JP2001182819A - Hydraulic control device for automatic transmission - Google Patents

Abstract

PROBLEM TO BE SOLVED: To prevent a draw of the torque or racing in an automatic transmission making shifting with re-setting of a frictional engagement element by enhancing the controlling accuracy of oil pressure. SOLUTION: The indication pressures corresponding to the requisite torque capacities on the release side and engagement side are changed by a filter having such a characteristic as (s2+2ζrealωreals+ωreal2)/(s2+2ζtgtωtgts+ωtgt2)×(ω2tgt/ω2real), where ζreal is the damping factor, ζtgt is its target value, ωreal is the natural frequency, and ωtgt is the target value of natural frequency, while the damping factor ζreal and the natural frequency ωreal are changed in accordance with the oil temperature, and also in precharging, are changed according to the oil temperature and the elapsed time (t) from the start of precharging.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a hydraulic control device for an automatic transmission, and more particularly, to a technique for correcting a command oil pressure according to a transmission characteristic between a command oil pressure and an actual oil pressure.

[0002]

2. Description of the Related Art Conventionally, the engagement and release of a friction engagement element is controlled by hydraulic pressure, and the speed is changed by changing the friction engagement element which simultaneously performs the engagement control and the release control of two friction engagement elements. In such an automatic transmission, a required torque capacity is determined according to an input shaft torque of a transmission mechanism or the like, and an instruction value of a hydraulic pressure corresponding to the required torque capacity is output. (See Japanese Patent Application Laid-Open No. 9-133205).

[0003]

By the way, even if the command pressure of the hydraulic pressure corresponding to the required torque capacity is output, there is a delay between the command pressure and the actual hydraulic pressure, so that the response to the required torque capacity is improved. Control could not be performed, resulting in a shift shock or the like.

[0004] In particular, in an automatic transmission having a configuration in which the shift is performed by changing the friction engagement element that simultaneously performs the engagement control and the release control of the two friction engagement elements, the interlock ( This causes a drop in the output shaft torque), and also results in an idling due to a delay in the increase in the hydraulic pressure on the engagement side, and the response delay in the torque capacity control has been an important problem that affects shift performance.

SUMMARY OF THE INVENTION The present invention has been made in view of the above problems, and an object of the present invention is to provide a hydraulic control device for an automatic transmission which can suppress a delay in a transient response in hydraulic control and can prevent the occurrence of interlock and idling. Aim.

[0006]

SUMMARY OF THE INVENTION Therefore, the present invention is directed to a hydraulic control apparatus for an automatic transmission for controlling engagement / disengagement of a friction engagement element by hydraulic pressure. The command oil pressure is corrected by a transfer function that approximates the transfer characteristic between the oil pressure and the hydraulic pressure by a second-order lag system, and the friction engagement element is controlled by the corrected command oil pressure. The natural frequency and the damping rate are changed according to the oil temperature.

According to this configuration, the actual oil pressure has a secondary delay with respect to the command oil pressure, and the command oil pressure is corrected by a transfer function corresponding to the transfer characteristic of the secondary delay system.
The natural frequency and the damping rate of the hydraulic control system change according to the oil temperature (ATF temperature) of the automatic transmission, thereby changing the transfer characteristics. Change according to temperature.

According to a second aspect of the present invention, in a hydraulic control device for an automatic transmission for controlling engagement / disengagement of a friction engagement element by hydraulic pressure, a transmission between a command hydraulic pressure and an actual hydraulic pressure for the friction engagement element is provided. The command oil pressure is corrected by a transfer function that approximates the characteristics by a second-order lag system, the friction engagement element is controlled by the corrected command oil pressure, and the natural frequency and the damping rate constituting the transfer function Is changed according to whether or not the precharge is performed at the time of starting the engagement of the friction engagement element.

According to this configuration, the friction engagement element whose hydraulic pressure is to be controlled is an element to be engaged from the disengaged state, and when performing the precharge control at the time of starting the engagement, the friction engagement element other than the precharge is Set different natural frequencies and damping rates.

According to a third aspect of the present invention, in a hydraulic control device for an automatic transmission for controlling engagement / disengagement of a friction engagement element by hydraulic pressure, transmission between an instruction hydraulic pressure and an actual hydraulic pressure for the friction engagement element is provided. The command oil pressure is corrected with a transfer function that approximates the characteristics in a secondary system, the friction engagement element is controlled by the corrected command oil pressure, and the natural frequency and damping rate that constitute the transfer function are adjusted. It is configured to change according to whether or not it is a precharge at the time of starting the engagement of the friction engagement element and the oil temperature.

According to such a configuration, the natural frequency and the damping rate constituting the transfer function are changed according to the oil temperature (ATF temperature) of the automatic transmission, and the friction engagement element to be controlled is at the time of precharging. The natural frequency and the damping rate are changed depending on whether or not.

