JPH10141484A - Controller for automatic transmission - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a controller for automatic transmission which controls shift with high precision even when a value of line pressure fluctuates. SOLUTION: When a clutch plate 1 and a bulkhead plate 2 are transferred from a release condition to an engagement condition, line pressure as original pressure may be reduced by filling hydraulic oil into a piston chamber 5 rapidly when a hydraulic pump 20 rotates at low speed. By detecting the number of revolutions of the hydraulic pump 20 and calculating line pressure which is reduced based on the number of revolutions thereof and a flow rate of consumption hydraulic oil which is filled in the piston chamber 5, it is possible to control shift with high precision in accordance with line pressure which is reduced.

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,
Various vehicle conditions, such as vehicle speed, engine speed,
Automatic transmission control computer (hereinafter referred to as
The shift control is performed in accordance with an instruction from the "automatic transmission control computer" (ECU). In such shift control, the control pressure of the friction engagement element is set from a line pressure which is an original pressure, and the line pressure is set by a line pressure control valve according to a command pressure from the ECU. Also, EC
The optimal control value of each control signal sent from U to the solenoid valve and the like is found from the results obtained by experiments according to the above-described conditions such as the vehicle speed, the engine speed, and the liquid temperature, and is stored in the memory in the ECU. It is stored as a matching map. The control value read from the matching map is used while being complemented.

[0003]

However, when the drive hydraulic pressure of the hydraulic piston for engaging the friction engagement element is increased by the pressing force to shift the friction engagement element from the released state to the engaged state, In the rapid filling control for rapidly filling the gap between the drive-side rotator and the driven-side rotator of the combined element, the line pressure may be lower than the command pressure from the ECU.
When the line pressure is lower than the command pressure, the control pressure set using the line pressure as the original pressure also decreases, so that there is a possibility that an engagement failure such as an increase in the engagement time or an insufficient engagement force may occur.

Further, as described above, in the method of obtaining an optimum control value from an experiment for each condition, it is necessary to oscillate parameters according to the operating state of the engine to obtain the control value, so that an enormous amount of time is required for the experiment. It costs. Further, the map formed from the obtained control values requires a large storage capacity, which is an obstacle to reducing the memory size of the ECU. 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.

An object of the present invention is to provide a control device for an automatic transmission that performs high-accuracy shift control even when the value of the line pressure fluctuates. It is another 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 that shifts from a released state to an engaged state.

[0006]

As a result of various experiments conducted by the present inventors, the decrease in the line pressure in the rapid filling control is not caused by the responsiveness of the line pressure control valve but by the liquid as a supply source of the working fluid. It has been found that it is determined by the balance between the discharge amount of the pressure pump and the consumption flow rate of the working fluid supplied to the piston chamber. Immediately after the working fluid was suddenly supplied to the piston chamber in the filling control, the line pressure control valve was closed and operated to increase the line pressure, but even in this state, the line pressure was reduced and did not increase. . The decrease in the line pressure increased to the command pressure when the discharge amount of the working fluid increased by increasing the rotation speed of the hydraulic pump. In other words, it has been found that the decrease in the line pressure during the filling control is caused by an insufficient discharge amount of the hydraulic pump at the time of low rotation.

The present inventors have also constructed an equation of motion of a hydraulic piston that drives the friction engagement element when the friction engagement element shifts from the released state to the engaged state, and stores the equation in advance in the ECU. Attempts to calculate the required filling time by detecting the specifications of the control device, which is the parameter of the filled time, the fluid temperature of the working fluid, and the engine speed, and sequentially calculating the equation of motion in the ECU. Was.

The movement of the hydraulic piston is determined by the balance between inertial resistance, viscous resistance of the working fluid, the urging force of the return spring of the hydraulic piston, and an external force such as working hydraulic pressure. Among them, the influence of the inertial resistance of the piston and the working fluid on the movement of the hydraulic piston is extremely small with respect to other forces,
As a result of examination, it has become clear that the range is practically negligible. 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 only the parameters are substituted into the calculation expression and the operation is performed without performing a transient operation for solving a motion equation as a differential equation.
As a result of the trial calculation, it was found that the filling time can be calculated with an error of about 5% with respect to the transient calculation by the steady calculation. 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 of the present invention, the line pressure is calculated from the rotation speed of the hydraulic pump and the consumption flow rate of the working fluid. Therefore, for example, even if the line pressure is reduced due to a shortage of the discharge amount of the hydraulic pump at the time of low rotation, the reduced line pressure can be calculated. Hydraulic pressure control is possible.

