JPH08258595A - Control device for continuously variable transmission and internal combustion engine - Google Patents

Abstract

PURPOSE: To eliminate the lowering of combustion efficiency and the increase of harmful components in exhaust gas and prevent the wear of a pulley and a driving belt caused by the slip of a CVT belt while realizing excellent accelerating sensation. CONSTITUTION: A cylinder direct injection engine 9 is used, and a delay characteristic generating means is provided so as to have a delay characteristic with a time constant, larger than the response time constant of second oil pressure Po, between the fuel injection quantity target value and the fuel injection quantity command value for determining the actual fuel injection quantity when the fuel injection quantity target value Fc to a fuel injection device 9a changes. Accordingly, even in the case of a driver suddenly depressing an accelerator pedal 51 to cause the sudden change of the fuel injection quantity target value, the change of the fuel injection quantity to the engine is suppressed so as to prevent engine torque Te from exceeding CRT transmittable torque.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a continuously variable transmission (hereinafter referred to as "CVT") for an automobile and an internal combustion engine (hereinafter referred to as "CVT").
(Referred to as "engine"), and particularly to a control device for a CVT and an engine that can control the CVT and the engine in an integrated manner to effectively control the hydraulic pressure for the CVT without degrading combustion efficiency and exhaust gas components. It is about.

[0002]

2. Description of the Related Art Generally, in a CVT in which a gear ratio is continuously controlled via two movable pulleys, a request is made when the input torque changes so that the transmissible torque does not become smaller than the input torque to the CVT. By adjusting the operating hydraulic pressures on the primary side (input side) and secondary side (output side) according to the gear ratio, the transmissible torque is controlled so that the drive belt between the pulleys does not cause slip wear. There is.

However, when the output torque of the engine suddenly changes, the input torque to the CVT may temporarily exceed the transmissible torque of the CVT. In order to avoid such problems, a method of correcting the engine output so that the engine output does not exceed the transmissible torque of the CVT, as in the control method disclosed in Japanese Patent Laid-Open No. 4-50440, for example. Is proposed.

FIG. 17 is a block diagram schematically showing a conventional CVT and engine control device as referred to in the above publication. In the figure, 8 is a starting clutch for transmitting the engine torque Te as CVT input torque Tc, and 9 is a starting clutch.
Is an engine having a plurality of cylinders and outputting an engine torque Te.

Reference numeral 51 denotes an accelerator pedal which outputs an accelerator depression amount α by a driver's operation, and 52 denotes an accelerator depression amount α.
A throttle that adjusts the intake air amount Q to the engine 9 by setting the opening θ according to the above, 55 is a CVT that transmits the CVT input torque Tc to the wheels as the final output torque To, and 56 is the detected throttle opening. The CVT controller outputs a hydraulic valve drive signal DV to the CVT 55 based on the degree θ and the engine speed Ne of the engine 9.

Reference numeral 57 is an ECU (electronic control unit) consisting of a microcomputer for controlling the parameters (ignition timing, fuel injection, etc.) of the engine 9, and includes the following elements 57a to 57d.

57a is an estimated value Ts of the engine torque Te.
Is an engine map for outputting the engine torque estimated value Ts based on the detected engine speed Ne and the detected intake air amount Q. 57b is a CVT map which outputs the transmissible torque Tm by the CVT 55, and the transmissible torque Tm is map-calculated based on the measured hydraulic pressure Po of the CVT 55.

Reference numeral 57c is an engine corrector for outputting a correction signal Gc for the parameters of the engine 9 (ignition timing, air-fuel ratio, etc.), and a torque deviation ΔT between the estimated engine torque value Ts and the transmittable torque Tm (the slip of the CVT 55 is The correction signal Gc is output so that the engine torque Te becomes optimal based on the torque margin until it is generated. 57d is an engine controller that outputs a control command G for parameters of the engine 9 (ignition timing, fuel injection, etc.), and controls the engine torque Te based on the detected intake air amount Qo and the correction signal Gc.

FIG. 18 is a block diagram showing a general automobile torque transmission system including the CVT 55 shown in FIG.
9, 55, 56, DV, Po, Te, Tc and To are the same as described above. Reference numeral 10 denotes a hydraulic pressure sensor (hydraulic pressure measuring means) arranged in a secondary hydraulic pressure chamber (described later) of the CVT 55 to measure the hydraulic pressure Po. A similar hydraulic sensor (not shown) may be arranged in the primary hydraulic chamber (described later).

The CVT 55 has a speed change mechanism 1 associated with a starting clutch 8. The speed change mechanism 1 has a primary pulley 2 and a secondary pulley 3 each of which has one end axially movable, and between the pulleys 2 and 3. And a drive belt 1a that is stretched over. 2a and 3a are fixed pieces of the pulleys 2 and 3, 2b and 3b are movable pieces of the pulleys 2 and 3,
2c and 3c are hydraulic chambers provided in the movable pieces 2b and 3b and filled with hydraulic oil. Arrows A and B are
It is the moving direction of each movable piece 2b and 3b.

The fixed piece 2a of the primary side pulley 2 is connected to a starting clutch 8, and the starting clutch 8 interrupts an engine torque Te output from an engine 9 at the time of starting and stopping, thereby providing an input torque Tc. Fixed piece 2a
Give to. The fixed piece 3a of the secondary pulley drives the vehicle body (not shown) via various gears and shafts (not shown) by the final output torque To.

Further, the CVT 55 has a primary side valve 4 and a secondary side valve 6 which are controlled by respective drive signals DV1 and DV2 from the CVT controller 56, and the respective valves 4 and 6 are respectively the primary side oil passages. 5 and the secondary side oil passage 7, the movable piece 2b of each pulley 2 and 3
And hydraulic oil are supplied to 3b.

The primary-side valve 4 and the primary-side oil passage 5 apply the primary-side hydraulic pressure P1 for changing the gear ratio to the primary-side pulley 2
The secondary side valve 6 and the secondary side oil passage 7 provide a secondary side hydraulic pressure P2 to the secondary side pulley 3 for preventing the drive belt 1a from slipping. Constitutes a hydraulic means.

That is, the secondary side valve 6 is driven by the second hydraulic valve drive signal DV2, applies the secondary side hydraulic pressure P2 to the hydraulic chamber 3c via the secondary side oil passage 7, and the secondary side pulley 3 The movable piece 3b is adjusted in pressure and moved in the direction of arrow B (axial direction). The primary-side valve 4 is driven by the first hydraulic valve drive signal DV1 to further regulate the secondary-side oil pressure P2 diverted from the secondary-side oil passage 7 to generate the primary-side oil passage 5
The primary-side hydraulic pressure P1 is applied to the hydraulic chamber 2c via the, and the movable piece 2b of the primary-side pulley 2 is regulated to move in the arrow A direction (axial direction).

As a result, the contact position between the pulleys 2 and 3 and the drive belt 1a changes in the radial direction of the rotary shaft, and the CV
The gear ratio of T55 is changed. At this time, the secondary hydraulic pressure P
2 is controlled by the second hydraulic valve drive signal DV2 so that the drive belt 1a does not always slip as a line pressure. Further, the primary hydraulic pressure P1 is controlled by the first hydraulic valve drive signal DV1 so that the gear ratio is adjusted as required.

The secondary side oil pressure P2 is detected by the oil pressure sensor 10
Is measured as the hydraulic pressure Po by the CVT in the ECU 57.
It is input to the map 57b and used as a parameter correction signal for the engine 9. As a result, the ECU controls the parameters such as the ignition timing of the engine 9 and the air-fuel ratio (fuel injection) to be optimum in terms of combustion efficiency and exhaust gas components.
It is controlled by the engine controller 57d in 57.

By the way, in the control device disclosed in the above-mentioned Japanese Patent Application Laid-Open No. 4-50440, an engine of the type that injects fuel into the intake port is used as the engine 9, and this type of engine 9 Intake air amount Qo
The engine torque Te is determined using as a main parameter. Further, control parameters such as the ignition timing of the engine 9 and the air-fuel ratio (ratio between the fuel injection amount and the intake air amount) are controlled by the engine controller 57d to be optimum in terms of combustion efficiency and harmful components in exhaust gas. To be done.