In the invention described in claim 4, in the precharge, the natural frequency and the damping rate are changed according to the elapsed time from the start of the precharge.
According to this configuration, when the friction engagement element to be controlled is at the time of precharging, the natural frequency and the damping rate are changed according to the time from the start of precharging.

According to a fifth aspect of the present invention, the transfer function is determined from an actual transmission characteristic between a command hydraulic pressure and an actual hydraulic pressure for the friction engagement element and a target transmission characteristic.

According to this configuration, the transfer function is determined from the transmission characteristics of the actual oil pressure when the command oil pressure is applied and the target transmission characteristics indicating how the actual oil pressure changes with respect to the command oil pressure. decide.

According to the sixth aspect of the present invention, the attenuation rate is set to ζrea
l, Zetatgt the target value of the damping factor, natural frequency Omegareal, the target value of the natural frequency when the Omegatgt, the transfer function (s <sup>2
+ 2ζrealωreals + ωreal 2) / (</sup> s 2 + 2ζtgtωtgt
s + ωtgt ² ), and the filter gain GAINatf is G
The command oil pressure is converted by a filter that sets AINatf = ω ² tgt / ω ² real.

According to this configuration, (s ² + 2ζrealωr
eals + ωreal ² ) / (s ² + 2ζtgtωtgts + ωtgt ² )
The indicated hydraulic pressure is converted by a filter set to a transfer function of × (ω ² tgt / ω ² real).

[0017]

According to the first aspect of the present invention, the transfer function can be set in accordance with the fluctuation of the natural frequency and the damping rate depending on the oil temperature, and there is an effect that the change in the response due to the oil temperature can be prevented.

According to the second aspect of the present invention, at the time of precharging, the natural frequency and the damping rate are affected by the air mixed in the hydraulic path, so that the actual natural frequency and the damping rate are changed. The command hydraulic pressure can be processed with the corresponding transfer function, and the control accuracy of the hydraulic pressure during precharge can be improved.

According to the third aspect of the present invention, the transfer function is set in accordance with the fluctuation of the natural frequency and the damping rate depending on the oil temperature and the fluctuation of the natural frequency and the damping rate depending on whether or not the precharge is performed. Thus, there is an effect that a change in responsiveness due to oil temperature and whether or not precharge is performed can be prevented.

According to the fourth aspect of the present invention, the amount of air mixed in the hydraulic path changes with the lapse of time from the start of the precharge, and accordingly, the transmission is performed in response to the fluctuation of the natural frequency and the damping rate. There is an effect that the function can be changed appropriately.

According to the fifth and sixth aspects of the present invention, the actual hydraulic pressure is changed by the target change characteristic with respect to the indicated oil pressure in accordance with the difference between the actual oil pressure change characteristic with respect to the indicated oil pressure and the target oil pressure change characteristic. There is an effect that it can be changed.

[0022]

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 composed of 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 has a reverse clutch R /
C is configured to be connectable to the input shaft IN, and is configured to be fixable by the brake bands 2 & 4 / B.

The sun gear S2 of the planetary gear set G2 is 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 r of the planetary gear set G2 is connected.
2 is configured to be connectable to the carrier c1 of the planetary gear set G1 by a low clutch L / C, and is further configured to be coupled to the carrier c of the planetary gear set G1 by a low & reverse brake L & R / B.
1 can be fixed.

The ring gear r1 of the planetary gear set G1 and the carrier c2 of the planetary gear set G2 are directly and integrally 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 a fastened state, and a part without a symbol indicates a released state.
The fastening state indicated by a black circle of & R / B indicates fastening only in one range.

According to the combination of the engagement states of the respective clutches and brakes shown in FIG. 2, for example, during a downshift from the fourth speed to the third speed, the brake bands 2 & 4 / B are released and the low clutch L / C is engaged. During the downshift from the third speed to the second speed, the high clutch H / C is released and the brake bands 2 & 4 / B are engaged. At the time of the upshift from the second speed to the third speed,
The brake bands 2 & 4 / B are released and the high clutch H / C is engaged. At the time of the upshift from the third speed to the fourth speed, the low clutch L / C is released and the brake bands 2 & 4 / B are engaged. That is, as described above, the shift in which the engagement and disengagement of the clutch / brake (friction engagement element) is simultaneously controlled to change the friction engagement element is referred to as a shift shift.

Each of the clutches and brakes (friction engagement elements) is operated by 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 for controlling various solenoid valves of the solenoid valve unit 11 includes an A / T oil temperature sensor 13, an accelerator opening sensor 14,
Detection signals from the vehicle speed sensor 15, the turbine rotation sensor 16, the engine rotation sensor 17, the air flow meter 18, and the like are input, 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 as 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.

Before giving a detailed description, an outline of the control will be described with reference to the time chart of FIG. 4 and the block diagram of FIG. First, the input shaft torque estimating unit 101 estimates the input shaft torque of the transmission mechanism, and the release-side F / F control unit 102
The engagement-side F / F control unit 103 calculates a feedforward F / F of the transmission torque capacity of the release-side friction engagement element and the engagement-side friction engagement element based on the input shaft torque. Here, while the transmission torque capacity of the disengagement side frictional engagement element is gradually reduced, the engagement side frictional engagement element can share the torque capacity shortage with the release side frictional engagement element alone. The transmission torque capacity of the engagement element is increased.