According to the control device for an automatic transmission according to the second aspect of the present invention, when the friction engagement element is shifted from the released state to the engaged state, the drive hydraulic pressure to the hydraulic piston is controlled by the hydraulic control valve. Is rising. Therefore, if the drive hydraulic pressure is increased to a value equivalent to the line pressure by the hydraulic pressure control valve, the friction engagement element can be quickly shifted from the released state to the engaged state. That is, quick filling control can be performed.

According to the control device for an automatic transmission according to the third aspect of the present invention, the hydraulic pressure piston is pressed by the hydraulic piston from the start of the transition from the released state to the engaged state, and the driven rotating body and the driven side of the friction engaging element are driven. By calculating the filling time required until the gap with the 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.

[0012]

Embodiments of the present invention will be described below with reference to the drawings. FIG. 1 shows a control device for a hydraulically controlled automatic transmission according to an embodiment of the present invention.
The clutch plate 1 as the driving-side rotator and the partition plate 2 as the driven-side rotator constitute a frictional engagement element, and the clutch plate 1 and the partition plate 2 are engaged to move from the driving side to the driven side. Is transmitted to the motor. 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 from the volume portion 21 of the hydraulic oil adjusted to the line pressure. . Solenoid valve 41 as a fluid pressure control valve is a three-way solenoid valve controlled duty ratio, the hydraulic oil of the line pressure P _L as source pressure according to the duty ratio of the control signal sent from the ECU50 as an electric control unit Is adjusted to a predetermined control pressure P ₁ and 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 hydraulic oil, and sends control signals to the respective solenoid valves. The oil temperature of the hydraulic oil 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) Detect the hydraulic oil temperature from the detection signal of the temperature sensor,
With the detected oil temperature to calculate the viscosity coefficient μ and the damping coefficient C _P, the engine rotational speed from the sensor as a rotation speed detecting means, that is for detecting the speed N _e of the hydraulic piston 20 (step 103). Further, a constant described later is read from the memory (step 104).

(4) Step 103 and Step 104
The constant value obtained in the above, the detected value as a variable, and the variable value calculated from the detected value are substituted into the following Expressions 1 to 5. In Equation 2, P ₁ = D _r P _L (D _r is the pressure ratio of the hydraulic pressure control valve 41). Therefore, Equations 2 and 4 form a simultaneous equation of Q _P and P _L. Calculating a consumed flow rate Q _P and the line pressure P _L by solving the simultaneous equations (step 105). The consumption flow rate Q _P represents a unit time amount of the working oil filled in the piston chamber 5. In this embodiment, the consumption flow rate Q _P is obtained by substituting the line pressure P _L shown in Equation 5 obtained by solving the simultaneous equations of Equations 2 and 4 into Equation 2, which is a solution method. it, and the rotational speed of the hydraulic piston 20 N _e and consumption rate Q _P
The line pressure P _L is calculated from the above equation. Further, the filling time t _S is calculated by substituting Q _P into Equation 1 (Step 106).

In step 105, equations 2 and 4
The program processing for solving the simultaneous equations is "consumption flow rate calculation means" and "line pressure calculation means". Solving the simultaneous equations also means to calculate a value by substituting parameters into a calculation formula.

[0021]

(Equation 1)

Here,

[0023]

(Equation 2)

The discharge oil amount Q of the hydraulic pump 20_OUTIs the consumption
Q is the flow rate_P, The amount of oil leaking from the flow path Q _aIs given by
expressed.

[0025]

(Equation 3)

Equation 3 can be summarized into the following equation 4.