Therefore, when the ignition timing and the air-fuel ratio are changed from the optimum state, the engine torque Te changes a little, but the combustion efficiency decreases and the harmful components in the exhaust gas increase. The intake air amount Q with respect to the engine 9 changes with a characteristic that is substantially linearly delayed with respect to the change in the throttle opening θ.

Further, since the fuel injected into the engine 9 once adheres to the intake port of each cylinder, it has a property of substantially first-order lag until it is actually vaporized and becomes combustible. Further, the engine torque Te changes with a first-order lag characteristic in which a time constant is a sum of both time constants with respect to a change in the throttle opening θ due to the intake air amount Q and a delay in fuel vaporization.

The control operation of the conventional CVT 55 and the engine 9 (in particular, the first-order lag characteristic of the engine torque Te) will be specifically described below with reference to FIGS. 17 and 18. First, when the driver of the vehicle operates the accelerator pedal 51, the throttle 52 is driven in the opening direction according to the accelerator depression amount α, and the intake air amount Q corresponding to the throttle opening θ is supplied to the engine 9.

At this time, the engine torque Te generated from the engine 9 is determined by the intake air amount Q passing through the throttle 52 and the rotation speed Ne of the engine 9. The engine torque Te is directly input to the CVT 55 when the output shaft of the engine 9 is in a state where the starting clutch 8 is directly connected (other than starting and stopping), and the engine torque Te is input to the CVT 55.
Is multiplied by the gear ratio, and finally becomes the output torque To that drives the vehicle body.

On the other hand, the CVT controller 56 outputs a hydraulic valve drive signal DV based on the detected throttle opening θ and the engine speed Ne to determine the primary side hydraulic pressure P1 and the secondary side hydraulic pressure P2 for the CVT 55.
The hydraulic pressures P1 and P2 are generated by the valves 4 and 6 in the CVT 55 driven by the hydraulic valve drive signal DV, as described above.

At this time, the drive belt 1a in the CVT 55
In order to prevent the occurrence of slippage, the hydraulic pressure Po (each hydraulic pressure P1 and P2) acting on the secondary pulley 3 (and the primary pulley 2) is measured by the hydraulic pressure sensor 10 or the like, and the ECU 57
It is fed back to the CVT map 57b inside.

The CVT map 57b shows the measured oil pressure P.
A transmissible torque Tm is calculated based on o. Also,
The engine map 57a is based on the engine speed Ne of the engine 9 and the intake air amount Qo, and the estimated engine torque T
Calculate s. Further, the engine corrector 57c uses the correction signal G to optimize the engine torque Te based on the torque margin ΔT (= Ts−Tm), which is the torque deviation between the estimated engine torque value Ts and the transmittable torque Tm.
Output c.

That is, when the torque margin ΔT of the CVT 55 is below a predetermined value, the correction signal Gc for the target air-fuel ratio and the ignition timing is generated, and when the torque margin ΔT is above the predetermined value, the combustion of the engine 9 is performed. It is input to the engine controller 57d that optimally controls the efficiency and the like, whereby the target air-fuel ratio and the ignition timing are corrected to correct the engine torque Te.

By the way, the oil pressure sensor 10 in the CVT 55
High-frequency vibration and noise are superposed on the hydraulic pressure Po measured by (1) due to vibration of a pump (not shown) that is a hydraulic pressure generation source, vibration of piping including the oil passages 5 and 7, and the like.

Therefore, as a means for preventing superposition of high frequency vibration and noise on the measurement signal of the hydraulic pressure Po, a method of passing the measurement signal of the hydraulic pressure Po through a filter has been conventionally adopted. However, the filter output has a property that it lags behind the actual behavior, and especially when dealing with a short-term phenomenon of a transient response of the engine torque Te and the hydraulic pressure Po, the delay due to the filter cannot be ignored.

In particular, the CVT as referred to in the above publication
In the control device for the engine 55 and the engine 9, since the engine is not a direct injection type cylinder engine, there is a control delay due to the time required for fuel vaporization. Even if an in-cylinder direct injection engine is used, in-cylinder direct injection control is not taken into consideration, so the combustion efficiency is reduced and the harmful components in the exhaust gas are increased with the correction of the ignition timing. Or may cause vibration of the engine torque Te.

The hydraulic pressures P1 and P2 of the CVT 55 will be described below.
And a phenomenon related to the slip of the drive belt 1a will be specifically described. The line pressure for preventing the drive belt 1a from slipping may be applied to either the primary side pulley 2 or the secondary side pulley 3, but as shown in FIG. It shall be given as the secondary side hydraulic pressure P2.

Usually, when the gear ratio of the CVT 55 is controlled to be constant, or the movable pieces 2b and 2 of each pulley are
When the moving speed of c is sufficiently small, the respective hydraulic pressures P1 and P2
The response (change with time) of is the following differential equation (1),
It is represented by (2).

DP1 / dt = B (Pv1-P1) / (V1 · R1) (1) dP2 / dt = B (Pv2-P2) / (V2 · R2) (2)

However, in the equations (1) and (2), V
1 and V2 are the volumes of the hydraulic chambers 2c and 3c, B is the bulk modulus of the hydraulic oil, R1 and R2 are the piping resistances of the oil passages 5 and 7, Pv1 and Pv2 are the outlet pressures of the valves 4 and 6 (= Line pressure).

From the equations (1) and (2), the primary side hydraulic pressure P1 and the secondary side hydraulic pressure P2 show a primary delay response to changes in the outlet hydraulic pressures Pv1 and Pv2 of the respective valves. That is, the response time constant of the primary delay is respectively the primary side hydraulic pressure P
It is represented by V1 · R1 / B for 1 and V2 × R2 / B for the secondary hydraulic pressure P2.

Further, since the pipe resistances R1 and R2 of the oil passages 5 and 7 change depending on the temperature of the hydraulic oil (oil temperature), the time constant of the hydraulic response changes depending on the oil temperature. In addition, each hydraulic pressure P related to each outlet hydraulic pressure Pv1 and Pv2
In 1 and P2, the opening degrees of the valves 4 and 6 (that is, corresponding to the rising time with respect to the target hydraulic pressure value) generally change rapidly, so that they show approximately a first-order lag response to the hydraulic pressure target value. . Furthermore, as described above, actually, high-frequency vibrations such as vibrations of the pump, which is a hydraulic pressure generation source, and vibrations of piping are superimposed.

[0035]

As described above, the conventional CVT and engine control device uses the engine torque T based on parameters such as the ignition timing of the engine 9 and the air-fuel ratio.
Since e is corrected, there is a possibility that combustion failure will occur, especially when the throttle opening θ changes significantly, and combustion efficiency will temporarily decline and harmful components in the exhaust gas will increase.

Further, since the correction signal Gc is determined based on the oil pressure Po measured by the oil pressure sensor 10, the engine torque Te is affected by the high frequency signal and noise superimposed on the measurement signal of the oil pressure Po. Therefore, there is a problem that in addition to the reduction of combustion efficiency and the increase of harmful components in the exhaust gas, the vibration of the engine torque Te is caused.

Further, as a means for preventing superposition of high frequency vibration and noise on the measurement signal of the hydraulic pressure Po, the hydraulic pressure Po is used.
When the method of passing the measurement signal of (3) through the filter is adopted, there is a problem that the response of the control is deteriorated because the output of the filter is delayed with respect to the actual behavior.

The present invention has been made to solve the above-mentioned problems, and wear due to slippage of the drive belt in the CVT without a reduction in combustion efficiency and an increase in harmful components in exhaust gas. It is an object of the present invention to obtain a control device for a CVT (continuously variable transmission) and an engine (internal combustion engine) that prevents the above.

[0039]

A control device for a CVT and an engine according to claim 1 of the present invention uses an in-cylinder direct injection type engine and is provided with delay characteristic generating means.
The delay characteristic generating means, when the fuel injection amount target value for the fuel injection device changes, between the fuel injection amount target value and the fuel injection amount command value that defines the actual fuel injection amount, the second hydraulic pressure response. It has a delay characteristic of a time constant larger than the time constant.

