JP2002295661A - Gear shift control device of automatic transmission - Google Patents

Abstract

PROBLEM TO BE SOLVED: To make feedback control on the rotational speed of a turbine, during an inertia phase to a target rotation speed at a high response, while ensuring superior robustness. SOLUTION: The rotational speed of the turbine during the inertia phase is controlled by a sliding mode control, based on a switching function σ. Here, a border layer, in which a control input is continuously approximated, is introduced to a switching surface, and a width ϕ of the border layer is made narrower than a reference value at a predetermined time, immediately after the start of the inertia phase (feedback control). A convergent responsivity to the switching surface is enhanced, by narrowing the width ϕ of the border layer, whereas at the sliding mode after arriving at the switching surface, superior robustness can be ensured by enlarging the width ϕ of the border layer to a reference value.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a shift control device for an automatic transmission, and more particularly to a sliding mode control technique suitable for feedback control of an input shaft rotation speed during an inertia phase.

[0002]

2. Description of the Related Art Conventionally, feedback is performed by proportional / integral / differential operation based on a deviation between a target value and an actual value so that an input shaft rotation speed during a gear shift or a rate of change of the input shaft rotation speed follows a target. 2. Description of the Related Art There is known a shift control device for an automatic transmission configured to perform control (see Japanese Patent Application Laid-Open Nos. 2000-110924 and 11-31317).

In the feedback control based on the above-described proportional / integral / differential operation (PID control), A
Since the requirements of the feedback gain differ depending on the conditions such as the temperature of TF (Automatic Transmission Fluid) and the vehicle speed, a gain map adapted for each condition such as the ATF temperature and the vehicle speed is provided. Is searched for and used.

[0004] Therefore, there is a problem that the number of feedback gains that need to be adapted is large and the number of adaptation steps becomes enormous.
Since the gain is fixed, it is not possible to cope with a state change during gear shifting, and there is a possibility that convergence to a target may be deteriorated. As a control for compensating for the above-mentioned disadvantages of the PID control, a sliding mode control which has excellent robustness and a relatively simple controller design is known (Japanese Patent Laid-Open No. 2000-2000).
-35120).

[0005]

In the above-mentioned sliding mode control, it is general to introduce a boundary layer whose control input is continuously approximated to the switching surface in order to prevent chattering. Is wider, the control system becomes more robust. However, when the width of the boundary layer is increased, the response is reduced, and in particular, there is a problem that convergence to the switching surface immediately after the start of feedback control is delayed. .

SUMMARY OF THE INVENTION The present invention has been made in view of the above-mentioned problems, and provides a shift control device for an automatic transmission that can converge to a switching surface with good response immediately after the start of feedback control while ensuring excellent robustness. The purpose is to do.

[0007]

Therefore, according to the first aspect of the present invention, a sliding mode control based on a switching function using the state quantity of the automatic transmission as a variable is performed so as to make the state quantity during shifting match the target value. In a shift control device for an automatic transmission, which is configured to control hydraulic pressure supply to an engagement element, and in which a boundary layer in which a control input is continuously approximated to a switching surface is introduced, in a predetermined period immediately after the start of feedback control, The width of the boundary layer was reduced.

With this configuration, in the predetermined period immediately after the start of the feedback control, the sliding mode control is performed with a narrower boundary layer width than after the predetermined period has elapsed. According to the second aspect of the invention, the width of the boundary layer is gradually increased in accordance with the elapsed time from the start of the feedback control.

According to this configuration, the width of the boundary layer is relatively narrow immediately after the start of the short elapsed time from the start of the feedback control, and the width of the boundary layer is increased as time elapses. In the invention according to claim 3, the width of the boundary layer is held at an initial value smaller than a reference value within a predetermined holding time from the start of the feedback control, and after a lapse of the holding time, a predetermined increase control time has elapsed. It was configured to gradually increase to the reference value.

According to this configuration, the width of the boundary layer is held at an initial value smaller than the reference value until a predetermined holding time elapses from the start of the feedback control, and high responsiveness is exhibited during the holding time. After that, the control value is gradually increased with the increased control time to a reference value at which excellent robustness can be secured. According to the fourth aspect of the present invention, the hydraulic pressure supplied to the friction engagement element is feedback-controlled so that the input shaft rotation speed matches the target rotation speed in the inertia phase during the shift.

