JPH10141486A - Control device for automatic transmission - Google Patents

Abstract

PROBLEM TO BE SOLVED: To obtain charging time easily by constructing an equation of motion of a hydraulic piston for driving a friction engaging element, and detecting only parameters for the temperature of an operating fluid and the like to compute. SOLUTION: A viscosity coefficient, a damping coefficient and a constant previously computed from oil temperature are substituted for each parameter of an equation 2, and charging time ts is computed from an expression 1. In the expression 1, Vp is the increase capacity of a piston chamber at the time of a clutch piston being displaced to the maximum from a reference position in a released state onto the engaging side, Qp is the unit time quantity of operating oil fed to a piston chamber. In the equation 2, Ap is the pressure receiving area of the clutch piston, L is the total choke length in a piston housing, (d) is a choke diameter, ρ is operating oil density, dc is an orifice diameter in the piston housing, Fset is the set load of a return spring, C is a flow coefficient, μis a viscosity coefficient, ΔXp is the maximum displacement quantity of the clutch piston, Cp is a damping coefficient, and P1 is control pressure. The equation is common even though an engine is different, and appropriate charging time can be computed.

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 a hydraulic control type automatic transmission.

[0002]

2. Description of the Related Art Conventionally, in a control device for a hydraulic control type automatic transmission, which switches a plurality of shift speeds by engaging or disengaging a friction engagement element according to an engine operating state,
When a shift is started according to an instruction from an automatic transmission control computer (hereinafter, “automatic transmission control computer” is referred to as an ECU), a frictional engagement element of a clutch or a brake that shifts from a disengaged state to an engaged state. Rapid filling control is performed so that the gap between the driving-side rotating body and the driven-side rotating body that has been formed is quickly and without any engagement impact. The optimum value of the time required for the quick filling control is determined by varying many parameters such as the temperature of the working fluid, the supply pressure of the working fluid, and the specifications of the friction engagement element, for example, the diameter of the clutch plate and the working stroke. It has been determined for each parameter by conducting experiments.

In the technique disclosed in Japanese Patent Laid-Open Publication No. Hei 5-3066761, the filling time for each parameter obtained by experiments is stored in a storage device of a controller, for example, as a map, and stored in a map of the storage device according to the engine operating state. The optimal filling time has been taken out. Then, an appropriate correction is applied to the extracted filling time, and the actuator of the friction engagement element is driven according to the corrected filling time.

[0004]

However, as described above, when the optimum value of the charging time is obtained from an experiment for each parameter as described above, it is necessary to determine the charging time by varying the parameters according to the operating state of the engine. It takes an enormous amount of time for experiments. Furthermore, since the map composed of the obtained filling times requires a large storage capacity, E
This is an obstacle in reducing the memory size of the CU.
Also, different engines require different maps, so it is necessary to experiment each time a new engine is developed. Therefore, it is difficult to reduce the number of development steps and development costs.

SUMMARY OF THE INVENTION It is an object of the present invention to provide a control device for an automatic transmission for easily obtaining a charging time of a frictional engagement element which shifts from a released state to an engaged state.

[0006]

SUMMARY OF THE INVENTION The present inventors have constructed an equation of motion of a hydraulic piston that drives a friction engagement element when the friction engagement element transitions from a released state to an engaged state. In the ECU, the parameters of the control device, which is a parameter of the charging time, are stored in advance, and only the parameters such as the temperature of the working fluid which have a great influence on the movement of the hydraulic piston are detected to calculate the equation of motion in the ECU. An attempt was made to calculate the required filling time by performing successive calculations.

The movement of the hydraulic piston is determined by inertial resistance, viscous resistance of the working fluid, and the balance between the urging force of the return spring of the hydraulic piston and an external force such as working hydraulic pressure. Among these, it has been clarified by examination that the influence of the inertial resistance of the piston and the fluid on the movement of the hydraulic piston is extremely small with respect to other forces, and is within a practically negligible range. That is, the movement of the hydraulic piston is determined by the hydraulic pressure and the urging force of the return spring.