Further, a control device for a CVT and an engine according to a second aspect of the present invention is, in the first aspect, a steady torque calculating means for calculating a steady torque of the engine from a fuel injection amount and an engine speed, and a steady torque calculation means. A transient torque estimating means for estimating a transient response value of the engine torque from the calculated value, a hydraulic pressure measuring means for measuring a second hydraulic pressure, and a CV
A gear ratio detecting means for detecting a gear ratio of T, a transmissible torque computing means for computing a transmissible torque of CVT from the second hydraulic pressure and the gear ratio, an estimated value of a transient response value and a computed value of the transmissible torque. And a torque margin calculating means for obtaining a comparison calculation value of 1., the delay characteristic generating means is enabled only when the comparison calculation value becomes equal to or less than a predetermined value.

According to a third aspect of the present invention, in the CVT and engine control device according to the second aspect, the time constant of the delay characteristic is set as a function of the comparison calculation value.

Further, a control device for a CVT and an engine according to a fourth aspect of the present invention is provided with a hydraulic pressure response estimating means (hydraulic response model) according to the second aspect, and the hydraulic response estimating means is a hydraulic pressure for the second hydraulic pressure. Second based on the target value
The hydraulic pressure response estimated value having a first-order lag characteristic having the hydraulic pressure response time constant as the time constant is obtained, and the transmittable torque calculation means uses the hydraulic response estimated value as the second hydraulic pressure.

Further, a control device for a CVT and an engine according to a fifth aspect of the present invention, in the second aspect, is provided with a two-input type hydraulic pressure estimating means (hydraulic pressure estimating observer), and the hydraulic pressure estimating means is a second hydraulic pressure. Is multiplied by a constant and a hydraulic pressure target value to obtain an estimated hydraulic pressure value having a first-order lag characteristic whose time constant is the sum of the response time constant of the second hydraulic pressure and the constant. Indicates that the estimated hydraulic pressure is used as the second hydraulic pressure.

A control device for a CVT and an engine according to claim 6 of the present invention is the oil temperature detection device for detecting the oil temperature of the hydraulic oil supplied to the CVT according to any one of claims 1 to 5. Means and a water temperature detecting means for detecting the cooling water temperature of the engine are provided, and the delay characteristic generating means changes the time constant of the delay characteristic based on at least one of the oil temperature and the cooling water temperature. .

According to a seventh aspect of the present invention, a control device for a CVT and an engine according to any one of the first to fifth aspects, wherein the engine speed is integrated from the time when the engine start command is issued. Means are provided, and the delay characteristic generating means changes the time constant of the delay characteristic based on the integrated value of the number of revolutions.

[0046]

According to the first aspect of the present invention, when the fuel injection amount target value changes, the second CVT is set between the fuel injection amount target value and the fuel injection amount command value for determining the actual fuel injection amount. The engine output torque (that is, the input torque to the CVT) exceeds the CVT transmissible torque even when the target value of the fuel injection amount suddenly changes by providing the delay characteristic of the time constant larger than the response time constant of the hydraulic pressure of. Wear of the pulley and belt due to CVT belt slip is prevented. In addition, fuel such as gasoline, methanol, and diesel is injected directly into the cylinder in a combustible state, and there is no delay in fuel vaporization, so transitional engine torque rises quickly and unvaporized fuel exists. Therefore, unless the ignition timing is changed from the optimum state, the combustion efficiency and the optimality of harmful components in the exhaust gas are maintained.

Further, according to a second aspect of the present invention, the steady torque of the engine is calculated from the fuel injection amount and the engine speed, and the transient response value of the engine torque is estimated from the calculated value of the steady torque. Of the CVT is calculated from the second hydraulic pressure and the gear ratio, and a comparative calculated value (torque margin) between the estimated value of the transient response value and the calculated value of the transmissible torque is predetermined. The delay characteristic generating means is enabled only when the value becomes less than or equal to the value, so that the response of the engine torque is suppressed only when there is a risk that the CVT belt slips, and the response speed of the engine torque is reduced more than necessary. To prevent

According to the third aspect of the present invention, the time constant of the delay characteristic is set as a function of the comparison calculation value (torque margin) to suppress the response of the engine torque according to the torque margin and to reduce the torque margin. Specify the optimum engine output torque rise characteristics according to the.

In the fourth aspect of the present invention, C
Based on the hydraulic pressure target value for the second hydraulic pressure of VT, the second
Of the hydraulic pressure response time constant of the CVT is used as the second hydraulic pressure, and the hydraulic response estimated value is used as the second hydraulic pressure. CV is calculated from the estimated hydraulic pressure value of the delay characteristic and the detected gear ratio value.
By obtaining the transmissible torque of T, it is possible to detect a hydraulic pressure value on which high frequency vibration and noise are not superimposed without using a filter. Further, since it is not necessary to pass the oil pressure measurement signal through the filter, there is no delay with respect to the actual behavior, and in particular, the control cannot be ignored when the torque margin for the CVT belt slip is obtained from the transient response between the engine torque and the oil pressure. Prevent delays.

In the fifth aspect of the present invention, C
Based on the value obtained by multiplying the second hydraulic pressure of VT by a constant and the hydraulic pressure target value, an estimated hydraulic pressure value having a first-order lag characteristic whose time constant is the sum of the response time constant of the second hydraulic pressure and the constant is obtained. By calculating the transmittable torque using the estimated hydraulic pressure value as the second hydraulic pressure, an intermediate detected value between the estimated hydraulic response value and the measured hydraulic pressure value according to claim 4 is obtained, and the characteristic of the detection signal is set to the frequency. The hydraulic pressure value is detected in the optimum condition according to the actual signal condition.

In the sixth aspect of the present invention, attention is paid to the fact that the time constant of the hydraulic response changes depending on the oil temperature,
The time constant of the delay characteristic is changed based on at least one of the oil temperature of the hydraulic oil supplied to the CVT and the cooling water temperature of the engine. Thereby, the response characteristic of the engine is changed according to the change of the hydraulic response characteristic due to the actual change of the oil temperature of the CVT, and the controllability is improved.

Further, in claim 7 of the present invention, paying attention to the fact that the oil temperature rises according to the integrated value of the engine speed, the engine speed from the time when the engine start command is issued is integrated, and the integrated value of the engine speed is obtained. Based on this, the time constant of the delay characteristic is changed. Thus, the approximate value of the oil temperature is detected without using the temperature sensor, and the response characteristic of the engine is changed according to the change of the hydraulic response characteristic.

[0053]

【Example】

Example 1. Embodiment 1 of the present invention will be described below with reference to the drawings. 1 is a block diagram showing a first embodiment of the present invention, and FIG. 2 is a configuration diagram showing a torque transmission system including a CVT 55 and an engine 9 in FIG.
51, 52, 55 and 57d are the same as described above. 57A corresponds to the ECU 57, and 57f
Corresponds to the CVT controller 56.

The ECU 57A is the engine controller 5
In addition to the 7d and the CVT controller 57f, a fuel injection amount corrector 57e that corrects a command from the engine controller 57d and outputs a fuel injection signal J is provided.

Reference numeral 9a denotes a fuel injection device (injector) for supplying the fuel F into the engine 9 in response to the fuel injection output J.
Is. In this case, an in-cylinder direct injection type engine is used as the engine 9, and the fuel F is directly injected in a combustible state by the fuel injection device 9a arranged in the cylinder.

Hydraulic chamber 3c of the secondary pulley 3 (see FIG. 2)
The secondary side oil pressure P2 in the inside is measured by the oil pressure sensor 10.
CVT controller 5 in the ECU 57
It is taken into 7f and used as a feedback signal when controlling the secondary hydraulic pressure P2. In addition, the accelerator depression amount α
Detected value αo of the engine controller 57d and C
It is taken into the VT controller 57f and used to control the engine 9 and the CVT 55.