With this configuration, when the feedback control of the input shaft rotation speed (turbine rotation speed) is started in the inertia phase, the sliding mode control is performed with a relatively narrow boundary layer width in the initial predetermined period.

[0012]

According to the first aspect of the present invention, the responsiveness is enhanced by making the width of the boundary layer relatively narrow immediately after the start of the feedback, so that the response converges on the switching surface with good response. As a result, an excellent robustness is exhibited by expanding the range, so that there is an effect that the state quantity during shifting can be feedback-controlled to the target with high responsiveness and high stability.

According to the second and third aspects of the present invention, the width of the boundary layer is changed according to the elapsed time from the start of the feedback control, so that the width of the boundary layer is narrowed by a time necessary and sufficient for convergence on the switching surface. Therefore, there is an effect that convergence responsiveness to the switching surface can be improved without sacrificing robustness. According to the fourth aspect of the present invention, there is an effect that the input shaft rotational speed during gear shifting can be feedback-controlled to the target rotational speed with high responsiveness and high stability.

[0014]

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 (not shown) is transmitted to a transmission mechanism 2 via a torque converter 1.

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

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

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 speed change mechanism 2 having the above-described configuration, as shown in FIG.
This is realized by a combination of the release states.

In FIG. 2, a circle indicates a fastened state, and a part without a symbol indicates a released state.
A fastening state indicated by a black circle of & R / B indicates fastening in only one range. Each of the clutches and brakes (friction engagement elements) is engaged and disengaged by supply hydraulic pressure. The supply hydraulic pressure for each clutch and brake is controlled by a solenoid valve unit 11 shown in FIG.
Are individually controlled by a solenoid valve included in the control unit.

An A / T controller 12 for controlling each solenoid valve 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 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 control by the T controller 12 will be described with reference to the time chart of FIG. 5 by taking an example of an upshift (hereinafter, referred to as a power-on upshift) in a state where the driving torque of the engine 20 is applied. This will be described with reference to the flowchart of FIG.

In the flowchart of FIG. 4, 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 the vehicle speed VSP and the accelerator opening (throttle opening). For example, a search is made from the current shift stage and the shift map. Different from the gear stage, and
If it is 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 output shaft rotation speed No [rpm] of the transmission mechanism is changed to the gear ratio before the shift (gear ratio = turbine rotation speed Nt (input shaft rotation speed). ) / Output shaft rotation speed No)
The shift to the torque phase is determined by determining whether or not the turbine rotation speed Nt [rpm] is higher than the sum of the reference turbine rotation obtained by multiplying the reference turbine speed and the hysteresis value HYS stored in advance. I do.

In the present embodiment, the release control is advanced earlier than the engagement control to induce the idling, and the transition to the torque phase is determined based on the occurrence of the idling. is there. Until the transition to the torque phase is determined in step S2, step S3
To execute the preparation phase process.

In the preparation phase process of step S3, the command oil pressure of the friction engagement element (hereinafter, referred to as a release-side friction engagement element) to be released from the engaged state before the shift is gradually reduced to the release initial pressure in a predetermined time. Then, while gradually decreasing the release initial pressure at a predetermined speed from the release initial pressure, the command hydraulic pressure of a friction engagement element (hereinafter, referred to as engagement-side friction engagement element) to be engaged from the release state before the shift is changed to the standby pressure after the precharge. To be held.

When the shift to the torque phase is determined in step S2, the process proceeds to step S4, where the gear ratio is changed to F / B.
(Feedback) It is determined whether or not the gear ratio has changed in the upshift direction beyond the start gear ratio. Then, the torque phase process of step S5 is performed until the gear ratio exceeds the F / B start gear ratio and changes in the upshift direction.

In the torque phase process, following the preparation phase process, the command oil pressure of the disengagement side frictional engagement element is gradually reduced, and the command oil pressure of the engagement side frictional engagement element is gradually increased from the standby pressure. If it is determined in step S4 that the gear ratio has exceeded the F / B start gear ratio, the process proceeds to step S6, in which the gear ratio is changed to the F / B end gear ratio (<F / B start gear ratio).
Is determined.