Therefore, the filling time can be calculated by a steady-state operation in which the parameters are substituted into the calculation expression and the operation is simply performed without performing a transient operation for solving a motion equation as a differential equation. As a result of the trial calculation, it has been found that the filling time can be obtained with an error of about 5% from the transient operation by the steady operation. As described above, since the filling time can be calculated by a simple calculation by substituting the parameters into the calculation formula, there is an effect that the calculation time and the memory capacity can be reduced.

According to the control device for an automatic transmission according to the present invention, the drive-side rotating body of the friction engagement element is pressed by the hydraulic piston from the start of the transition from the release state to the engagement state. By calculating the filling time required until the gap with the driven-side rotating body becomes substantially zero by substituting parameters into the calculation formula, the filling time can be calculated in a short time without requiring a large storage capacity. Can be.

According to the control device for an automatic transmission according to the third aspect of the present invention, the charging time is calculated using the line pressure, which is the original pressure of the working fluid, as a parameter, so that the charging can be properly performed even when the line pressure fluctuates. Time can be calculated. According to the control device for an automatic transmission according to the fourth aspect of the present invention, the line pressure is calculated from the rotation speed of the hydraulic pump, and the filling time is calculated using the line pressure. It is especially effective.

When the frictional engagement element is shifted from the released state to the engaged state, when the rotation speed of the hydraulic pump is low, that is, when there is no margin in the fluid supply amount from the hydraulic pump, the hydraulic piston When the working fluid is supplied to the piston chamber in a short time, the line pressure, which is the original pressure of the working fluid, may decrease. If the filling time is calculated without considering such a decrease in the line pressure and the filling control is performed, a gap remains between the driving-side rotating body and the driven-side rotating body even after a predetermined filling time has elapsed. There is a problem that the accuracy of the combined control is reduced.

[0012]

Embodiments of the present invention will be described below with reference to the drawings. FIG. 1 shows a control device for an automatic transmission according to an embodiment of the present invention. Clutch plate 1 as driving-side rotator and partition plate 2 as driven-side rotator
Constitute a frictional engagement element, and when the clutch plate 1 and the partition plate 2 are engaged, torque is transmitted from the driving side to the driven side. The clutch plate 1 is urged by the return spring 3 in a direction away from the partition plate 2, that is, in a release side of the friction engagement element.

A clutch piston 4 as a hydraulic piston
Has a pressing portion 4a on the clutch plate side, and when the pressing portion 4a presses the clutch plate 1 toward the partition plate 2, the frictional engagement element is engaged. When the force received by the clutch piston 4 from the hydraulic oil in the piston chamber 5 exceeds the urging force of the return spring 3, the clutch piston 4 moves toward the clutch plate 1. An O-ring 6 for preventing leakage of hydraulic oil from the piston chamber 5 is fitted around the outer periphery of the clutch piston 4. The orifice 7 indicates a hydraulic oil leak from the piston chamber 5.

The choke 8 represents a portion corresponding to a choke formed in the piston housing 10 extending from the inflow port 10a to the piston chamber 5, and the orifice 9 represents a portion corresponding to an orifice formed in the piston housing 10. The choke 8 and the orifice 9 suppress the sudden flow of hydraulic oil into the piston chamber 5. The hydraulic oil supplied to the piston chamber 5 is discharged from a hydraulic pump 20 as a hydraulic pump. The hydraulic oil discharged from the hydraulic pump 20 is adjusted to a predetermined line pressure by a line pressure control valve 30 as a line pressure control means.

The line pressure control valve 30 has a command control chamber 31 and a line pressure feedback chamber 32. A command pressure is applied from the solenoid valve 42 to the command control chamber 31 via the choke 24, and the adjusted line pressure hydraulic oil is fed back to the line pressure feedback chamber 32 via the choke 23. The balance between the force received by the spool valve 33 from the command control chamber 31 and the line pressure feedback chamber 32 and the urging force of the return spring 34 moves the spool valve 33 to adjust the release amount of hydraulic oil to the drain 35, and the hydraulic pump The hydraulic oil discharged from the pressure controller 20 is adjusted to a line pressure corresponding to a command pressure from the solenoid valve 42.

