JPH07119815A - Oil pressure controller of automatic transmission - Google Patents

Abstract

PURPOSE:To more precisely predict the variation in an input revolving speed of a speed change mechanism so as to properly control the working oil pressure at the time of a speed change by controlling in an automatic transmission, the working oil pressure at the time of a speed change, which is supplied to a friction element that forms a power transmitting route in the speed change mechanism, based on both the input torque to the transmission and the variation in the input revolving speed of the speed change mechanism at the time before and after the speed change. CONSTITUTION:A revolving speed at the completion of a speed change is predicted based on both an output revolving speed of an automatic transmission 20, which is detected by an output revolving speed sensor 35, and a gear ratio of the speed stage after the speed shift. The variation in the turbine revolving speed at the time before and after the speed change is set, based on both the revolving speed at the completion of the speed change and a turbine revolving speed before the speed change, which is detected by a turbine revolving speed sensor 34.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a hydraulic control system for an automatic transmission, and more particularly to a change in operating torque for engaging frictional elements during shifting, input torque to the transmission, and change in input rotational speed of the transmission before and after shifting. The present invention relates to the control based on the quantity and the quantity.

[0002]

2. Description of the Related Art Automatic transmissions installed in automobiles are
It is configured so that the gear stage is automatically switched according to the operating state by switching the power transmission path of the transmission mechanism to which the engine output torque is input via the torque converter by selectively engaging a plurality of friction elements. The automatic transmission of this type is provided with a hydraulic control circuit that generates a line pressure for engaging the friction element. In that case, if the line pressure generated by this hydraulic control circuit is too low with respect to the input torque to the friction element, the torque capacity of the friction element becomes insufficient, and the required torque cannot be reliably transmitted. It will be. On the other hand, if the line pressure is too high with respect to the input torque to the friction element, a so-called shift shock occurs when the friction element is engaged, and the torque for driving the oil pump becomes unnecessarily large. The output will be consumed unnecessarily.

To address this, for example, Japanese Unexamined Patent Publication No. 4-337.
As disclosed in Japanese Patent No. 158, the line pressure in normal time is set according to the input torque to the automatic transmission, and at the time of shifting, the input rotational speed of the transmission mechanism before and after shifting, that is, the turbine rotational speed. It is considered to correct the line pressure based on the change amount of

According to this, the line pressure in the normal time appropriately corresponds to the input torque of the friction element,
It becomes possible to effectively reduce the pump drive loss of the engine output while ensuring the required torque capacity of the friction element, and at the time of gear shifting, the moment of inertia caused by the rotational change of the gear shift mechanism is appropriately absorbed. Therefore, it is expected that a good shift feeling will be obtained.

[0005]

However, even in the prior art described in the above publication, there still remain the following problems to be solved.

That is, in the prior art described in the above publication, the change amount ΔN of the turbine rotation speed Nt before and after the shift is changed.
t is calculated using the following relational expression, and the line pressure at the time of shifting is corrected and controlled based on this change amount ΔNt.

ΔNt = Nts (Gs / Ge-1) Here, Nts is the turbine speed at the time of shift determination,
Gs indicates a gear ratio before shifting, and Ge indicates a gear ratio after shifting.

By the way, there is a so-called torque demand shift as a form of shift control in this type of automatic transmission.
This is performed so as to downshift the shift stage when the operating state passes through the shift down line that constitutes the shift pattern as the engine load increases, such as when the accelerator pedal is depressed.

In this case, in the torque demand shifting in the normal drive state in which the vehicle is driven by the engine output, as shown by the solid line in FIG. 13, the turbine rotation speed Nt changes from the shift start rotation speed Nts1 to the shift end rotation speed Nte. In the torque demand shifting during inertial traveling of the vehicle with inertia of the vehicle body, the turbine rotation speed Nt is, for example, from the idle rotation speed Nts0 (<Nts1) to The speed changes to the speed change end rotation speed Nte.

However, if the turbine revolution speed change amount ΔNt is predicted by substituting the idle revolution speed Nts0 into the value of Nts in the above relational expression, the shift end revolution speed Nte 'indicated by the predicted value ΔNt (p) is shown in the figure. The value is as shown by the chain line. That is, the predicted value ΔNt (p) of the change amount ΔNt
Will be calculated to be a value smaller than the actual value ΔNt (r). Therefore, the line pressure at the time of shifting corrected based on the predicted value ΔNt (p) does not correspond to the required hydraulic pressure, and the shift feeling is deteriorated such as the occurrence of a large shock.