Similarly to the above, the engine controller 57d controls the ignition timing and the air-fuel ratio so as to satisfy the conditions of improving the combustion efficiency and suppressing the harmful components in the exhaust gas.

Further, the fuel injection amount corrector 57e includes a delay characteristic generating means, as will be described later, and includes the fuel injection device 9a.
When the target fuel injection amount target value Fc changes with respect to the fuel injection amount target value Fc between the fuel injection amount target value Fc and the fuel injection amount command value Fu (fuel injection signal J) that determines the actual fuel injection amount. It has a delay characteristic with a larger time constant.

Here, one ECU 57A is used as the engine controller 57d and the CVT controller 5.
Although the configuration including 7f is adopted, each controller function may be configured by an individual ECU so that data signals and control signals are exchanged between the ECUs.

Next, the operation of the first embodiment of the present invention shown in FIGS. 1 and 2 will be described. First, when the driver operates the accelerator pedal 51, the intake air amount Q corresponding to the accelerator depression amount α is supplied to the engine 9, and a desired rotation speed Ne is obtained. On the other hand, the engine controller 57d in the ECU 57A calculates the target fuel injection amount Fc and the ignition timing according to the detected value αo of the accelerator depression amount α and the detected value Qo of the intake air amount Q, and outputs the ignition timing signal G. It is output as a parameter command for the engine 9.

Further, the fuel injection amount corrector 57e corrects the fuel injection amount (target value) Fc calculated by the engine controller 57d, and outputs the fuel injection signal J as the final fuel injection amount command value. 9a. At this time, the fuel injection amount corrector 57e is provided with a first-order delay characteristic having a time constant larger than the hydraulic response of the CVT 55.

As a result, the fuel injection device 9a injects fuel into the engine 9 based on the fuel injection signal J, and the engine 9 outputs the engine torque Te determined by the fuel injection amount and the rotation speed Ne. Engine torque Te
As described above, the CVT 55
Input torque Tc, and finally output torque To via CVT 55, and a vehicle body (not shown) is driven via gears and the like.

Next, the operation of the first embodiment of the present invention will be specifically described with reference to the flowchart of FIG. FIG. 3 shows a fuel control processing routine in the ECU 57A. First, when the ignition signal is turned on by the driver's operation, in step S101, an initial value Fu (0) of the fuel injection amount command value is read from a ROM (not shown) in the ECU 57A, and a memory including a RAM or the like is read. To memorize. The initial value F of the fuel injection amount command value
u (0) is preliminarily stored in the ROM in the ECU 57A.

Further, in step S102, the fuel injection amount target value F generated by the engine controller 57d.
c (k) is read and stored in the memory in the same manner. Then, in step S103, the fuel injection amount target value Fc
(K) and the previous value of the fuel injection amount command value Fu (k) (k
In the case of = 0, based on the initial value Fu (0)), the current fuel injection amount command value Fu (k +) is calculated by the following equation (3).
Calculate 1) and store this in memory.

Fu (k + 1) = a1 · Fu (k) + b1 · Fc (k) (3)

In the equation (3), the first loop (k =
When calculating the fuel injection amount command value Fu (1) in 0),
The previous value becomes the initial value Fu (0).

Further, a1 and b1 in the equation (3) are coefficients for the filtering calculation, and give the time constant of the first-order delay to the engine 9 which responds too early. Each coefficient a1 and b1 is uniquely and discretely determined based on the time constants V1.R1 / B and V2.R2 / B in the above equations (1) and (2) and the sampling interval. It is a value and satisfies the relationship of a1 + b1 = 1.

Then, in step S104, the drive pulse length Fp for the fuel injection device 9a is determined based on the fuel injection amount command value Fu (k + 1). Finally, in step S105, the engine controller 57d
In accordance with the pulse output command from, the drive pulse for the fuel injection device 9a, that is, the fuel injection signal J is output.

Thereafter, the process returns to step S102 for reading the fuel injection amount target value Fc (k), and steps S102 to S10.
The processing loop up to 5 is repeated. After the second loop, in step S103, the fuel injection amount command value Fu (k +
When calculating 1), the previous value is sequentially updated from the initial value Fu (0) to the fuel injection amount command value Fu (k) at the time of the previous sample, and the processing is executed.

Next, the torque response operation according to the first embodiment of the present invention will be described with reference to the characteristic diagram of FIG. 4 showing the relationship between the accelerator depression amount α and the engine torque Te and the CVT transmissible torque.

Now, assuming that the accelerator depression amount α changes rapidly at time to as shown in FIG. 4A,
At this time, assuming that the fuel injection amount command value Fu is set as it is according to the fuel injection amount target value Fc, FIG.
As shown in (b), the uncorrected engine torque Teo (dotted line) temporarily exceeds the transmissible torque Tm (solid line) of the CVT 55, and the drive belt 1a (see FIG. 2) slips. Become.

Therefore, as in step S103,
By providing a delay characteristic having a time constant larger than the response time constant of the hydraulic pressure Po of the CVT 55 between the fuel injection amount target value Fc and the fuel injection amount command value Fu (fuel injection signal J), the corrected engine The torque Tec (broken line) becomes a response curve subjected to the first-order lag processing.

Therefore, the engine torque Te is CVT.
It does not exceed the transmissible torque Tm of 55, and it is possible to prevent wear of the pulleys 2 and 3 and the drive belt 1a due to the slip of the drive belt 1a.

In particular, as shown in FIG. 2, in the engine 9 having the in-cylinder direct injection type fuel injector 9a, fuels such as gasoline, methanol and diesel are directly injected into the cylinder in a combustible state. There is no delay in fuel vaporization,
The transitional rise is fast. There is also an engine 9 of the type in which the throttle 52 is excluded, and in this case, the rise of the intake air amount Q becomes faster, so that the delay of the injection amount of the fuel F with respect to the operation of the accelerator pedal 51 becomes further smaller.

However, the fuel injection control can be optimized by the response delay processing step S103. Further, in the engine 9, the state in which the fuel F is not vaporized disappears, so that the optimality in terms of improvement of combustion efficiency and suppression of harmful components in exhaust gas is maintained unless the ignition timing is changed from the optimal state.

Example 2. In the first embodiment, the CV
Although the transmissible torque Tm of T55 was not calculated, the transmissible torque Tm based on the gear ratio of CVT55 is used.
To calculate the estimated engine torque and the transmittable torque Tm
The fuel injection signal J may be corrected based on the deviation (torque margin) from

For example, in the first embodiment, the change in the fuel injection amount is always suppressed by the first-order lag characteristic, but the torque margin ΔT may be larger than the predetermined value even when the accelerator depression amount αo becomes large. In order to suppress the change in the fuel injection amount, the accelerator pedal 51 of the engine 9 is used.
May deteriorate the responsiveness to.

A second embodiment of the present invention in which the fuel injection signal J is corrected based on the torque margin to prevent deterioration of responsiveness of the accelerator pedal 51 will be described below with reference to the drawings. FIG. 5 is a block diagram showing a second embodiment of the present invention.
a, 51, 52, 55 and 57d to 57f are the same as described above. Further, 57B corresponds to the ECU 57A, and the configuration of the torque transmission system including the CVT 55 is as shown in FIG.

The ECU 57B is the engine controller 5
7d, a fuel injection amount corrector 57e, and a CVT controller 57f, the following elements 57g to 57j are provided. 57 g is the rotation speed N1 of the primary pulley 2 of the CVT 55.
And a gear ratio calculator that calculates the gear ratio RT based on the rotational speed N2 of the secondary pulley 3.

57h is an engine torque estimator which calculates an estimated value Tes of the transient response value of the engine torque Te based on the accelerator depression amount Qo, and 57i is a comparative calculated value of the estimated value Tes and the transmissible torque Tm (in this case, the torque). Deviation) is calculated as a torque margin ΔT (= Tm-Tes), and a torque margin calculator 57j is a gear ratio RT of the CVT 55.
And a transmissible torque calculating means for calculating a transmissible torque Tm based on the hydraulic pressure Po. The torque margin ΔT is input to the fuel injection amount corrector 57e.