Reference numeral 21 denotes a volume occupied by the hydraulic oil which is adjusted to the line pressure by the line pressure control valve 30, and a choke 22 indicates leakage of the hydraulic oil adjusted to the line pressure. The solenoid valve 41 as a hydraulic control valve is a three-way solenoid valve whose duty ratio is controlled.
According to the duty ratio of the control signal sent from the 50 control the hydraulic oil of the line pressure P _L as source pressure to a predetermined control pressure P ₁ supplied to the piston chamber 5. It is also possible to use a proportional valve as the solenoid valve 41.

The ECU 50 inputs detection signals such as an engine speed, a turbine speed as a driven-side rotator, a throttle opening, an oil temperature of a working fluid, and sends control signals to the respective solenoid valves. The oil temperature of the working fluid is detected by a temperature sensor as liquid temperature detecting means. Next, the operation of the frictional engagement element that shifts from the released state to the engaged state will be described with reference to FIGS. The steps described below are the steps in the flowchart of FIG. The flowchart of FIG. 2 is processed by a program stored in a memory serving as a storage unit in the ECU 50.

(1) The ECU 50 determines that shifting is necessary according to the operation of the select lever by the driver or the operating state of the engine (step 100). (2) According to the judgment of the ECU 50, the shift process is started (step 102). (3) The oil temperature of the operating oil is detected from the detection signal of the temperature sensor (step 103), and a constant described later is read from the memory (step 104).

(4) The viscosity coefficient μ and the damping coefficient C _P as variables previously calculated from the oil temperature and constants other than these variables are substituted into the respective parameters of the following equation 2 to obtain V _P.
And Q _P are calculated, and the filling time t _S is calculated from Equation 1 (step 105). V _P represents the increased capacity of the piston chamber 5 when the clutch piston 4 is displaced from the reference position in the disengaged state to the engagement side at the maximum, and Q _P is the piston chamber 5
Shows the unit time amount of hydraulic oil supplied to.

Steps 103 and 10 in the ECU 50
The program processes 4 and 105 correspond to “time calculating means” described in the claims.

[0021]

(Equation 1)

Here,

[0023]

(Equation 2)

Where A _P : Pressure receiving area of clutch piston 4 L: Total choke length in piston housing 10 d: Choke diameter in piston housing ρ: Hydraulic oil density d _C : Orifice diameter in piston housing 10 F _SET : set load of the return spring 3 C: flow coefficient mu: viscosity coefficient (= ρν) Δx _P: maximum displacement C _P of the clutch piston 4: damping coefficient (proportional function for [nu) P _1: control pressure oil temperature by increasing If, viscosity coefficient μ and the damping coefficient C _P decreases, the number the viscosity coefficient μ and the damping coefficient C _P is decreased 2
Since Q _P increases from, fill time t _S if increasing the oil temperature as shown in FIG. 3 is decreased.

(5) The duty ratio of the solenoid valve 41 is increased for the determined filling time t _S , and the hydraulic oil having a high control pressure P ₁ equivalent to the line pressure is applied to the piston chamber 5 as shown in the middle part of FIG. Supply (step 106). Then, the clutch piston 4 receives the force from the high-pressure hydraulic oil in the piston chamber 5 and quickly moves toward the clutch plate 1 from the position shown in FIG. The clutch piston 4 is further pressed toward the partition plate 2 and moves until the gap between the clutch plate 1 and the partition plate 2 becomes almost zero. The charging time t _S is from the time when the duty ratio of the solenoid valve 41 is increased and the oil pressure of the hydraulic oil in the piston chamber is increased to the time when the gap between the clutch plate 1 and the partition plate 2 becomes substantially zero. The processing performed during the filling time t _S is the filling control in step 106.

