec ² ] In the above equation, the inertia INS (moment of inertia) [Nm / rad
/ sec ² ] is a value determined according to the type of shift.

The target turbine angular acceleration [rad / sec ² ]
Is calculated as target turbine angular acceleration [rad / sec ² ] = 2 × π × target turbine acceleration [1 / sec ² ] / 60, and the target turbine acceleration [1 / sec ² ]
Is the target turbine acceleration [1 / sec ² ] = (Nt × gear step)
/ (Target shift time [sec]) In the above equation, the gear step is a value calculated as gear step = 1− (gear ratio after shift / gear ratio before shift), and Nt [rpm]
Is the turbine rotation speed at the start of the inertia phase.

In step S812, the engagement side hydraulic pressure Pc7 is calculated according to the following equation based on the target inertia torque Tinr. Pc7 = K2 × Tt × Tr × Tr-c + Prtn-c + K2 × Tr
-c × Tinr In addition to the above basic control, in step S82, rotation feedback (F / B) control is executed.

The rotation F / B control will be described with reference to the flowchart of FIG. In step S821, a target turbine rotation speed [rpm] is calculated. The target turbine rotation speed is determined by the turbine rotation speed Nt [rpm] at the start of the inertia phase and the target turbine acceleration [1 / se].
c ² ], the target turbine acceleration [1 / se] is calculated from the turbine rotation speed Nt [rpm] at the start of the inertia phase.
c ² ] (target turbine speed (n) = target turbine speed (n−1) + target turbine acceleration).

In step S822, the target turbine rotation speed [rpm] and the actual turbine rotation speed Nt [rpm] are determined.
A feedback correction is calculated by a PID (proportional / integral / differential) operation based on the deviation from the above. Step S82
3, the feedback correction amount is set to the engagement side hydraulic pressure P.
The result added to c7 is output as engagement-side hydraulic pressure Pc8.

If it is determined in step S6 of the flowchart of FIG. 5 that the gear ratio has become smaller than the F / B end gear ratio, the process proceeds from step S6 to step S8, where the gear ratio is changed to the F / B end gear ratio. Then, it is determined whether or not a predetermined time TIMER7 has elapsed from the first time the value has become smaller. If the time is within the predetermined time TIMER7, the process proceeds to step S9, and an end phase process is performed.

In the termination phase process for the disengagement-side friction engagement element, a process for continuously maintaining the command oil pressure (oil pressure = 0) is performed. On the other hand, the termination phase process of the engagement-side friction engagement element is shown in the flowchart of FIG. 22. In step S111, the predetermined time period from the hydraulic pressure corresponding to the critical coupling torque to the hydraulic pressure corresponding to 1.2 times the critical coupling torque is used. The ramp gradient Rmp-Tr2 to be raised in the TIMER 7 is set. The predetermined time TIMER7 is set according to the type of the speed change and the friction engagement element.

In step S112, the engagement side instruction pressure Pc
9 as follows: Pc9 = K2 × Tt × Tr-c × (1 + 0.2 × Rmp-Tr2) + P
It is calculated as rtn-c + K2 × Tr-c × Tinr. Then, when the predetermined time TIMER7 has elapsed, the command pressure on the fastening side is step-changed from Pc9 to the maximum pressure.

[Brief description of the drawings]

FIG. 1 is a diagram showing a transmission mechanism of an automatic transmission according to an embodiment.

FIG. 2 is a diagram showing a correlation between a combination of engagement states of frictional engagement elements in the transmission mechanism and a shift speed.

FIG. 3 is a system diagram showing a control system of the automatic transmission.

FIG. 4 is a time chart showing a state of shifting by changing the friction engagement element in the embodiment.

FIG. 5 is a flowchart showing a main routine of a shift change control of a friction engagement element in the embodiment.

FIG. 6 is a flowchart illustrating a preparation phase process of a disengagement-side friction engagement element.

FIG. 7 is a flowchart showing a disengagement initial hydraulic pressure calculation in a disengagement-side friction engagement element preparation phase process.

FIG. 8 is a flowchart illustrating a sharing ratio ramp control in a preparation phase process of a release-side friction engagement element.