The ECU 57B in FIG. 5 is the ECU in FIG.
A gear ratio calculator 57g to a transmittable torque calculator 57j is added to 57A to provide a torque margin ΔT (comparative calculation value).
Is to ask for. In this case, the fuel injection amount corrector 57e enables the delay characteristic generating means only when the torque margin ΔT becomes equal to or less than the predetermined value.

Next, the operation of the second embodiment of the present invention shown in FIG. 5 will be described with reference to FIG. First, after the accelerator pedal 51 is depressed, the engine torque T
The operation until e is output is the same as described above.

Engine torque estimator 5 in the ECU 57B
7h is the engine torque T based on the detected accelerator depression amount αo and engine speed Ne using a map or the like.
A steady value of e is obtained, and this is input to a first-order lag system having a time constant estimated by experiments or the like to obtain an estimated value Tes of the transient response value of the engine torque Te.

In addition, the gear ratio calculator 57 in the ECU 57B
g is a measurement signal of the secondary hydraulic pressure P2 of the CVT 55 (hydraulic pressure P
o) and the ratio of the rotational speeds N1 and N2 of the primary side pulley 2 and the secondary side pulley 3 to obtain the speed change ratio RT, and the transferable torque calculating means 57j can transfer the CVT 55 from the speed change ratio RT and the hydraulic pressure Po. The torque Tm is calculated.

Further, the torque margin estimator 57i obtains a comparative calculation value (torque deviation) as a torque margin ΔT from the estimated value Tes of the transient response value of the engine 9 and the transmissible torque Tm of the CVT 55, and determines this as the fuel margin ΔT. Corrector 57
Enter in e.

The fuel injection amount corrector 57e has a first-order lag characteristic having a time constant larger than the response of the hydraulic pressure Po of the CVT 55 only when the accelerator depression amount αo becomes large and the torque margin ΔT becomes a predetermined value or less. Is given. The predetermined value that serves as a comparison standard for the torque margin ΔT is
It should be set to the lowest value that ensures safety.

Next, referring to the flowcharts of FIGS. 6 and 7, the ECU 57B according to the second embodiment of the present invention.
The internal processing operation will be described. First, when the ignition signal is turned on by the driver's operation, the estimated value Te of the engine torque Te is calculated in step S201.
The initial value Teu (0) of s is stored. Further, similar to step S101 described above, in step S202, the initial value Fu (0) of the fuel injection amount command value is stored in the memory.

Subsequently, in step S203, the detected engine speed Ne is read and stored in the memory, and similarly, in step S204, the accelerator depression amount αo is read and stored in the memory. Next, in step S205, a steady value of the engine torque Te is calculated from the engine map as an estimated value Tes (k) for the engine speed Ne and the accelerator depression amount αo, and this is stored in the memory.

Further, in step S206, based on the estimated value Teu (k) calculated in the previous sampling and the estimated value Tes (k) of the steady value, the current estimation of the engine torque Te is calculated from the following equation (4). Value Tes (k +
1) is found and stored in memory.

Tes (k + 1) = a2 · Tes (k) + b2 · Tes (k) (4)

However, in the formula (4), a2 and b2
Is a coefficient used in the filter calculation of the first-order lag characteristic, and satisfies the relationship of a2 + b2 = 1. Further, in the equation (4), at the time of the first sampling (k = 0), the initial value Teu is set as the previous estimated value Teu (k).
(0) is used.

Next, in step S207, the calculated output of the gear ratio calculator 57g, that is, the gear ratio RT is read and stored in the memory, and in step S208, the measured value of the secondary oil pressure P2, that is, the oil pressure Po, is read and stored in the memory. To do. In step S209, the gear ratio R
Transferable torque Tm of CVT 55 from T and hydraulic pressure Po
Is calculated and stored in memory.

Subsequently, in step S210, the estimated value Tes is subtracted from the transmittable torque Tm to obtain the torque margin Δ.
Calculate T and store it in memory. Next, in step S211, it is determined whether or not the torque margin ΔT is equal to or greater than a predetermined value (for example, preset in ROM or the like).

If it is determined that the torque margin ΔT is equal to or greater than the predetermined value (that is, YES), the fuel injection amount target value Fc (k) generated by the engine controller 57d is set to the command value Fu (k + 1) in step S212. And store it in the memory. In addition, step S21
In 3, the fuel injection device drive pulse length Fp is determined based on the command value Fu (k + 1) and stored in the memory.

Next, in step S214, a pulse output command (target value F from the engine controller 57d is generated.
The drive pulse (fuel injection signal) J to the fuel injection device 9 is output according to (corresponding to c). Below, step S
Returning to step 203, the next sampling process is repeated.

On the other hand, if it is determined in step S211 that the torque margin ΔT is less than the predetermined value (that is, NO), as in steps S102 and S103 described above,
The process proceeds to the process of providing a first-order lag characteristic of a time constant larger than the response time constant of the hydraulic pressure Po between the fuel injection amount target value Fc and the previous fuel injection amount command value Fu.

That is, first, in step S215, the fuel injection command value Fu (k +
1) is read and stored in the memory as the previous value Fu (k). Succeedingly, in a step S216, the fuel injection amount target value Fc generated by the engine controller 57d is read and stored in the memory, and in a step S217, the first-order lag characteristic is calculated by the above-described filter calculation formula (3).

After that, the routine proceeds to step S213, and the same processing as that when the torque margin ΔT is equal to or larger than a predetermined value is performed. Hereinafter, the process returns to step S203 to start the next sampling process. Although the torque margin ΔT is obtained by subtracting the estimated value Tes from the transmissible torque Tm here, the torque margin ΔT may be obtained by dividing the transmissible torque Tm by the estimated value Tes.

As described above, by adopting the configuration shown in FIG. 5, the change in the engine torque Te can be suppressed only when there is a risk that the drive belt 1a of the CVT 55 slips, and the engine output Te is unnecessarily increased. It is possible to prevent the situation in which the response speed is reduced to deteriorate the responsiveness.

Example 3. Although the time constant of the first-order lag characteristic of the fuel injection amount corrector 57e is constant in the second embodiment, the time constant may be changed according to the torque margin ΔT. A third embodiment of the present invention in which the time constant of the first-order lag characteristic is changed will be described below with reference to the drawings.

In this case, in the fuel injection amount corrector 57e, the time constant of the first-order lag characteristic changes according to the torque margin ΔT calculated by the torque margin estimator 57i. Since it is the same as the second embodiment, FIG.
Will be described with reference to. The time constant changes so that it decreases when the torque margin ΔT is large and increases when the torque margin ΔT is small, but it may be given by a map calculation for the torque margin ΔT or may be given as a function.

Next, the ECU 5 according to the third embodiment of the present invention
The processing operation of 7B (see FIG. 5) will be described with reference to the flowchart of FIG. Note that steps S201 to S2 shown in FIG. 6 and FIG. 7 are performed until the torque margin ΔT is calculated by subtracting the estimated value Tes from the transmissible torque Tm and stored in the memory (step S210).
The description is omitted because it is the same as 10.

Following step S210, first, in step S311, the time constant τc of the first-order lag characteristic of the fuel injection amount corrector 57e is obtained from the map from the value of the torque margin ΔT and stored in the memory. Then, in step S312, the coefficients a1 and b1 in the above equation (3) are obtained from the value of the time constant τc, for example, by a map and stored in the memory.

Hereinafter, steps S102 to S10 in FIG.
After the drive pulse J for the fuel injection device 9a is output by the same process steps S314 to S316 as in step 4, the process returns to the above-mentioned step S203 (see FIG. 6) to start the next sampling process.

Here, after the time constant τc is obtained from the map (step S311), the values of the coefficients a1 and b1 are calculated (step S312).
The values of the coefficients a1 and b1 may be directly obtained from the value of.
Although the torque margin ΔT is calculated by subtracting the estimated value Tes from the transmissible torque Tm, the torque margin ΔT may be calculated by dividing the transmissible torque Tm by the estimated value Tes.