As shown in the middle part of FIG._SBut
Elapsed time t in FIG.₁, The solenoid valve 41
Supply from the solenoid valve 41 to the piston chamber 5
Control pressure P₁Sharply decrease. This allows the
The clutch plate 1 is moved to the partition plate 2 by the pressing force of the
Collision can be avoided. (6) By reducing the duty ratio of the solenoid valve 41
Suddenly reduced control pressure P ₁Gradually increase the clutch plate 1
And the partition plate 2 are slid. This is torque phase control
(Step 107). In the torque phase control, the clutch plate 1
Is pressed by the partition plate 2 and the clutch plate 1 and the partition plate 2 slide.
However, the torque of the clutch plate 1 is transmitted to the partition plate 2.
Not. Therefore, the rotation speed of the partition plate 2, that is,
Rotation speed N_tDoes not change. Torque phase control takes time t₁Or
T_TwoBetween.

(7) Control pressure P₁Further increase
Then, the torque of the clutch plate 1 is transmitted to the partition plate 2.
You. This is the time t shown in FIG._TwoAnd from here the partition plate
The number of revolutions of 2 starts to decrease. And the control pressure P₁Further
Time t_ThreeThen, the clutch plate 1 and the partition plate 2
Are engaged (step 108). This time t
_TwoTo t_ThreeIs inertia phase control.

(8) If the clutch plate 1 and the partition plate 2 shift to the engaged state, the shift control ends (step 10).
9). In the present embodiment, without taking into account the variation of the line pressure P _L during the filling control in step 106, to calculate the fill time t _S of the control pressure P ₁ is the output hydraulic pressure of the solenoid valve 41 as a constant. However, if the hydraulic oil is rapidly supplied to the low-pressure piston chamber 5 during the filling control in step 106,
Depending engine operating condition as shown in the lower part of FIG. 4 is that the line pressure P _L as the original pressure of the control pressure P ₁ temporarily decreases during the fill time t _S. Thus, the line pressure P
_{When L} falls below the command pressure of the solenoid valve 42, the solenoid valve 41
From the piston chamber 5 control pressure P ₁ of the hydraulic fluid supplied to the decreases, the clutch piston 4 to a predetermined position where the gap is substantially zero and the clutch plate 1 and the partition plate 2 may become impossible to move.

It has been found that such a decrease in the line pressure P _L is determined by the discharge amount of the hydraulic pump 20, the supply amount of hydraulic oil to the piston chamber 5, and the balance of oil leakage from each oil passage. because there can be obtained various values as constants, and the rotational speed of the hydraulic pump 20, i.e. the line pressure P _L to be lowered from the engine speed is predicted. Seek control pressure P ₁ is the output hydraulic pressure of the solenoid valve 41 from the obtained line pressure P _L, filled in consideration of reduction of the line pressure P _L in the control filling by substituting the control pressure P ₁ to the number 2 Time t _S can be determined. As shown in FIG. 5, the filling time t _S when the line pressure P _L decreases is longer.
Thus filling time taking into account the decrease in the line pressure P _L t _S
, The gap between the clutch plate 1 and the partition plate 2 can be reduced to almost zero when the filling control in step 106 is completed even if the line pressure P _L decreases during the filling control.

Specifications as constants and hydraulic pump 2
Line pressure P falling from 0 rpm _LTo predict and determine
Required line pressure P_LFrom the output hydraulic pressure of the solenoid valve 41
Control pressure P₁The process for obtaining is as shown by the dotted line in FIG.
Next, oil temperature detection (step 103), constant reading (step 103)
The process in parallel with step 104) is
This is especially effective when using a hydraulic pump
You. The rotation speed of the hydraulic pump 20, that is, the engine rotation speed
Is detected from a sensor as rotation speed detecting means. Also,
It is also possible to directly detect line pressure with a pressure sensor
is there.

On the other hand, if there is a margin in the discharge amount of the working oil even if the rotation speed decreases, the line pressure P _L
Without considering the reduction of the filling time t the control pressure P ₁ as a constant
_S can be calculated. In the present embodiment described above, many maps are stored in the memory by performing constant calculations by substituting the constants stored in the memory and the values of the variables obtained from the oil temperature into Equations 1 and 2 as parameters. The filling time t _S can be calculated easily and in a short time.