[0027]

(Equation 4)

In equations 1 to 4, V_P : Increase of the piston chamber 5 at the time of transition to the engaged state
Capacity Q_P ： Consumption flow rate C_V : Internal leak coefficient of the hydraulic pump 20 d_a : Diameter L of choke 22_a : Length of chalk 22 q_th : Theoretical capacity of the hydraulic pump 20  A_P : Pressure receiving area of clutch piston 4 L: Length of choke 8 d: Diameter of choke 8 ρ: Hydraulic oil density d_C : Diameter of orifice 9 F_SET: Set load of return spring 3 C: Flow coefficient μ: Viscosity coefficient (= ρν, ν is kinematic viscosity coefficient) Δx_P: Maximum displacement of clutch piston 4 C_P : Damping coefficient (proportional function to ν) N_e : Engine speed (speed of hydraulic pump 20) P₁ : Control pressure

[0029]

(Equation 5)

Steps 103 to 10 in the ECU 50
The program processing of No. 6 corresponds to “time calculating means” described in the claims. (5) The duty ratio of the solenoid valve 41 is increased for the calculated filling time t _S , and the piston chamber 5 is moved as shown in the middle part of FIG.
Supplying hydraulic oil having a line pressure corresponding high control pressure P ₁ (step 107). Then clutch piston 4
Receives the force from the high-pressure hydraulic oil in the piston chamber 5, moves quickly from the position shown in FIG. 1 toward the clutch plate 1, and comes into contact with the clutch plate 1. 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. After increasing the duty ratio of the solenoid valve 41 to increase the hydraulic pressure of the hydraulic oil in the piston chamber, the gap between the clutch plate 1 and the partition plate 2 becomes substantially zero.
Is the filling time t _S. The processing performed during the filling time t _S is the filling control in step 107.

As shown in the middle part of FIG. 3, when the charging time t _S has elapsed and the time t ₁ in FIG. 3 has been reached, the duty ratio of the solenoid valve 41 is reduced and the control pressure supplied from the solenoid valve 41 to the piston chamber 5 is reduced. rapidly lowering the P _1. Thus, it is possible to prevent the clutch plate 1 from colliding with the partition plate 2 due to the pressing force of the clutch piston 4. Here, temporary between line pressure P _L is the fill time t _S by the engine operating conditions as shown in the lower part of FIG. 3 by supplying a rapidly operating oil in the piston chamber 5 of low pressure during the filling control at step 106 May decrease. Without considering the reduction in the line pressure P _L, that is, the source pressure command line pressure from ECU50 without solving the simultaneous equations of Equation 2 and Equation 4 to calculate the control pressure P _1, the number of the control pressure P ₁ substituted 2 calculates a consumption flow rate Q _P, when calculating the Q _P are substituted into Equation 1 the filling time t _S, the actual filling time t _S calculated from a high control pressure P ₁ than the reduction The required filling time is shorter than the required filling time. Therefore, the clutch piston 4 does not move to the predetermined position within the obtained filling time t _S , and the gap between the clutch plate 1 and the partition plate 2 cannot be reduced to substantially zero. If the next torque phase control and inertia phase control are performed in such a state, an engagement failure such as an increase in the time required for engagement or an insufficient engagement force occurs. As further shown in FIG. 4, the low-speed region of the engine irrespective of the filling control there is a region where the line pressure P _L than the command line pressure decreases the rotation speed.

The consumption flow rate Q _P , the line pressure P _L , and the charging time t _S obtained from the simultaneous equations of the equations (2) and (4) are independent of the reduction of the line pressure due to the filling of the working oil into the piston chamber 5 and the filling control. This is obtained in consideration of the decrease in the line pressure in the determined low engine speed region. Therefore, by calculating the charging time t _S from the line pressure P _L calculated in consideration of the reduced amount using the engine speed as a parameter, the charging control in step 107 is terminated even when the line pressure decreases from the command pressure. At this point, the gap between the clutch plate 1 and the partition plate 2 can be reduced to almost zero.

(6) The control pressure P _1, which has been sharply reduced, is
And the clutch plate 1 and the partition plate 2 are slid. This is the torque phase control (step 108). Also in the torque phase control, so generates a control pressure P ₁ the line pressure P _L obtained in step 105 as a source pressure, it is possible to perform highly accurate speed control.