Further, when the occurrence of the idling, that is, the occurrence of the torque phase, is detected by the idling determining unit 104, the soft OWC control unit 105 corrects the idling to suppress the occurrence of the idling due to the insufficient torque capacity. The torque capacity is set, and this is added to the feedforward F / F of the release-side friction engagement element (and the engagement-side friction engagement element) as a feedback correction amount.

When the inertia phase judging section 106 detects that the inertia phase has been reached, the rotational feedback control section 107 sets a feedback correction amount for making the turbine rotational speed coincide with the target speed, and sets the feedback correction amount. Feedforward component F of engagement side frictional engagement element
/ F.

As described above, when the transmission torque capacity of each of the release-side friction engagement element and the engagement-side friction engagement element is determined, the torque-hydraulic conversion unit 108 converts the transmission torque capacity to hydraulic pressure. , And this oil pressure is
In step 09, dynamic characteristics compensation is performed, and the hydraulic pressure after the processing is converted into a control duty of a solenoid valve by a hydraulic-duty conversion unit 110, and the energization of each solenoid valve is controlled by the control duty.

Here, the details of the torque-hydraulic converter 108 and the inverse filter 109 will be described with reference to the control block diagram of FIG. The transmission torque capacity T of each of the disengagement side friction engagement element and the engagement side friction engagement element and the friction coefficient μ of the friction engagement element (clutch) are input to the torque-hydraulic conversion unit 108. .

The friction coefficient μ is set based on a clutch speed v which is set based on the type of shift and the turbine rotational speed Nt. The torque-to-hydraulic conversion unit 108 determines the transmission torque capacity T and the friction coefficient μ, the clutch area A,
From the return spring force Frtn, the number of clutches N, and the clutch diameter D, the command oil pressure P is calculated as P = 1 / A (Frtn + kT / NμD): (k is a constant).

On the other hand, an inverse filter (transient oil pressure compensation filter) 109 for processing the command oil pressure P sets the damping rate of the hydraulic control system to ζreal, sets the target value of the damping rate to ζtgt, and sets the natural frequency of the hydraulic control system to ωreal. When the target value of the natural frequency is ωtgt, the conversion function (transfer function) is calculated by using the Laplace transform as (s ² + 2ζreal ωreals + ωreal ² ) / (s ² +
2ζtgtωtgts + ωtgt ² ), and the filter gain GAI
This is a filter that sets Natf as GAINatf = ω ² tgt / ω ² real.

The damping rate ζreal and the natural frequency ωreal of the hydraulic control system are set according to the ATF temperature (oil temperature) at that time. In general, the dynamic characteristics of the actual oil pressure with respect to the command oil pressure have a dead time and a second-order lag.
The next delay is approximated by a transfer function using the natural frequency and the damping rate as parameters, and the resonance at the natural frequency deteriorates the hydraulic response. Therefore, a system-identified model (actual transfer characteristic) and a reference model that does not show resonance in the transient response (target transfer characteristic) in order to cancel the resonance point.
, The hydraulic response is improved by controlling the solenoid valve by processing the indicated value of the hydraulic pressure with the inverse filter.

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.

For the engagement-side friction engagement element whose hydraulic pressure is set to 0 before the gear shift, the hydraulic pressure is precharged at the start of the gear shift as described later. In the precharge, air is supplied to the oil path. Therefore, the natural frequency ωreal is lower than that during the torque phase, and the natural frequency ωreal changes depending on the elapsed time from the start of precharge. For this reason, the decay rate プ リ real
And the natural frequency ωreal are stored in a separate map according to the ATF temperature (oil temperature) and the elapsed time t from the start of the precharge correlating to the change in the amount of air entrapment, and the attenuation rate retrieved from this map during the precharge ζreal and natural frequency ωrea
Is used to ensure a hydraulic response in precharge.

Next, the details of the input shaft torque estimating section 101 will be described with reference to the block diagram of FIG. In the input shaft torque estimating unit 101, the engine rotation speed Ne [rp
m] and the intake air flow rate Qa [liter / h], a cylinder intake air amount Tp is obtained, and the cylinder intake air amount Tp is calculated.
And the engine rotation speed Ne, the engine generated torque [N
m].

On the other hand, the engine friction is estimated based on the temperature (hereinafter referred to as oil temperature) of the hydraulic oil (ATF) of the automatic transmission, and the engine generated torque is subtracted and corrected by the engine friction.

Further, an engine inertia torque is obtained from a change in the engine rotation speed Ne, and is added to the engine generated torque. Then, the engine generated torque is subjected to delay correction based on dynamic characteristics (primary delay and dead time) between the engine rotation speed Ne and the intake air flow rate Qa and the actual generated torque.