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. In the inertia phase process, the command oil pressure of the disengagement side frictional engagement element is decreased stepwise to 0, while the command oil pressure of the engagement side frictional engagement element is set at the target turbine rotation speed Nt (input shaft rotation speed). Feedback control is performed so as to match the rotation speed.

The target rotational speed is set as a value that gradually changes from the turbine rotational speed at the start of the inertia phase to a turbine rotational speed that matches the gear ratio after the shift within a predetermined shift time. When the gear ratio is F
/ B end gear ratio is smaller than the gear ratio in step S
When the determination is made in step 6, the process proceeds from step S6 to step S8, and it is determined whether or not a predetermined time TIMER7 has elapsed from the time when the gear ratio first became smaller than the F / B end gear ratio.

If the time is within the predetermined time TIMER7, the flow advances to step S9 to perform an end phase process.
In the end phase process, the command oil pressure of the disengagement side frictional engagement element is maintained at the oil pressure at the end of the inertia phase (= 0), while the command oil pressure of the engagement side frictional engagement element is increased to the maximum pressure.

Here, the feedback control of the turbine rotational speed Nt in the inertia phase process in step S7 will be described in detail. FIG. 6 is a block diagram of a system for performing the feedback control.
A deviation err between the target rotation speed and a signal indicating the actual turbine rotation speed Nt taken out from the drive system 102 is input to the controller 101, and the hydraulic system 103 (the engagement-side friction clutch) is controlled by sliding mode control based on the deviation err. A control signal u (instruction oil pressure) output to a solenoid valve that controls the oil pressure of the combined element is calculated.

As described above, by performing feedback control of the turbine rotation speed by the sliding mode control, a control system having excellent robustness can be formed, and the number of steps for matching the gain can be greatly reduced, and the design of the control system can be reduced. It can be done easily. In addition, the minimum dimension observer 1
04 (Corona “Sliding mode control” 1996
(See pages 160-178, issued April 10, 2008)
A mechanism for estimating, from measurable data, a state quantity that cannot be directly measured among system state quantities (state variables) required for realizing the sliding mode control, and outputs the estimation result to the SMC controller 101. I do.

In designing the SMC controller 101, the drive system 102 and the hydraulic system 103 were modeled as follows. [Hydraulic system model]

[0033]

(Equation 1)

The hydraulic system model has an ATF temperature of 80.
° C as the nominal model of the hydraulic system. [Drive system model]

[0035]

(Equation 2)

[0036] Note that in Equation 2, T _t is turbine torque, T _REV is reverse clutch torque, the T _HC is high clutch torque, T _LC is low clutch torque, Omegaod
ot (“dot” indicates a superscripted point of a modifier symbol; hereinafter the same), ωtdot indicates turbine angular acceleration (input shaft angular acceleration), and I indicates inertia. FIG. 7 shows a system block diagram based on the above model formula.

In FIG. 7, Frtn is the return spring pressure of the engagement side frictional engagement element, u is the control input, r is the target rotation speed, and err is the turbine rotation speed Nt and the target rotation speed r.
Is the deviation from In the above system model, the state variable (state quantity) x used for the switching function σ = Cx is

[0038]

(Equation 3)

[0039] By including x5 which is an integral value of the deviation in the state variable (state quantity) x, the steady deviation can be absorbed. Note that x1, x2, x3, x4 are state variables shown on the system block diagram of FIG.

At this time, the state equation is described as follows.

[0041]

(Equation 4)

X1, x2, x3, x4 of the above state variables x
Cannot be measured directly, so that the minimum dimension observer 10
4 Then, the switching function σ in the phase space by the state variable x is expressed as follows: σ = Cx + βr C = [c1, c2, c3,1, c4], x = [x1, x2, x3, x4, x5] σ = (c1 × x1 + c2 · X2 + c3 · x3 + x4 + c4 · x5) +
βr.

The second term βr of the switching function σ is a term for adding a zero point, and the control parameter β for multiplying the target rotational speed r is changed according to the ATF temperature.
In a controller that does not add a zero point, the turbine torque T
Since the influence of t cannot be suppressed and the transient response deteriorates, the transient response is improved by adding a zero point by βr and increasing the gain in the high frequency region.