In the torque phase control, the clutch plate 1 is a partition plate.
2, the clutch plate 1 and the partition plate 2 are sliding.
However, the torque of the clutch plate 1 is not transmitted to the partition plate 2.
No. Accordingly, the rotation speed of the partition plate 2, that is, the turbine rotation
Number N_tDoes not change. Torque phase control takes time t₁To t_Two
Between. (7) Control pressure P₁As you further increase the clutch plate
1 is transmitted to the partition plate 2. This is Figure 3
Time t shown in_TwoFrom which the rotation speed of the partition plate 2 is reduced.
Start a little. And the control pressure P₁Further increase the time
t_ThreeIs reached, the clutch plate 1 and the partition plate 2 are engaged.
(Step 109). This time t _TwoTo t_ThreeBetween
Is the inertia phase control. For inertia phase control
The line pressure P obtained in step 105_LFrom
Control pressure P as pressure₁High precision shifting
Control can be performed.

(8) If the clutch plate 1 and the partition plate 2 shift to the engaged state, the shift control ends (step 11).
0). In the embodiment of the present invention described above, the engine rotation speed, that is, the rotation speed of the hydraulic pump 20 is detected from a sensor as the rotation speed detecting means, and the rotation speed of the hydraulic pump 20 is substituted into a simultaneous equation to calculate the consumption flow rate Q. _P , line pressure P _L ,
The filling time t _S is calculated. Therefore, even if the line pressure P _L than the command line pressure by a decrease in the engine rotational speed is reduced, it is possible to calculate the reduced line pressure P _L, the appropriate gear based on the calculated line pressure P _L Control can be performed with high accuracy.

In this embodiment, the data stored in the memory
Using the value of the variable calculated from the number and oil temperature as a parameter
By substituting into Equations 1 to 5 for steady-state calculation, many
Easy and short time without storing the map in memory
Filling time t_SCan be calculated. Also, the oil temperature
Detect and use it as a parameter for filling time calculation
Thus, the viscosity coefficient μ and the damping coefficient C_P
Filling time t with high accuracy even if _SCan be calculated.

Equations (1) to (5) are used in common even if the engine is different. Therefore, by storing constants in a memory, the structure is uniquely determined even when developing a different engine. Only by changing the constants on the memory, the appropriate consumption flow rate Q _P , line pressure P _L , and charging time t _S can be calculated for each engine. 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 coefficient by previously assigning constants to the equations (1) to (5) and calculate a coefficient, and to calculate the filling time t _S by substituting a variable into the 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 changes in turbine speed, control pressure, and line pressure from the start of gear shifting.

4 is a characteristic diagram showing the relationship between the engine speed and the line pressure P _L.

[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)

──────────────────────────────────────────────────の Continued on front page (51) Int.Cl. ⁶ Identification code FI F16H 63:12

Claims (3)

[Claims] 1. A control device for an automatic transmission for switching a gear by engaging or disengaging at least one friction engagement element by hydraulic pressure of a working fluid, comprising: a hydraulic pump for discharging a working fluid; Line pressure control means for controlling the line pressure of the working fluid discharged from the hydraulic pump; rotation speed detection means for detecting the rotation speed of the hydraulic pump; and a consumption flow rate of the working fluid adjusted to the line pressure. For an automatic transmission, comprising: a consumption flow rate calculation means for calculating; and a line pressure calculation means for calculating a line pressure from an output of the rotation speed detection means and a consumption flow rate calculated by the consumption flow rate calculation means. Control device. 2. A hydraulic piston that engages a driving-side rotator and a driven-side rotator of the friction engagement element 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 pressure control valve, wherein the frictional engagement element is changed from a released state to an engaged state. 2. The control device for an automatic transmission according to claim 1, wherein, when the shift is performed, the drive hydraulic pressure to the hydraulic piston is increased by the hydraulic control valve. 3. When the friction engagement element is shifted from a release state to an engagement state, the hydraulic pressure to the hydraulic piston is increased by the hydraulic pressure control valve, and then the drive side rotating body and the A time calculating unit that calculates a filling time until the gap with the driven-side rotating body becomes substantially zero by substituting a parameter into a calculating formula, wherein the time calculating unit includes an output of the rotation speed detecting unit; 3. The control device for an automatic transmission according to claim 2, wherein the charging time is calculated by substituting the consumption flow rate calculated by the consumption flow rate calculation means into the calculation formula.