The transfer function in the delay correction is represented by e ^-T1s
/ (1 + T2s), and the dead time time constant T1 and the first-order lag time constant T2 are respectively set according to the engine speed Ne.

Further, the speed ratio of the torque converter is calculated from the engine speed Ne and the turbine speed Nt,
The torque ratio of the torque converter is obtained from the speed ratio. Then, the turbine torque is obtained by multiplying the engine generated torque subjected to the delay correction by the torque ratio, and further, the turbine torque is corrected by a shift inertia torque corresponding to a change in rotation during the shift to change the turbine torque. Input shaft torque.

The shift inertia torque is determined from the inertia (inertia moment) corresponding to the type of shift and the target acceleration obtained based on the target shift time, the gear ratio change, and the turbine speed at the start of the inertia phase. Is calculated.

Next, setting control of the transmission torque capacity of each of the release-side friction engagement element and the engagement-side friction engagement element, that is, the release FF control section 102, the engagement FF control 103,
Details of the software OWC control unit 105 and the rotation feedback control unit 107 will be described with reference to a time chart of FIG.
This will be described below.

In the following description, it is assumed that the conversion of the transmission torque capacity to the hydraulic pressure is simply performed using a constant. The flowchart of FIG. 8 shows a main routine of torque capacity control common to the engagement-side friction engagement element and the release-side friction engagement element.

In step S1, a shift determination for a power-on upshift is performed. The A / T controller 12 previously stores a shift map in which a shift stage is set according to a vehicle speed VSP and an accelerator opening (throttle opening). If the shift speed retrieved from the map is different, and it 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, where the gear ratio before shifting to the output shaft rotation speed No [rpm] of the transmission mechanism (gear ratio = turbine rotation Nt / output shaft rotation No). It is determined whether or not the input shaft rotation speed (turbine rotation speed) Nt [rpm] of the transmission mechanism is higher than an addition value of the reference turbine rotation obtained by multiplying the above and the hysteresis value HYS stored in advance.

If the turbine rotation speed Nt is equal to or less than the sum of the reference turbine rotation and the hysteresis value HYS, it is determined that the release-side friction engagement element has not been released, and the preparation phase process of step S3 is performed. Let it run.

The preparation phase processing in step S3 is as follows.
The process is divided into the process on the release side and the process on the fastening side. The flowchart of FIG. 9 shows the main routine of the preparation phase process of the disengagement side frictional engagement element. In step S31, the main routine is stored in advance according to the type of shift, the type of the frictional engagement element to be disengaged, and the oil temperature. Predetermined time TIMER1
It is determined whether or not only the shift determination has elapsed.

If the time is within the predetermined time TIMER1, the process proceeds to step S32 to calculate a release initial oil pressure. The release initial hydraulic pressure is an initial pressure for performing release control, and is a predetermined time TI from the hydraulic pressure during non-shift to the release initial hydraulic pressure.
It should be lowered in MER1.

The calculation of the initial disengagement hydraulic pressure in step S32 is shown in detail in the flowchart of FIG. 10. 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-shifting oil pressure Po0 is expressed as Po0 = K1 × (Tt × Tr-o) × initial allowance + 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 transmission mechanism. 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 torque capacity to the critical transmission torque capacity, and is stored in advance as a value of, for example, about 3.0. Prtn-o is the release-side standby pressure (release-side return spring pressure) and is stored in advance for each friction engagement element.

In step S322, the margin 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 for the above-mentioned t, and if the allowance is substituted for 1.2, the gain is obtained. α is determined, and a margin for each elapsed time t is obtained by using the gain α.

The target value of the margin after the lapse of the predetermined time TIMER1 is set to a value that enables the release-side friction engagement element to maintain the engaged state even if the estimated error of the input shaft torque occurs within the expected range. Is set.

In step S323, the allowance for each elapsed time t obtained as described above is used to determine 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 As described above, the hydraulic pressure on the release side is gradually decreased within the predetermined time TIMER1, and then, in step S33, the reference turbine rotation ( Until the turbine rotation speed Nt is determined to be higher than the sum of (No × gear ratio) and the hysteresis value HYS, the process proceeds to step S34 and subsequent steps.

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 341, a predetermined time TIMER stored in advance according to the type of shift and the type of friction engagement element to be released is controlled.
Within 2, allowance (1) to allowance (2) (for example, 0.8)
For a predetermined time TIM
A margin within ER2 is determined (see FIG. 12).

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) × Margin + Prtn-o The margin (2) (= 0.8) is released even if the estimation error of the input shaft torque occurs within the expected range. The value is set as a value that can surely shift the side friction engagement element to the release state.

In step S35, the sharing ratio ramp is limited. Details of the sharing ratio ramp limitation in step S35 are shown in the flowchart of FIG.
At 351, it is determined whether or not the input shaft torque Tt is equal to or less than a predetermined value.