As described above, according to the third embodiment of the present invention, the time constant τ of the first-order lag characteristic of the fuel injection amount corrector 57e.
By changing c with respect to the torque margin ΔT, it is possible to specify the optimum rising characteristic of the engine torque Te according to the torque margin ΔT.

Example 4. In the third embodiment described above, the time constant τc of the first-order lag characteristic in the fuel injection amount corrector 57e.
The transmissible torque Tm is calculated based on the hydraulic pressure Po in order to change the variable torque according to the torque margin ΔT. However, the transmissible torque Tm may be calculated based on an estimated target value of the hydraulic pressure Po.

Next, a fourth embodiment of the present invention in which the transmittable torque Tm is calculated based on the hydraulic response estimated value will be described with reference to the drawings. FIG. 9 is a block diagram showing a fourth embodiment of the present invention, 57C corresponds to the ECU 57B, and other configurations (each indicated by the same reference numeral as above) are the same as those in FIG.

In the ECU 57C in FIG. 9, 57k is a hydraulic response model for generating the estimated hydraulic pressure value Pos from the target value Poc of the hydraulic pressure Po, and constitutes hydraulic pressure response estimation means that cooperates with the CVT controller 57f. ECU 57
C is the ECU 57B in FIG. 5 with a hydraulic response model 57k added, and the hydraulic response model 57k is the hydraulic pressure P.
An estimated hydraulic pressure value Pos having a first-order lag characteristic having a response time constant of o as a time constant is input to the transmittable torque calculation means 57i.

In this case, the CVT is set in the hydraulic response model 57h.
Secondary hydraulic pressure target value Poc calculated by the controller 57f
Is input to estimate the hydraulic pressure, and the estimated hydraulic pressure value Pos is input to the transmissible torque calculating means 57j together with the gear ratio RT. At this time, the time change of the estimated hydraulic pressure value Pos generated from the hydraulic response model 57k is given by the following equation (5).

DPos2 / dt = B (Poc2-Pos2) / (V2 · R2) (5)

However, in the equation (5), Pos2 is the estimated oil pressure on the secondary side, Poc2 is the target hydraulic pressure on the secondary side, B is the bulk elastic coefficient of the hydraulic oil, and V2 is the secondary hydraulic chamber 3c (Fig. Two
R) is a pipe resistance of the secondary oil passage 7.
Since the measured hydraulic pressure Po is the secondary hydraulic pressure P2 here, the secondary hydraulic pressure estimated value Pos2 and the secondary hydraulic pressure target value Poc2 in the equation (5) are the estimated hydraulic hydraulic pressure Po and the target hydraulic pressure Po, respectively. Equivalent to.

Next, the processing operation in the ECU 57C according to the fourth embodiment of the present invention will be described with reference to the flowchart in FIG. In this case, the difference from the above-described second and third embodiments lies only in the hydraulic pressure detection method, so only different processing operations will be described. Therefore, FIG. 10 shows only the estimated hydraulic pressure value calculation operation using the hydraulic response model 57k.

First, in step S401, the initial value Pos (0) of the estimated oil pressure value Pos is read from the ROM and stored in the memory. Similarly, in step S402,
The hydraulic pressure target value Poc is read and stored in the memory. Subsequently, in step S403, the estimated hydraulic pressure value Pos (k) calculated in the previous sampling and the hydraulic target value Poc are calculated.
Based on (k), the current estimated hydraulic pressure value Pos (k + 1) is calculated as in the following equation (6) and stored in the memory.

Pos (k + 1) = a3 · Pos (k) + b3 · Poc (k) (6)

However, in the formula (6), a3 and b3
Is a coefficient used for the filter calculation, and is a3 + b
The relationship of 3 = 1 is satisfied. At the time of the first sampling, the initial value Pos (0) is used as the previous estimated hydraulic pressure value Pos (k). Hereinafter, the process returns to step S402 for reading the hydraulic pressure target value Poc, and the next sampling process is started.

Using the estimated oil pressure value Pos thus obtained and the gear ratio RT obtained from the gear ratio calculator 57g,
The transmittable torque Tm is calculated as described above, and a torque margin is further provided. In addition, here, the CVT controller 5
Although the oil pressure Po measured by the CVT 55 is used as the oil pressure detection value input to 7f, the oil pressure estimated value Pos generated from the oil pressure response model 57k may be used.

As described above, as the hydraulic pressure value (secondary hydraulic pressure value) for calculating the transmittable torque Tm of the CVT 55,
By using the estimated hydraulic pressure value Pos obtained by inputting the desired hydraulic pressure value Poc into the hydraulic response model 57k, the hydraulic pressure value can be detected without superimposing high frequency vibration and noise. Therefore, the vibration of the engine torque Te caused by the disturbance of the fuel injection amount command signal (fuel injection signal J) is not caused.

Example 5. In the fourth embodiment, the hydraulic response model 57k is used to obtain the estimated hydraulic pressure Pos. However, the estimated hydraulic pressure observer is used to determine the estimated hydraulic pressure Pos based on the hydraulic target value Poc and the hydraulic pressure Po. Good. A fifth embodiment of the present invention using a hydraulic pressure estimation observer will be described below with reference to the drawings.

FIG. 11 is a block diagram showing a fifth embodiment of the present invention, 57D corresponds to the above-mentioned ECU 57C, and other configurations (each indicated by the same reference numeral as the above) are the configurations in FIG. Is the same as.

In the ECU 57D in FIG. 11, 57 m
Is the estimated hydraulic pressure value Po from the hydraulic pressure target value Poc and the hydraulic pressure Po.
It is a hydraulic pressure estimation observer that generates s, and constitutes a two-input hydraulic pressure estimation means that cooperates with the CVT controller 57f. The ECU 57D has a hydraulic pressure estimation observer 57m in place of the hydraulic pressure response model 57k in the ECU 57C.
An estimated hydraulic pressure value Po having a first-order lag characteristic whose time constant is the sum of a response time constant of o and a constant (multiplier for hydraulic pressure Po)
s is input to the transmittable torque calculation means 57i.

In this case, the oil pressure estimation observer 57m is
The hydraulic target value Po calculated by the CVT controller 57f
c and the hydraulic pressure Po measured by the hydraulic pressure sensor 10 (see FIG. 2) of the CVT 55, the estimated hydraulic pressure value Poc is obtained,
This can be transmitted together with the gear ratio RT.
Enter in j. At this time, the temporal change of the hydraulic pressure estimated value Pos generated from the hydraulic pressure estimation observer 57m is given by the following equation (7) corresponding to the above equation (5) using the secondary side hydraulic pressure estimated value Pos2.

DPos2 / dt =-(β + g) Pos2 + β · Poc2 + g · Po ... (7)

However, in the equation (7), β represents the reciprocal of the rising time constant and is expressed as follows using the coefficient in the equation (5).

Β = B / (V2 · R2)

In the equation (7), g is an arbitrary coefficient and is set within a range satisfying the following relationship in order to stabilize the estimated oil pressure value Pos2 shown in the equation (7).

Β + g> 0

Next, the processing operation in the ECU 57D according to the fifth embodiment of the present invention will be described with reference to the flowchart of FIG. First, in steps S401 and S402 similar to the above, the initial value Pos (0) of the estimated hydraulic pressure and the hydraulic target value Poc are read.

Then, in step S503, the oil pressure Po from the oil pressure sensor 10 is read and stored in the memory. Next, in step S504, the estimated hydraulic pressure value Pos (k) calculated in the previous sampling and the desired hydraulic pressure value Po
Based on c (k) and the measured oil pressure Po (k), the current oil pressure estimated value Pos (k + 1) is calculated by the following equation (8).
Ask for.

Pos (k + 1) = a4 · Pos (k) + b4 · Poc (k) + d4 · Po (8)

However, in the formula (6), a4 and b4
Is a coefficient used in the filter calculation of the first-order lag characteristic, and d4 is a constant by which the hydraulic pressure Po is multiplied. Also,
At the time of the first sampling, the previous estimated oil pressure value P
The initial value Pos (0) is used as os (k).