Further, calculated by using as a parameter of the fill time t _S calculated by detecting the oil temperature, even viscosity coefficient μ and the damping coefficient C _P is changed filling time with high precision t _S by a change in the oil temperature it can. Furthermore, to predict the reduction in the line pressure P _L from the number of revolutions of the hydraulic pump 20 as described above, high precision by obtaining the calculated fill time t _S of the control pressure P ₁ of the solenoid valve 41 from the line pressure Shift control can be realized. Even in the torque phase control and the inertia control, it is more precise to control the duty ratio of the solenoid valve 41 in consideration of the fluctuation of the line pressure due to the rotation speed of the hydraulic pump 20 and adjust the hydraulic pressure of the working oil supplied to the piston chamber 5 with higher accuracy. It is desirable from the viewpoint of realizing a proper shift control.

Equations (1) and (2) are commonly used even if the engine is different. Therefore, by storing constants in the memory, even when developing a different engine, the equation is uniquely determined from the structure. An appropriate charging time t _S can be calculated for each engine only by changing the constant on the memory.
Therefore, development time and development man-hours are reduced, and development costs can be reduced.

In the present invention, it is also possible to calculate a filling time t _S by substituting constants into Equations 1 and 2 in advance to calculate a coefficient and calculating a coefficient, and substituting variables into this calculation equation. It is. In this case, since it is not necessary to store the constant in the memory, the memory capacity can be reduced.

[Brief description of the drawings]

FIG. 1 is a schematic configuration diagram showing a control device for an automatic transmission according to one embodiment of the present invention.

FIG. 2 is a flowchart illustrating shift control of the present embodiment.

FIG. 3 is a characteristic diagram showing a relationship between oil temperature and filling time.

FIG. 4 is a characteristic diagram showing changes in turbine speed, control pressure, and line pressure from the start of gear shifting.

FIG. 5 is a characteristic diagram showing a relationship between a line pressure and a filling time.

[Explanation of symbols]

 Reference Signs List 1 clutch (friction engagement element) 2 partition plate (friction engagement element) 3 return spring 4 clutch piston (hydraulic piston) 5 piston chamber 20 hydraulic pump (hydraulic pump) 30 line pressure control valve (line pressure control means) 41 control valve (hydraulic pressure control valve) 50 ECU (electric control unit)

Claims (4)

[Claims] 1. A control device for an automatic transmission, wherein at least one friction engagement element is engaged or disengaged by a hydraulic pressure of a working fluid to switch a gear position, wherein a drive-side rotating body of the friction engagement element and A hydraulic piston that engages the driven-side rotator by pressing force, a hydraulic control valve that controls a hydraulic pressure that drives the hydraulic piston, and an electric control device that sends a control signal to the hydraulic control valve. When the friction engagement element is shifted from the release state to the engagement state, the drive hydraulic pressure to the hydraulic piston is increased by the hydraulic control valve, and then the drive side is pressed by the hydraulic piston. A control device for an automatic transmission, comprising: time calculating means for calculating a filling time until a gap between a rotating body and the driven-side rotating body becomes substantially zero by substituting parameters into a calculating formula. 2. A liquid temperature detecting means for detecting a liquid temperature of a working fluid, and a storage means for storing specifications of the hydraulic piston, wherein the time calculating means includes an output of the liquid temperature detecting means. The control device for an automatic transmission according to claim 1, wherein the charging time is calculated by using the specification value and the specification value. 3. A hydraulic pump for discharging a working fluid, and line pressure control means for controlling a line pressure of the working fluid discharged from the hydraulic pump, wherein the time calculating means includes a line pressure controller. 3. The control device for an automatic transmission according to claim 1, wherein the charging time is calculated using a value of the line pressure controlled by the means. 4. A rotating speed detecting means for detecting a rotating speed of the hydraulic pump, wherein the time calculating means calculates a line pressure value by using an output of the rotating speed detecting means. 4. The control device for an automatic transmission according to claim 3, wherein the charging time is calculated by using the following.