If the input shaft torque Tt exceeds a predetermined value, the release hydraulic pressure Po calculated in step S34 is calculated.
In order to use 2 as it is, steps S352 to 354 are jumped and terminated, but if the input shaft torque Tt is equal to or smaller than a predetermined value, the process proceeds to step S352.

In step S352, the margin (2) is changed to a smaller value. For example, if the standard value is 0.8, change this to 0.6. With the above change, the change speed of the allowance (the hydraulic pressure Po2 on the release side) is further increased, thereby preventing the shift time from being prolonged at low torque.

In step S353, the allowance within the predetermined time TIMER2 is re-determined based on the changed allowance (2) in the same manner as in step S341. In step S354, the release hydraulic pressure Po2 is calculated based on the newly determined allowance.

In step S36, a sharing ratio ramp learning is performed. Details of the sharing ratio ramp learning in step S36 are shown in the flowchart of FIG.
At 361, it is determined whether or not the torque estimation learning for correcting the estimation error of the input shaft torque Tt has converged. The torque estimation learning will be described later.

If it is determined in step S361 that the torque estimation learning has converged, the flow advances to step S362 to increase the allowance (1) and the allowance (2) by 1.0.
, And the gradient of the margin within the predetermined time TIMER2 is reduced. For example, the allowance (1) is changed from 1.2 to 1.1, and the allowance (2) is changed from 0.8 to 0.9. By changing the margin, the rotation change at the beginning of the torque phase can be moderated, and the controllability in the torque phase can be improved.

In step S363, the allowance within the predetermined time TIMER2 is determined again in the same manner as in step S341 based on the changed allowances (1) and (2).
In step S364, the release hydraulic pressure Po2 is calculated based on the newly determined allowance.

It should be noted that, with the change in the allowance (1), the change in the allowance within the predetermined time TIMER1 is also changed. As described above, when the oil pressure on the release side is gradually decreased within the predetermined time TIMER2 in accordance with the setting for decreasing the allowance, the turbine rotation is more than the sum of the reference turbine rotation (No × gear ratio) and the hysteresis value HYS. By detecting the idling state of the engine with the high speed Nt, it is possible to indirectly know that the transmission torque capacity on the release side has dropped to near the criticality.

Here, it is ideal that the turbine rotation speed Nt becomes higher than the sum of the reference turbine rotation (No × gear ratio) and the hysteresis value HYS when the allowance is close to 1.0. If there is an estimation error of the input shaft torque Tt, the engine will run idle after the allowance margin is larger than 1.0 or becomes smaller than 1.0, and in consideration of the estimation error of the input shaft torque Tt, The change range of the allowance within the predetermined time TIMER2 is 1.0
(For example, 1.2 to 0.8).

For example, if the gear ratio starts to change at the release hydraulic pressure corresponding to the margin allowance = 1.1, the input shaft torque is estimated to be smaller than the actual value. If it is determined that the gear ratio has begun to slip even though the oil pressure can maintain the state, and conversely, for example, if the gear ratio starts to change at the release oil pressure corresponding to the allowance = 0.9, the input shaft torque is estimated from the actual value. Is estimated to be large, it is determined that the start of slippage has been delayed even though the hydraulic pressure (the transmission torque capacity) that cannot maintain the original engaged state has already been reduced.

Therefore, when the turbine rotation speed Nt becomes higher than the sum of the reference turbine rotation (No × gear ratio) and the hysteresis value HYS for the first time, step S3 is executed.
7 and perform torque estimation learning for obtaining a correction coefficient for correcting the input shaft torque estimation value based on the margin at that time. Details of the torque estimation learning in step S37 are described in FIG.
In step S371, an allowance at the time when the turbine rotation speed Nt first becomes higher than the sum of the reference turbine rotation (No × gear ratio) and the hysteresis value HYS is determined. It should be noted that since the detection of the idling is delayed, a predetermined time before the time when it is determined that the turbine rotation speed Nt first becomes higher than the sum of the reference turbine rotation (No × gear ratio) and the hysteresis value HYS. It is preferable that the allowance is set as the allowance at the time of occurrence of the idling.

In step S372, as shown in FIG. 16, a table in which the correction coefficient Ktt of the input shaft torque is stored in advance in accordance with the deviation (Tr-1) between 1.0 and the margin Tr when the engine is idling is generated. Step S
The correction coefficient Ktt is obtained by referring to the table based on the margin Tr obtained at 371.

The correction coefficient Ktt is such that the margin Tr is 1.
When the margin Tr is smaller than 1.0, the value is set to 1.0. When the margin Tr is larger than 1.0, the value is set to a value larger than 1.0.
so that when r is 1.0, the engine blows,
Correct the estimated value of the input shaft torque.

When the correction coefficient Ktt is set, learning is performed so as to estimate the input shaft torque including a correction request based on the correction coefficient Ktt. Further, the correction coefficient Ktt is limited within a predetermined upper and lower limit value,
The learning of the correction coefficient Ktt is performed when the ATF temperature is equal to or higher than a predetermined temperature.