Thereafter, the process returns to step S402 for reading the hydraulic pressure target value Poc, and the next sampling process is started. Note that, as described above, instead of the hydraulic pressure Po, the CVT
The estimated hydraulic pressure value Pos may be input to the controller 57f.

Here, the validity of the estimated hydraulic pressure value Pos obtained by the hydraulic pressure estimation observer 57m will be specifically described. For example, from the above equation (2), the secondary side hydraulic model is expressed as the following equation (9) using the hydraulic pressure target value Poc2, the actual hydraulic pressure P2, and the reciprocal β of the rising time constant. .

DP2 / dt = −β · P2 + β · Poc2 (9)

The hydraulic pressure target value Poc2 in the above equation (9) is
Since the secondary side line pressure Pv2 in the equation (2) substantially matches, the equation (9) is substantially equivalent to the equation (2). Also,
Laplace conversion of each state quantity is shown with L,
The transfer function is expressed by the following expression (10).

LP2 = β · LPoc2 / (s + β) (10)

Here, in order to obtain the estimated hydraulic pressure value Pos2, an observer is constructed as in the above equation (7) for the secondary hydraulic model. Let e be the estimation error of the observer of equation (7) (actual hydraulic pressure value P2−estimated hydraulic pressure value Pos2), and estimate noise e when the noise Δpo is added to the measured hydraulic pressure Po. The following expression (11) is obtained.

Le / ΔLP2 = g / {s + (β + g)} (11)

From equation (11), it can be seen that the smaller the coefficient g, the smaller the estimation error e in the low frequency region. When g = 0, the hydraulic pressure estimation observer 57m in FIG. 11 substantially matches the hydraulic response model 57k in FIG.

Further, it is assumed that the influence of modeling error of the hydraulic model is the form of the disturbance d, that is, the case where the actual hydraulic model is expressed by the following equation (12).

DP2 / dt = −β · P2 + β · Poc2 + d (12)

The equation (12) is obtained by adding the disturbance d to the equation (9). When the transfer function Le represents the estimation error when the time change of the hydraulic pressure P2 is expressed as in Expression (12), it becomes as in Expression (13) below using the Laplace-transformed disturbance Ld.

Le / Ld = 1 / {s + (β + g)} (13)

From equation (13), it can be seen that the larger the coefficient g, the smaller the estimation error in the low frequency region.

That is, when the noise component Δpo superimposed on the hydraulic pressure Po measured from the CVT 55 is large, the coefficient g is set small, and when the modeling error of the hydraulic model is large, the coefficient g is set large. e will be small. Therefore, the value of the hydraulic pressure Po can be detected in the optimum state by adjusting the coefficient g according to the actual state.

As described above mathematically, by inputting the oil pressure Po (measured value) together with the oil pressure target value Poc to the oil pressure estimation observer 57m, the oil pressure estimated value by the oil pressure response model 57k in FIG. 9 is obtained. And CVT5
An estimated oil pressure value Po intermediate to the oil pressure Po measured from 5
s is obtained.

Thus, the coefficient g is set small when the noise component ΔPo superposed on the hydraulic pressure Po (measured value) is large, and the coefficient g is set when the modeling error of the hydraulic model is large.
If the value is increased, the estimation error e is reduced and the estimated hydraulic pressure value Pos according to the actual state is obtained.

Example 6. In each of the above embodiments, the temperature is not taken into consideration, but the time constant of the delay characteristic may be changed according to the temperature. In general, the piping resistances R1 and R of the primary side oil passage 5 and the secondary side oil passage 7 (see FIG. 2).
Since 2 changes depending on the oil temperature, the time constants V1 · R1 / B and V2 · R2 / B of the hydraulic response change depending on the oil temperature from the above equations (1) and (2).

A sixth embodiment of the present invention in which the time constant of the delay characteristic is changed based on the temperature information will be described below with reference to the drawings. Here, the case where the oil temperature output from the oil temperature detecting means is used as the temperature information will be described.
For example, detection signals of the engine cooling water temperature or both the oil temperature and the cooling water temperature may be used.

FIG. 13 is a block diagram showing a sixth embodiment of the present invention. 57E corresponds to the ECU 57A, and only the point that the oil temperature Co of the CVT 55 is input to the fuel injection amount corrector 57e. This is different from FIG. 1 (Example 1).

In this case, the CVT 55 is provided with a temperature sensor (not shown) for detecting the oil temperature Co. In addition, the fuel injection amount corrector 57e is based on the oil temperature Co.
The time constant of the first-order lag characteristic (for example, the filter calculation coefficients a1 and b1 in the equation (3)) changes.

Next, the processing operation of the ECU 57E according to the sixth embodiment of the present invention will be described with reference to the flowchart of FIG. In FIG. 14, S101 to S1
Reference numeral 05 is the same step as in FIG. 3, except that an oil temperature detecting step S602 and a time constant changing step S603 are added.

First, when the ignition signal is turned on by the driver's operation, the initial value Fu (0) of the fuel injection amount command value is stored in the memory (step S101), and then in step 602, the CVT 55 oil is stored. The temperature Co is detected and read, and this is stored in the memory. continue,
In step S603, the coefficients a1 and b1 in the above equation (3) are obtained by, for example, map calculation according to the value of the oil temperature Co, and stored in the memory.

For example, if the oil temperature Co rises, the pipe resistances R1 and R2 decrease, and the response delay time constants of the hydraulic pressures P1 and P2 decrease. Therefore, the fuel injection signal J increases in accordance with the increase of the oil temperature Co. The time constant of the first-order lag characteristic may be reduced. Therefore, the coefficient a1 is set to a smaller value and the coefficient b1 is set to a larger value (small time constant) as the oil temperature Co increases, and conversely, the coefficient a1 increases and the coefficient b1 decreases as the oil temperature Co decreases. Value (large time constant)
Is set to

Thereafter, similarly to the above, the fuel injection amount target value Fc (k) from the engine controller 57d is read and stored in the memory (step S102). Subsequently, the fuel injection amount target value Fc (k), the fuel injection amount command value Fu (k) calculated in the previous sampling, and step S603.
The fuel injection amount command value Fu (k + 1) (corresponding to the fuel injection signal J) is calculated by the formula (3) based on the coefficients a1 and b1 obtained in step S3 and stored in the memory (step S103).

In step S103, at the time of the first sampling, the previous value Fu (k) of the fuel injection amount command value is set.
The initial value Fu (0) is used as After the fuel injection device 9a is driven by the fuel injection signal J having the drive pulse length Fp (step S104) according to the fuel injection amount command value Fu (k + 1) thus calculated (step S10).
5) The process returns to step S602 to start the next sampling process.

As described above, the time constant of the first-order lag characteristic in the fuel injection amount corrector 57e is changed based on the oil temperature Co, so that the hydraulic response characteristic due to the oil temperature change actually occurring in the CVT 55 is caused. The response characteristic of the engine 9 can be changed in accordance with the change of.

Incidentally, the above-mentioned time constant changing process step S6.
In the case of 03, the coefficients a1 and b1 from the value of the oil temperature Co
The value of fuel injection amount target value Fc
The time constant between the fuel injection amount command value Fu and the fuel injection amount command value J may be obtained from the map, and then the values of the coefficients a1 and b1 may be calculated according to the time constant.

Further, the case where the time constant of the first-order lag characteristic of the fuel injection signal J in the fuel injection amount corrector 57e is changed based on the oil temperature Co (and at least one of the engine cooling water temperatures) has been described. Similarly, the time constant of the first-order lag characteristic of the estimated oil pressure value Pos in the oil temperature estimation observer 57m (see FIG. 11) may be changed.

Example 7. Further, in the sixth embodiment, the time constant of the first-order lag characteristic is changed based on the temperature information, for example, the oil temperature Co, but the time constant is based on the integrated value (corresponding to the warm-up state) based on the rotation speed Ne of the engine 9. May be changed. Hereinafter, a seventh embodiment of the present invention in which the time constant is changed based on the integrated value of the engine speed Ne will be described with reference to the drawings.