According to the present invention, the working hydraulic pressure at the time of gear shifting, which is supplied to the friction element forming the power transmission path in the gear shifting mechanism,
This is to address the above problem in an automatic transmission that is controlled based on the input torque to the transmission and the amount of change in the input speed of the speed change mechanism before and after shifting. It is an object of the present invention to more accurately predict the amount of change in the hydraulic pressure and to appropriately control the hydraulic pressure during gear shifting.

[0012]

SUMMARY OF THE INVENTION According to the invention of the present application, an operating hydraulic pressure at the time of gear shifting, which is supplied to a friction element forming a power transmission path in the gear shifting mechanism, is input torque to the gearbox and a gear shifting mechanism before and after gear shifting. In an automatic transmission that is controlled based on the amount of change in the input rotation speed, a shift end rotation that predicts the shift end rotation speed based on the output rotation speed of the transmission and the gear ratio of the shift stage after the shift. A number predicting means, and an input rotational speed change amount setting means for setting the change amount of the input rotational speed based on the shift end rotational speed predicted by the predicting means and the input rotational speed before the shift. Characterize.

[0013]

According to the above construction, the following operation can be obtained.

That is, the shift end rotation speed is predicted based on the output rotation speed of the transmission and the gear ratio of the shift stage after the shift, and based on the predicted shift end rotation speed and the input rotation speed before the shift. Since the change amount of the input speed of the speed change mechanism before and after the shift is set, the change amount of the input speed can be accurately obtained even during the speed change from the inertia running state. Since the operating oil pressure is set appropriately corresponding to the oil pressure, a good gear shift operation with less shock is performed.

[0015]

EXAMPLES Examples of the present invention will be described below.

As shown in FIG. 1, an engine 10 according to this embodiment has an air flow sensor 12 for detecting an intake air amount from an upstream side in an intake passage 11 and a first throttle valve which is automatically controlled to open and close. 13 and a second throttle valve 14 that is controlled to be opened and closed by a driver's operation, and fuel injection valves 15 ... 15 for each cylinder.
, And spark plugs 16 ... 16 are provided. Also,
An automatic transmission 20 that constitutes a power unit together with the engine 10 includes a torque converter 21 connected to an output shaft 17 of the engine 10, a transmission mechanism 22 to which the output torque (turbine torque) is input, and a plurality of friction elements ( And a hydraulic control circuit 23 for switching the gear ratio (gear stage) of the transmission mechanism 22 by selectively supplying a line pressure to (not shown).

Further, a control unit 30 for various controls on the engine 10 and the automatic transmission 20 is provided, and a signal from the air flow sensor 12 in the intake passage 11 from the engine 10 is provided to the control unit 30. , A signal from a throttle opening sensor 31 that detects the opening of the second throttle valve 14, a signal from an engine rotation speed sensor 32 that detects the rotation speed of the engine output shaft 17, and a water temperature that detects the temperature of the cooling water. A signal from the sensor 33 is input, and the turbine speed sensor 3 that detects the output speed (turbine speed) of the torque converter 21 is input from the automatic transmission 20.
4, a signal from an output rotation speed sensor 35 that detects the output rotation speed of the transmission mechanism 22, and a signal from an oil temperature sensor 36 that detects the temperature of the hydraulic oil are input.

Then, the control unit 30 controls the fuel injection by the fuel injection valves 15 ... 15 in the engine 10 and the spark plug 16 based on the input signals.
Ignition control for 16 and first throttle valve 13
Throttle control for the automatic transmission 2
For 0, shift control by the solenoid valves 24 ... 24 provided in the hydraulic control circuit 23 and line pressure control by the duty solenoid valve 25 also provided in the hydraulic control circuit 23 are performed. .

Here, the structure of the portion of the hydraulic control circuit 23 related to the line pressure control will be described.