On the other hand, the preparation phase process on the fastening side is performed as shown in FIG.
7 is shown in the flowchart. In step S41,
Reference turbine rotation (No x gear ratio) and hysteresis value H
It is determined whether or not the turbine rotation speed Nt is higher than the value added to YS.

When it is determined that the turbine rotation speed Nt is equal to or less than the sum of the reference turbine rotation (No.times.gear ratio) and the hysteresis value HYS, in other words, engine idling occurs. Until the time, the process proceeds to step S42.

In step S42, the reference 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, the damping rate ζreal and the natural frequency ωreal used in the inverse filter (transient oil pressure compensation filter) 109 are retrieved from the precharge map according to the ATF temperature and the elapsed time t from the start of precharge. To do. Then, the reference precharge pressure (standby pressure) is processed by an inverse filter (transient oil pressure compensation filter) 109 based on the damping rate ζreal and the natural frequency ωreal obtained from the precharge map, and the result is finally obtained. Is output as the appropriate engagement side hydraulic pressure Po0.

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.

The details of the sharing ratio ramp control in step S45 are shown in the flowchart of FIG.
At 451, within a predetermined time TIMER2, a margin (1)
Assuming that the margin is increased at a constant speed from (for example, 0.8) to the margin (2) (for example, 1.2), the margin within a predetermined time TIMER2 is determined (see FIG. 19).

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 a standby pressure on the engagement side (return spring pressure on the engagement side) and is stored in advance for each friction engagement element.

Here, returning to the flow chart of FIG. 8, the description will be continued. In step S2, the turbine rotation speed Nt is higher than the sum of the reference turbine rotation (No × gear ratio) and the hysteresis value HYS. Is determined, it is determined whether the gear ratio has exceeded the F / B (feedback) start gear ratio and has changed in the upshift direction. Then, the torque phase process of step S5 is performed from the determination of the engine idling until the change in the upshift direction beyond the F / B start gear ratio.

In the torque phase processing (soft OWC control) of the release-side friction engagement element, the release-side hydraulic pressure Po2 determined by continuing the control for reducing the margin in the preparation phase at the same speed is determined according to the idling. Correction oil pressure Po
By adding 3, the final release-side hydraulic pressure Po4 is obtained.

More specifically, as shown in the flowchart of FIG. 20, first, in step S51, the release correction hydraulic pressure P corresponding to the differential value ΔNt of the actual turbine rotation speed Nt is set.
o3 is calculated according to the following equation.

Po3 = K1 × ｛INS × (2π / 60)
× ΔNt + 1 / g (Nt−No × i)｝ where INS is inertia (moment of inertia) determined for each type of shift, g is a gain for converting clutch torque into rotational speed, and i is a gear ratio before the shift. is there.

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 to:
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).

It should be noted that a limit is imposed so that the final release hydraulic pressure Po4 does not fall below the release hydraulic pressure Po2. Further, the differential value ΔNt of the turbine rotation speed
The value after the low-pass filter processing is used.

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 turbine rotation speed Nt is higher than the sum of the reference turbine rotation (No × gear ratio) and the hysteresis value HYS, the process proceeds to step S62, and the gear ratio is changed to the F / B start gear ratio. Is determined in the upshift direction beyond the above. If it does not exceed the F / B start gear ratio, the process proceeds to step S63.

In step S63, the engagement side hydraulic pressure Pc2 is determined based on the margin set by continuing the increase control of the margin in the preparation phase at the same speed.
In step S64, the engagement correction hydraulic pressure Pc3 is calculated according to the following equation, similarly to step S51.

Pc3 = K2 × ｛INS × (2π / 60)
× ΔNt + 1 / g (Nt−No × i)} Then, Pc2 + Pc3 = Pc4 to obtain the final engagement side hydraulic pressure Pc4.

The control up to the torque phase of the release-side friction engagement element and the engagement-side friction engagement element will be schematically described with reference to the block diagram of FIG. Note that the block diagram in FIG. 22 shows the release-side F / F control unit 102, the engagement-side F / F control unit 103, the idling determination unit 104, and the software OW in FIG.
The detailed configuration of the C control unit 105 will be described.

The hydraulic pressures on the release side and the engagement side are basically determined by adding a margin to the critical torque obtained from the input shaft torque and the critical torque ratio according to the type of shift. It should be noted that the hydraulic pressure on the engagement side is precharged at the start of shifting.

Then, while the hydraulic pressure (torque capacity) on the engagement side is increased and changed as the margin increases, the hydraulic pressure (torque capacity) on the release side is reduced by decreasing the margin, thereby sharing the required torque capacity. Gradually changes from the release side to the engagement side. In addition, a correction according to the magnitude of the differential value of the turbine rotational speed described as the soft OWC control in the present embodiment is dealt with against the idling due to the insufficient torque capacity.

In the control block diagram of FIG.
t indicates a turbine rotational angular velocity, and ω (dot) t is a differential value of the turbine rotational angular velocity ωt. If it is determined in step S4 of the flowchart of FIG. 8 that the gear ratio has exceeded the F / B start gear ratio, the process proceeds to step S6, where the gear ratio changes to the F / B end gear ratio (<F / B start gear ratio). It is determined whether or not it has exceeded.