Then, the control input u (instruction oil pressure) was set as follows. u = _ueq + _unl + _uf

[0045]

(Equation 5)

U _eq is an equivalent control input when the effects (disturbance) of the return spring pressure, turbine torque, low clutch torque and output shaft angular acceleration are ignored. Wherein u _nl is a switching function σ operation for restraining system switching surface on the basis of the (non-linear control input), where, by introducing a boundary layer switching surface, the continuity approximating the switching function in the boundary layer Is used as a saturation function to suppress chattering.

Incidentally, a configuration using a smoothing function instead of the saturation function may be adopted. Further, _uf is an offset input (disturbance compensation input) for eliminating the effects of the return spring pressure, the turbine torque, the low clutch torque, and the output shaft angular acceleration, which are assumed as disturbances. The _unl is
Feedback gain (nonlinear gain: relay gain)
Is q, outside the saturation region (σ> Φ), u _nl = −q
Σ / | σ |, within the saturation region (σ <Φ), u _nl = −
(Q / Φ) · σ.

Note that Φ is a value that defines the boundary layer width (inside and outside the saturated layer) as described above.
Is performed as a gain. Differentiating both sides of the switching function σ = Cx + βr,

[0049]

(Equation 6)

And solving for u at σdot = 0 gives

[0051]

(Equation 7)

Is as follows. Therefore, the control input is separated into the equivalent control input u _eq and the disturbance offset input _uf ,
Further, a non-linear control input u _nl continuously approximated to suppress chattering is adopted. When the switching function σ and the control input u are set as described above, the system when the sliding mode occurs is as follows.

[0053]

(Equation 8)

The transfer function G (s) from the target rotation speed r to the actual turbine rotation speed Nt is

[0055]

(Equation 9)

Is as follows. Here, the transfer function G of Equation 9 is obtained.
So that the response in (s) is the response characteristic of the request,
The control parameters c1, c2, c3, c4 (switching parameters) are determined. Specifically, it was designed such that the closed-loop dynamics was −60 (heavy pole) −80 (heavy pole) [rad / s] at an ATF temperature of 80 ° C. The zero point was set to -5 [rad / s].

At this time, in the present embodiment, the control parameters c1, c2, c3, c4 are as follows. c1 = −0.00467 c2 = 29200 c3 = 280 c4 = −0.00838 FIG. 8 shows a control input u = u _eq based on the switching function σ.
SMC controller 101 for calculating + u _nl + u _f and minimum dimension observer 104 for estimating x1, x2, x3, x4
FIG. 3 is a block diagram showing the configuration of FIG.

The minimum dimension observer 104 has an actual turbine rotation speed Nt, a control input u of the friction engagement element,
The return spring pressure, the estimated value of the turbine torque, the low clutch torque, and the output shaft angular acceleration are input, and the state variables x1, x2, x3, x4 are estimated based on these data. The estimated state variables x1, x2, x3, x4 are:
Target turbine speed r, estimated turbine torque, low clutch torque, output shaft angular acceleration, shift up / down signal Up / Down, power on / off signal PowerOn / Off,
A rotation error err, which is a difference between the actual turbine rotation speed Nt and the target r, and a phase signal Ph-flg are input to the SMC controller 101, and a control input u is calculated based on these signals.

FIG. 9 shows an operation block of the control input u in the SMC controller 101. The control parameters c1, c2, c3, c4, c4, and c3 are assigned to the state variables x1, x2, x3, x5 and the target rotational speed r. is multiplied by β, and the result of the multiplication and the state variable x4 are summed to calculate the value of the switching function σ = Cx + βr. Then, the nonlinear control input _unl is calculated using a saturation function based on the value of the switching function.

The return spring pressure F which is a disturbance
rtn, turbine torque Tt, the low clutch torque T _LC and the output shaft angle disturbance offset which is calculated by the acceleration ωodot is summed with the u _f is calculated, and further, error er
The equivalent control input _ueq is calculated by multiplying r, the state quantities x2, x3, x4, and the target rotational speed r by the respective gains and summing them.

FIG. 10 is a flowchart showing the flow of feedback control of the turbine rotational speed Nt by the above-mentioned sliding mode control. Step S21
Then, it is determined whether or not the shift is being performed, and when the shift is started, the process proceeds to step S22. In step S22, it is determined whether or not the inertia phase has been reached. When the inertia phase has been reached, the process proceeds to step S23.