FIG. 9 is a diagram showing a change in a margin in the sharing ratio ramp control.

FIG. 10 is a flowchart illustrating a preparation phase process of a fastening-side friction engagement element.

FIG. 11 is a flowchart showing a sharing ratio ramp control in a preparation phase process of the engagement-side friction engagement element.

FIG. 12 is a diagram showing a change in a margin in the sharing ratio ramp control of the engagement-side friction engagement element.

FIG. 13 is a flowchart showing a torque phase process of a disengagement-side friction engagement element.

FIG. 14 is a flowchart showing a torque phase process of the engagement-side friction engagement element.

FIG. 15 is a flowchart showing idling convergence control in a torque phase process of the engagement-side friction engagement element.

FIG. 16 is a diagram showing characteristics of a maximum corrected hydraulic pressure in the above-mentioned airflow convergence control.

FIG. 17 is a time chart showing a process of creating a correction hydraulic pressure in the above-mentioned airflow convergence control. .

FIG. 18 is a diagram showing a characteristic of a time constant in the airflow convergence control.

FIG. 19 is a flowchart showing an inertia phase process of the engagement-side friction engagement element.

FIG. 20 is a flowchart showing basic control in an inertia phase process of the engagement-side friction engagement element.

FIG. 21 is a flowchart showing rotation feedback control in an inertia phase process of the engagement-side friction engagement element.

FIG. 22 is a flowchart showing details of a termination phase process of the engagement-side friction engagement element.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 1 ... Torque converter 2 ... Transmission mechanism 11 ... Solenoid valve unit 12 ... A / T controller 13 ... A / T oil temperature sensor 14 ... Accelerator opening degree sensor 15 ... Vehicle speed sensor 16 ... Turbine rotation sensor 17 ... Engine rotation sensor 18 ... Air flow Meter 20: Engine G1, G2: Planetary gear H / C: High clutch R / C: Reverse clutch L / C: Low clutch 2 & 4 / B: 2nd / 4th speed band brake L & R / B: Low & reverse brake

Claims (7)

[Claims] A control device for an automatic transmission for a vehicle configured to perform a shift by switching an engagement state between a friction engagement element for a high-speed gear and a friction engagement element for a low-speed gear. Is detected in a stepwise manner up to a predetermined value at the time when it is detected, and thereafter, a correction value that decreases and changes at a speed corresponding to the changing speed of the idling is determined by the hydraulic pressure of the frictional engagement element on the release side and the frictional engagement element on the engagement side. A control device for an automatic transmission for a vehicle, wherein the control device is added to at least one of hydraulic pressure and hydraulic pressure. 2. A control device for an automatic transmission for a vehicle according to claim 1, wherein said predetermined value is determined according to an input shaft torque of a transmission mechanism. 3. The vehicle according to claim 2, wherein a value obtained by subtracting a reference oil pressure of the friction engagement element from a maximum oil pressure set according to an input shaft torque of the transmission mechanism is set as the predetermined value. Control device for automatic transmission. 4. The control device according to claim 1, wherein the predetermined value is determined in accordance with a change speed of the idling and / or an engine rotation speed. 5. The system according to claim 1, wherein said correction value is decreased from said predetermined value by a first-order lag, and a time constant of said first-order lag is set in accordance with a change speed of an air blow. The control device for an automatic transmission for a vehicle according to any one of the above-described items. 6. A differential value of a difference between a reference input shaft rotation speed calculated from a gear ratio before a gear shift and an output shaft rotation speed of a transmission mechanism and an input shaft rotation speed of the transmission mechanism, the differential value of the difference The control device for an automatic transmission for a vehicle according to any one of claims 1 to 5, wherein the calculation is performed as a parameter indicating a speed. 7. When the input shaft rotation speed of the speed change mechanism becomes higher than a reference input shaft rotation speed calculated from the gear ratio before the speed change and the output shaft rotation speed of the speed change mechanism by a predetermined value or more, the idling speed is reduced. The control device for an automatic transmission for a vehicle according to any one of claims 1 to 6, wherein occurrence is determined.