When the gear ratio is between the F / B start gear ratio and the F / B end gear ratio, an inertia phase process in step S7 is performed. The release-side inertia phase process is shown in the flowchart of FIG. 23. In step S71, a setting is made to hold the oil pressure at the end of the torque phase (oil pressure = 0).

The inertia phase process on the fastening side is shown in the flowchart of FIG. In the flowchart of FIG. 24, in step S81, the basic control shown in the flowchart of FIG. 25 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 based on the target inertia torque Tinr according to the following equation. 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 (inertia phase determination unit 106 and rotation feedback control unit 107)
Will be described with reference to the control block diagram of FIG. 26 and the flowchart of FIG.

In step S821, a target turbine rotation speed [rpm] is calculated. The target turbine rotation speed is the turbine rotation speed Nt at the start of the inertia phase.
[Rpm] and the target turbine acceleration [1 / sec ² ] based on the turbine rotation speed Nt at the start of the inertia phase.
It is calculated as a characteristic that decreases from [rpm] at the target turbine acceleration [1 / sec ² ] (target turbine speed (n) = target turbine speed (n-1) + target turbine acceleration).

In step S822, a feedback correction torque _TFB is calculated by feedback control based on the target turbine rotation speed [rpm]. Specifically, the feedback correction torque T _FB is calculated with a configuration as shown in the block diagram of FIG.

In consideration of the dynamic characteristic (first-order lag) between the clutch torque and the turbine rotation speed Nt, a reference model target Tgt # Nt # indicating a first-order lag with respect to the target turbine rotation speed Tgt # Nt [rpm]. Set kihan.

Here, the transfer function of the first-order lag is 1 /
(1 + T _tgt s), and a reference model target Tgt # Nt_kihan is obtained by processing the target turbine rotation speed Tgt # Nt [rpm] with the filter of the transfer characteristic, and the reference model time constant T _tgt Is set to, for example, 0.8 [sec].

The deviation between the reference model target Tgt # Nt_kihan and the actual turbine rotational speed Nt is obtained, and the feedback correction rotational speed FB # Nt is calculated by proportional / integral control based on the deviation.

On the other hand, based on the dynamic characteristics (first-order lag) between the actual clutch torque and the turbine rotation speed Nt and the reference model, an inverse filter having a transfer function (1 + T _real s) / (1 + T _tgt s) is formed. , The target turbine rotation speed Tgt #
Nt is converted to an inverse filter output Tgt # Nt # Inverse. Note that the time constant T _real indicating the actual dynamic characteristic may be a fixed value or may be set to a different value according to the turbine rotation speed Nt.

Then, a value obtained by subtracting the target turbine rotation speed Tgt # Nt from the inverse filter output Tgt # Nt # Inverse is set as a feedforward correction rotation speed FF # Nt. Here, the feedback correction rotation speed FB # Nt and the feedforward correction rotation speed FF # Nt are added, and the addition result is converted into a torque with a gain g corresponding to the turbine rotation speed Nt, thereby obtaining a feedback correction torque. _Find TFB.

[0111] At step S823, by multiplying the conversion factor K2 in the feedback correction torque T _FB into a feedback correction oil pressure and outputs the result obtained by adding the engagement-side oil pressure Pc7, as fastening-side hydraulic Pc8.

If it is determined in step S6 of the flowchart of FIG. 8 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 flow advances to step S9 to perform an end phase process.
The end phase process for the disengagement side frictional engagement element is shown in the flowchart of FIG. 28, and in step S91, the setting for maintaining the oil pressure at the end of the inertia phase is performed. That is, the oil pressure of the release-side friction engagement element is maintained at the value at the time when the gear ratio becomes smaller than the F / B start gear ratio in the inertia phase and the end phase.

On the other hand, the termination phase process of the engagement side frictional engagement element is shown in the flowchart of FIG.
At 101, it is determined whether or not within a predetermined time TIMER7 from the time when the gear ratio first becomes smaller than the F / B end gear ratio. Execute.

Details of the end phase process in step S101 are shown in the flowchart of FIG. 30. In step S111, the predetermined time from the hydraulic pressure corresponding to the critical engagement torque to the hydraulic pressure corresponding to 1.2 times the critical engagement torque is determined. 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 Pc9 = K2 × Tt × Tr-c × (1 + 0.2 × Rmp-Tr2) + P
It is calculated as rtn-c + K2 × Tr-c × Tinr.

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 control block diagram showing the entire control system of the automatic transmission.

FIG. 6 is a control block diagram showing a block for determining a command oil pressure from a required torque capacity.

FIG. 7 is a control block diagram showing a block for estimating an input shaft torque.

FIG. 8 is a flowchart showing a main routine of a shift change control of a friction engagement element in the embodiment.