As shown in FIG. 2, the hydraulic control circuit 23 includes a regulator valve 42 for adjusting the pressure of the hydraulic oil discharged from the oil pump 41 to a predetermined line pressure for line pressure control, and the regulator valve 42. A throttle modulator valve 43 that supplies a control pressure to the valve 42 is provided. This throttle modulator valve 43
A constant pressure line 46 led from a main line 44 through which the hydraulic oil is discharged from the oil pump 41 via a reducing valve 45 for reducing the hydraulic oil to a constant pressure.
Is connected to the throttle modulator valve 4
3, a pressure increasing line 47 is led to a pressure increasing port 42a provided at one end of the regulator valve 42, and a pilot line 48 branched from the constant pressure line 46 is provided at one end of the throttle modulator valve 43. It is connected to the control port 43a.

The duty solenoid valve 25 for controlling the line pressure shown in FIG. 1 is installed in the pilot line 48.
By introducing the pilot pressure according to the duty ratio (ON time ratio in one ON, OFF cycle) of the control port 43a of the throttle modulator valve 43, the constant pressure supplied from the line 46 is supplied to the pilot. The pressure or the pressure corresponding to the above duty ratio is adjusted, and this hydraulic pressure is adjusted by the line 47 to the regulator valve 4
It is adapted to be supplied to the second pressure increasing port 42a.
Therefore, the line pressure whose pressure is adjusted by the regulator valve 42 becomes a pressure according to the duty ratio.

Next, the line pressure control performed by the control unit 30 will be described with reference to the flow chart shown in FIG.

That is, the control unit 30 is
In step S1, the intake air amount, the opening of the second throttle valve 14, the engine speed, the cooling water temperature, the turbine speed, and the output rotation of the speed change mechanism 22, which are indicated by the signals from the sensors 12 and 31 to 36 shown in FIG. 1, respectively. Enter the number and the oil temperature of the hydraulic oil. Next, at step S2, the current driving state of the vehicle is a positive driving state (e <1) depending on the ratio of the turbine rotational speed Nt to the engine rotational speed Ne, that is, the value of the speed ratio e (= Nt / Ne) of the torque converter 21. ) Or reverse driving state (e> 1), that is, whether the vehicle is driven by the output of the engine 10,
Determine whether the vehicle is traveling due to inertia.

Then, in the case of the normal drive state, in steps S3 to S6, the output torque Te of the engine 10 and the output torque (turbine torque) of the torque converter 21.
Tt is sequentially calculated, and when in the reverse driving state,
In step S7, the reverse drive torque Tr input from the output side to the speed change mechanism 22 of the automatic transmission 20 is calculated.

In this case, the engine torque Te in the normal drive state is 2 with respect to the ignition timing Ig as shown in FIG.
It can be approximated as a quadratic function, and if this is expressed by a formula, the following approximation formula is obtained.

Te = −A · (Ig−B)² + C ... Here, A, B, and C depend on the operating state of the engine 10.
It is a coefficient that changes and is shown in FIGS. 5 (a), (b), and (c).
As shown, the engine speed Ne and the air filling effect are shown.
It is preset as a map with the ratio Ce as a parameter.
ing. Therefore, in step S3, first,
The current air from the engine speed Ne and the intake air amount Q
The filling efficiency Ce1 is calculated, and the air filling efficiency Ce
1 and the engine speed Ne1 at the present time,
From the top to respond to the current operating condition of the engine 10.
The numbers A1, B1 and C1 are obtained. And these coefficients A
1, B1, C1 and the current ignition timing Ig are expressed by the above equations.
The engine torque Te is calculated by substituting. This
As a result, the actual engine torque can be accurately estimated.
It will be.

Next, in step S4, it is determined whether or not the shift-down shift is performed. If it is determined that the shift-down shift is not performed, the process proceeds to step S5, based on the current speed ratio e of the torque converter 21. From the map showing the torque increase characteristic of FIG. 6, the torque increase ratio t (= Tt
/ Te), the torque increase ratio t and the above step S
The turbine torque Tt is obtained by substituting the engine torque Te obtained in 3 into the following relational expression.

Tt = Te · t On the other hand, when it is determined in step S4 that the downshift is performed, the torque increase ratio t and the engine torque Te are substituted into the following relational expression to determine the turbine torque Tt. Ask.

Tt = (Te−Idω / dt) · t Here, I represents the moment of inertia and ω represents the angular velocity, and therefore the second term on the right side of the above equation represents the change in torque due to the change in rotation. .