In step S23, a termer for measuring an elapsed time from the start of the inertia phase is counted up. In step S24, a table in which the boundary layer width Φ is stored in advance for each timer value is referenced, and the timer value at that time (elapsed time from the start of the inertia phase)
Is searched for the boundary layer width Φ.

In the table of the boundary layer width Φ, the boundary layer width Φ is held at an initial value smaller than the reference value until a predetermined holding time elapses from the start of the inertia phase. Thereafter, the control value is gradually increased from the initial value to the reference value at a constant speed during the increase control time, and after the holding time + the increase control time elapses, thereafter, the reference value is set as the boundary layer width Φ.

That is, the boundary layer width Φ is narrowed during a predetermined period immediately after the start of the feedback control, thereby improving the convergence responsiveness to the switching surface immediately after the start of the feedback control. In the sliding mode state, excellent robustness is secured by widening the boundary layer width Φ. Step S25
Then, return spring pressure Frnt, turbine torque T
t, low clutch torque T _LC and the output shaft angular acceleration ωodot
Are input, and in step S26, the state variables x1, x2, x3, x4 are estimated based on these.

In step S27, a deviation err (err = Nt-N) between the actual turbine rotational speed Nt and the target rotational speed r is obtained.
r) is calculated, and in step S28, the deviation err is added to the integration result up to that time to update the integration value (state variable x5). In step S29, the boundary layer width Φ set in step S24 and the control parameters c1, c2, c3, c4
(Switching parameters), the state variables x1, x2, x3, x4, x
5. Calculate the control input u based on disturbances such as the target rotation speed r and the return spring pressure Frtn.

In step S30, the control input u is output. In step S31, the end of the inertia phase is determined, and the process returns to step S23 to continue the feedback control (sliding mode control) until the end is determined. Then, if it is determined in step S31 that the inertia phase has ended, this routine is ended as it is.

In the above embodiment, the feedback control of the turbine rotational speed in the inertia phase at the time of the upshift is described as an example. However, the feedback control of the turbine rotational speed in the inertia phase at the time of the downshift can be performed in the same manner. Is clear.
Further, the switching function σ and the control input u are not limited to those described above.

[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 flowchart illustrating shift control in the embodiment.

FIG. 5 is a time chart showing changes in each phase and a command oil pressure in the shift control according to the embodiment;

FIG. 6 is a block diagram illustrating a feedback control system of the turbine rotational speed according to the embodiment.

FIG. 7 is a block diagram showing a model of the feedback control system.

FIG. 8 is a block diagram showing input / output signals of the feedback control system.

FIG. 9 is a block diagram showing an arithmetic circuit of a control input in an SMC controller constituting the feedback control system.

FIG. 10 is a flowchart showing a flow of feedback control of the turbine rotation speed in the embodiment.

[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 101 SMC controller 102 Drive system 103 Hydraulic system 104 Minimum dimension observer G1, G2 Planetary gear H / C High clutch R / C Reverse clutch L / C Low clutch 2 & 4 / B 2 Speed / 4 speed band brake L & R / B ... Low & reverse brake

Claims (4)

[Claims] 1. A configuration in which hydraulic pressure supply to a friction engagement element is controlled by a sliding mode control based on a switching function using a state quantity of an automatic transmission as a variable so that a state quantity during gear shifting matches a target value. A shift control device for an automatic transmission in which a boundary layer whose control input is continuously approximated is introduced to a switching surface, wherein the width of the boundary layer is reduced during a predetermined period immediately after the start of feedback control. Transmission control device for an automatic transmission. 2. A shift control device for an automatic transmission according to claim 1, wherein the width of said boundary layer is gradually increased in accordance with an elapsed time from the start of feedback control. 3. Within a predetermined holding time from the start of feedback control, the width of the boundary layer is held at an initial value smaller than a reference value, and after the holding time has passed, the reference value is increased by a predetermined increase control time. 3. The shift control device for an automatic transmission according to claim 2, wherein the speed is gradually increased to: 4. The system according to claim 1, wherein the hydraulic pressure supplied to the friction engagement element is feedback-controlled so that the input shaft rotation speed matches the target rotation speed in the inertia phase during the shift. A shift control device for an automatic transmission according to any one of the above.