FIG. 9 is a flowchart showing a preparation phase process of a release-side friction engagement element.

FIG. 10 is a flowchart showing a disengagement initial hydraulic pressure calculation in a disengagement-side friction engagement element preparation phase process.

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

FIG. 12 is a diagram showing a change in a margin in the sharing ratio ramp control.

FIG. 13 is a flowchart showing a sharing ratio ramp limit in a preparation phase process of a release-side friction engagement element.

FIG. 14 is a flowchart illustrating a sharing ratio ramp learning in a preparation phase process of a release-side friction engagement element.

FIG. 15 is a flowchart showing torque estimation learning in a preparation phase process of a disengagement-side friction engagement element.

FIG. 16 is a diagram showing characteristics of a correction coefficient of an input shaft torque in the torque estimation learning.

FIG. 17 is a flowchart showing a preparation phase process of the engagement-side friction engagement element.

FIG. 18 is a flowchart showing a sharing ratio ramp control in a preparation phase process of the engagement-side friction engagement element.

FIG. 19 is a diagram showing a change in a margin in the sharing ratio ramp control of the engagement-side friction engagement element.

FIG. 20 is a flowchart showing a torque phase process of a disengagement-side friction engagement element.

FIG. 21 is a flowchart showing a torque phase process of the engagement side frictional engagement element.

FIG. 22 is a control block diagram showing a block for performing setting of a required torque capacity for a feedforward direction and idling control;

FIG. 23 is a flowchart showing an inertia phase process of a disengagement-side friction engagement element.

FIG. 24 is a flowchart showing an inertia phase process of the engagement-side friction engagement element.

FIG. 25 is a flowchart showing basic control in an inertia phase process of the engagement-side friction engagement element.

FIG. 26 is a control block diagram illustrating a block for performing rotation feedback control in an inertia phase process of the engagement-side friction engagement element.

FIG. 27 is a flowchart showing rotation feedback control in an inertia phase process of the engagement side frictional engagement element.

FIG. 28 is a flowchart showing a termination phase process of a disengagement-side friction engagement element.

FIG. 29 is a flowchart showing an end phase process of the engagement side frictional engagement element.

FIG. 30 is a flowchart showing details of an end 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 and reverse brake

Claims (6)

[Claims] 1. A hydraulic control device for an automatic transmission for controlling engagement and disengagement of a friction engagement element by hydraulic pressure, wherein a transmission characteristic between an instruction hydraulic pressure and an actual hydraulic pressure for the friction engagement element is changed by a secondary delay system. The command oil pressure is corrected with a transfer function that is approximated by, and the friction engagement element is controlled by the corrected command oil pressure, and the natural frequency and damping rate that constitute the transfer function are adjusted according to the oil temperature. A hydraulic control device for an automatic transmission, characterized in that it is configured to change. 2. A hydraulic control device for an automatic transmission for controlling engagement / disengagement of a frictional engagement element by hydraulic pressure, wherein a transmission characteristic between a command hydraulic pressure and an actual hydraulic pressure for the frictional engagement element is changed by a secondary delay system. The command oil pressure is corrected by a transfer function that is approximated by, and the friction engagement element is controlled by the corrected command oil pressure, and the natural frequency and damping rate that constitute the transfer function are adjusted by the friction engagement. A hydraulic control device for an automatic transmission, wherein the hydraulic control device is configured to change according to whether or not a precharge is performed at the time of starting engagement of elements. 3. A hydraulic control device for an automatic transmission for controlling engagement / disengagement of a friction engagement element by hydraulic pressure, wherein a transmission characteristic between an instruction hydraulic pressure and an actual hydraulic pressure for the friction engagement element is determined by a secondary system. The command oil pressure is corrected by an approximate transfer function, the friction engagement element is controlled by the corrected command oil pressure, and the natural frequency and the damping rate constituting the transfer function are adjusted by the friction engagement element. A hydraulic control device for an automatic transmission, wherein the hydraulic control device is configured to change according to whether or not a precharge is made at the start of engagement of the vehicle and the oil temperature. 4. The hydraulic control device for an automatic transmission according to claim 2, wherein the natural frequency and the damping rate are changed according to an elapsed time from the start of the precharge. . 5. The transfer function according to claim 1, wherein a transfer function is determined from an actual transfer characteristic between a command hydraulic pressure and an actual hydraulic pressure for the friction engagement element and a target transfer characteristic. A hydraulic control device for an automatic transmission according to any one of the preceding claims. 6. An attenuation rate is ζreal, and a target value of the attenuation rate is ζtg.
t, the natural frequency is ωreal, and the target value of the natural frequency is ωtgt, the transfer function is (s ² + 2ζrealωreals + ωr
eal ² ) / (s ² + 2ζtgtωtgts + ωtgt ² ), and the filter gain GAINatf is defined as GAINatf = ω ² tgt / ω ² r
The hydraulic pressure control device for an automatic transmission according to claim 5, wherein the command hydraulic pressure is converted by a filter that is set to eal.