Further, in step S7, the reverse drive torque Tr input from the output side to the speed change mechanism 22 of the automatic transmission 20 is calculated from the output speed No of the speed change mechanism 22 corresponding to the vehicle speed and the gear position. .

In this way, the turbine torque Tt and the reverse drive torque Tr, which are the basis for setting the line pressure, are correctly calculated.

Next, in step S8, the control unit 30 calculates the required line pressure Pn from the shift speed and the turbine torque Tt or the reverse drive torque Tr obtained as described above. This required line pressure Pn is, as shown in FIG. 7, the line pressure of a friction element that requires the highest line pressure for a constant turbine torque Tt or reverse drive torque Tr for each shift stage. , The proportional constants K1 to K4, Kr between the turbine torque Tt or the reverse drive torque Tr are calculated in advance, and the proportional constants for the current shift stage are multiplied by the turbine torque Tt.

Then, in step S9, the control unit 30 corrects the required line pressure Pn according to the engine speed Ne to obtain the baseline pressure Pb.
In the correction of the line pressure, when the engine speed Ne is low, the discharge amount of the oil pump is insufficient, so that the target line pressure cannot be obtained even if the control amount corresponding to the predetermined target line pressure is output. In contrast to this, when the engine speed Ne is low in advance, the required line pressure Pn is corrected to a higher value, and the corrected line pressure is used as the baseline pressure Pb. Specifically, the current engine speed Ne1 and the required line pressure Pn1 are shown in the map shown in FIG. 8 with the engine speed Ne and the required line pressure Pn as parameters.
Then, the current baseline pressure Pb1 is calculated.

As described above, the baseline pressure Pb required for the friction element to transmit the predetermined input torque is obtained as the minimum necessary value without excess or deficiency.

Next, the control unit 30 determines in step S10 whether or not the shift flag Fs for switching the shift stage of the automatic transmission 20 is set. When the shift flag Fs is not set, the above-mentioned step S10 is performed. From step S11 to step S11, the baseline pressure Pb is directly used as the target line pressure Po, and the duty ratio D of the duty solenoid valve 25 corresponding to the target line pressure Po is calculated in step S12. In that case, since the line pressure obtained for the same duty ratio D varies depending on the oil temperature of the hydraulic oil, as shown in FIG. Depending on the target line pressure P
The duty ratio D corresponding to o is obtained.

Then, in step S13, the control unit 30 outputs the duty signal of the duty ratio D set as described above to the duty solenoid valve 25. As a result, the regulator valve 42 of the hydraulic control circuit 23 shown in FIG. 2 generates a line pressure that accurately corresponds to the input torque of the friction element at that time.

On the other hand, when the control unit 30 determines in step S10 that the shift flag Fs is set, it moves to step S14 and calculates the amount of change ΔNt of the turbine speed Nt before and after the gear shift.

That is, the control unit 30 takes in the turbine rotational speed Nts and the output rotational speed Nos at the present time, substitutes them into the following relational expression, and sets the calculation result as the turbine rotational speed change amount ΔNt. .

ΔNt = Nos · Ge−Nts ... Here, Ge represents the gear ratio of the shift stage after the shift.

Therefore, even if the engine speed Ne or the turbine speed Nt is reduced to the idle speed when the vehicle is coasting, the output speed No reflects the vehicle speed of the vehicle. The change in the turbine speed Nt can be accurately predicted.

Next, the control unit 30 executes step S15 to determine whether the shift signal is a shift up signal or a shift down signal. If it is a shift-up signal, step S16 is calculated. If it is a shift-down signal, step S17 is calculated.

That is, at the time of shift-up, the increase line pressure ΔP at shift is set from the map of FIG. 10 on the basis of the turbine rotation speed change amount ΔNt, while at the time of shift-down, it is based on the turbine rotation speed change amount ΔNt. As a result, the increased line pressure ΔP during shifting is set from the map of FIG. In that case, in the map during shift up, the increase line pressure ΔP increases as the turbine rotation speed change amount ΔN increases, and in the map during shift down, the increase line pressure ΔP decreases as the turbine rotation speed change amount ΔN decreases. It is set to be large. This is because when the amount of change in turbine speed ΔNt increases during a shift up, the moment of inertia that must be absorbed by the friction elements increases, so in order to complete the gear shifting operation within a certain period of time, increase the line pressure. This is because the torque capacity of the friction element must be increased. Also, at the time of downshift, the moment of inertia is absorbed by the engine and the friction element,
This is because when the turbine rotation speed change amount ΔNt is small, the increase amount of the engine rotation speed is small and the ratio of the inertia moment to the friction element is large.

At the time of upshifting, the control unit 30 calculates the target line pressure Po at step S11 by adding the increased line pressure ΔP during shifting to the baseline pressure Pb obtained at step S9. As in the case of the non-shift, the operations of calculating the duty ratio D and outputting the duty signal are performed according to steps S12 and S13.

Further, at the time of downshifting, the control unit 30 determines the target line pressure Po by adding the shift increasing line pressure ΔP to the baseline pressure Pb determined in step S9 in step S11. , In this case also, steps S12 and S
According to No. 13, each operation of calculating the duty ratio D and outputting the duty signal is performed.

In that case, in the downshift from the inertia running state, conventionally, as shown in FIG. 11, the calculated value ΔNt of the turbine rotation speed change amount ΔNt is calculated.
1 becomes a value smaller than the actual value ΔNt (r), and the set value ΔP1 of the increased line pressure ΔP at the time of shifting is set to a large value accordingly, so that as shown by the arrow (a) in FIG. The line pressure of is also controlled to be higher than the required hydraulic pressure.

However, in this embodiment, the calculated value ΔNt2 of the turbine rotation speed change amount ΔNt is the actual value ΔNt.
It almost coincides with (r). Therefore, the set value ΔP2 of the increased line pressure ΔP at the time of gear shift is set to a value substantially equal to the required value, and the line pressure at the time of gear shift properly requires the required hydraulic pressure as shown by the arrow (a) in FIG. It will be controlled to an appropriate value that reflects it. As a result, a shift shock due to the line pressure being too high during shifting is avoided.

[0047]

As described above, according to the present invention, the operating oil pressure at the time of gear shifting, which is supplied to the friction element forming the power transmission path in the gear shifting mechanism, is the input torque to the gearbox and the gear shifting mechanism before and after the gear shifting. In the automatic transmission, which is controlled based on the change amount of the input rotation speed of, while predicting the shift end rotation speed based on the output rotation speed of the transmission and the gear ratio of the shift stage after the shift, Since the change amount of the input speed is set based on the predicted shift end speed and the input speed before the shift, the change amount of the input speed is changed even during the shift from the inertia running state. It will be required with high precision, and this will set the operating oil pressure that appropriately corresponds to the required oil pressure.
Good shift operation with less shock will be performed.

[Brief description of drawings]

FIG. 1 is a system diagram showing a control system of an embodiment.

FIG. 2 is a circuit diagram showing a main part of a hydraulic control circuit.

FIG. 3 is a flowchart showing the operation of line pressure control.

FIG. 4 is a characteristic diagram of engine torque with respect to ignition timing.

FIG. 5 is a map for obtaining each coefficient in an approximate expression of engine torque.

FIG. 6 is a map of torque increase ratio characteristics of the torque converter.

FIG. 7 is a map of a line pressure correction coefficient.

FIG. 8 is a map for obtaining a baseline pressure.

FIG. 9 is a map for obtaining a duty ratio.

FIG. 10 is a map for obtaining an increased line pressure during a shift when shifting up.

FIG. 11 is a map for obtaining an increased line pressure during a shift during a downshift.

FIG. 12 is a time chart showing the operation of the embodiment.

FIG. 13 is a time chart diagram showing conventional problems.

[Explanation of symbols]

 20 automatic transmission 22 transmission mechanism 30 control unit 34 turbine speed sensor 35 output speed sensor

Claims (1)

[Claims] 1. A hydraulic pressure during a gear shift, which is supplied to a friction element forming a power transmission path in the gear shift mechanism, is based on an input torque to the transmission and a change amount of an input rotational speed of the gear shift mechanism before and after the gear shift. A hydraulic pressure control device for an automatic transmission, wherein the shift end rotation number predicting means predicts a shift end rotation number based on an output rotation number of the transmission and a gear ratio of a shift stage after a shift. And input rotation speed change amount setting means for setting the change amount of the input rotation speed based on the shift end rotation speed predicted by the prediction means and the input rotation speed before shifting. Hydraulic control device for automatic transmission.