FIG. 15 is a block diagram showing a seventh embodiment of the present invention, and 57F corresponds to the ECU 57E. In this case, the ECU 57F has, in addition to the configuration shown in FIG. 1 (Embodiment 1), a rotation speed integration unit 57n that integrates the rotation speed Ne from the start command of the engine 9 and outputs the rotation speed integration value SNe. However, the fuel injection amount corrector 57e is configured to change the time constant of the first-order lag characteristic based on the rotation speed integrated value SNe.

Next, the processing operation of the ECU 57F according to the seventh embodiment of the present invention will be described with reference to the flowchart of FIG. In FIG. 16, S101 to S1
Reference numeral 05 denotes a step similar to that of FIG. 3, except for the addition of time constant changing steps S702 to S705 based on the rotation speed integrated value.

First, the initial value Fu of the fuel injection amount command value
After storing (0) (step S101), step S
In 702, the integrated value SNe of the engine speed Ne
The initial value SNe (0) of (k) is stored in the memory, and
In step S703, the engine speed Ne (k)
Is read and stored in memory.

Subsequently, in step S704, the rotation speed integrated value SNe (k) calculated in the previous sampling, and
The detected rotation speed Ne (k) is added to calculate the current rotation speed integrated value SNe (k + 1) as in the following Expression (14).

SNe (k + 1) = SNe (k) + Ne (k) (14)

In this way, the rotation speed integrated value SNe (k + 1) is updated and stored in the memory for each sampling.
In the equation (14), at the time of the first sampling, the initial value SNe is set as the previous value SNe (k) of the rotation speed integrated value.
Ne (0) is used.

Next, in step S705, the coefficients a1 and b1 in the equation (3) are obtained by, for example, map calculation based on the value of the revolution speed integrated value SNe (k + 1),
Store in memory. In this case, the increase in the rotation speed integrated value SNe corresponds to the increase in the oil temperature Co in the above-described sixth embodiment, and the fuel injection amount corrector 5 increases as the rotation speed integrated value SNe increases.
The coefficient a1 is set to a small value and the coefficient b1 is set to a large value so that the time constant of the first-order lag characteristic of 7e becomes small.

Thereafter, the fuel injection amount target value Fc from the engine controller 57d is read and stored in the memory (step S102), and the fuel injection amount target value Fc (k) and the fuel injection amount command value calculated in the previous sampling are stored. Fu (k)
Based on (fuel injection signal J) and the modified coefficients a1 and b1, the current fuel injection amount command value Fu (k + 1) is calculated and stored in the memory (step S10).
3).

At the time of the first sampling, the initial value Fu (k) is set as the previous value Fu (k) of the fuel injection amount command value.
(0) is used. Hereinafter, after the driving pulse is output to the fuel injection device 9a by the same process steps S104 and S105 as described above, the process returns to step S703 to start the next sampling process.

Although the values of the coefficients a1 and b1 are directly obtained based on the revolution speed integrated value SNe here, the fuel injection amount target value Fc and the fuel injection amount command value Fu are determined.
The values of the coefficients a1 and b1 may be calculated after obtaining the time constant between (fuel injection signal J) and the map.

As described above, by changing the time constant of the fuel injection amount corrector 57e based on the integrated value SNe of the engine speed Ne which is always input, without using a means for measuring the oil temperature Co, The same effect as that of the sixth embodiment can be obtained.

In each of the above embodiments, the fuel injection amount target value Fc is set based on the accelerator depression amount αo.
It may be set based on the intake air amount Qo passing through the throttle 52, and the accelerator depression amount αo and the intake air amount Qo may be set.
It may be set based on at least one of o.

In each of the above embodiments, the input signals to the engine controller 57d and the CVT controller 57f are, for example, the engine controller 57d.
For d, the accelerator depression amount αo and the intake air amount Qo
For the CVT controller 57f, the depression amount αo of the accelerator pedal 51, the rotation speed Ne of the engine 9 and the hydraulic pressure Po of the CVT 55 are used.
The controller 57f can be changed to an arbitrary input signal by using the pulley rotation speeds N1 and N2 of the CVT 55.

Further, the effects of the above respective embodiments are particularly effective when the engine torque Te becomes large,
For example, only when the fuel injection amount target value Fc becomes large,
The fuel injection amount corrector 57e may be operated.

[0175]

As described above, according to claim 1 of the present invention, a direct injection type engine is used and a delay characteristic generating means is provided. The delay characteristic generating means is a fuel injection amount target for the fuel injection device. When the value changes, a delay characteristic having a time constant larger than the response time constant of the second hydraulic pressure is provided between the fuel injection amount target value and the fuel injection amount command value that determines the actual fuel injection amount. Thus, even when the driver suddenly depresses the accelerator pedal and the fuel injection amount target value suddenly changes, the change in the fuel injection amount to the engine is suppressed, and the engine torque (CVT input torque) becomes the CVT transmittable torque. Since it does not exceed C, it does not decrease the combustion efficiency and increases the harmful components in the exhaust gas, and
It is possible to obtain a control device for a continuously variable transmission and an internal combustion engine that prevents wear of the pulley and the drive belt due to VT belt slip and realizes a good acceleration feeling.

According to claim 2 of the present invention, in claim 1, a steady torque calculating means for calculating a steady torque of the engine from the fuel injection amount and the engine speed,
Transient torque estimating means for estimating the transient response value of the engine torque from the calculated value of the steady torque, hydraulic pressure measuring means for measuring the second hydraulic pressure, gear ratio detecting means for detecting the CVT gear ratio, and second hydraulic pressure. And a transmissible torque calculating means for calculating a transmissible torque of the CVT from the gear ratio, and a torque margin calculating means for obtaining a comparative calculated value between the estimated value of the transient response value and the calculated value of the transmissible torque. Since the delay characteristic generating means is enabled only when the value becomes equal to or less than the predetermined value, the response of the engine torque is suppressed only when there is a risk that the CVT belt slips, and the engine torque is unnecessarily increased. There is an effect that a control device for a continuously variable transmission and an internal combustion engine that prevents the response speed from being reduced can be obtained.

Further, according to claim 3 of the present invention, in claim 2, the time constant of the delay characteristic is set as a function of the comparison calculation value (torque margin). Therefore, the engine torque is set according to the torque margin. It is possible to obtain a control device for a continuously variable transmission and an internal combustion engine, which can suppress the response of the above and can specify an optimum engine output torque rising characteristic according to the torque margin.

Further, according to claim 4 of the present invention, in claim 2, a hydraulic response estimation means (hydraulic response model) is provided, and the hydraulic response estimation means, based on the hydraulic target value for the second hydraulic pressure, A hydraulic response estimated value having a first-order lag characteristic having a response time constant of the second hydraulic pressure as a time constant is obtained, and the transmissible torque calculation means uses the hydraulic response estimated value as the second hydraulic pressure to determine the hydraulic response time. By obtaining the CVT transmissible torque from the estimated hydraulic pressure value of the first-order lag characteristic having a constant as a time constant and the gear ratio detection value, it is possible to detect the hydraulic pressure value in which high frequency vibration and noise are not superimposed without using a filter. Therefore, in particular, when a torque margin for CVT belt slip is obtained from a transient response between the engine torque and the hydraulic pressure, a continuously variable transmission and an internal control which prevent a non-negligible control delay are obtained. The effect of the engine control device is obtained.

According to claim 5 of the present invention, in claim 2, a two-input type hydraulic pressure estimating means (hydraulic pressure estimating observer) is provided, and the hydraulic pressure estimating means is a value obtained by multiplying the second hydraulic pressure by a constant. Based on the hydraulic pressure target value, an estimated hydraulic pressure value having a first-order lag characteristic whose time constant is the sum of the response time constant of the second hydraulic pressure and the constant is obtained, and the transmissible torque calculation means
The transmittable torque is calculated using the estimated hydraulic pressure as the second hydraulic pressure, and an intermediate detection value between the estimated hydraulic response without the first-order lag characteristic and the measured hydraulic pressure is obtained, and the characteristic of the detection signal is determined in the frequency domain. Since the setting is made, there is an effect that a control device for a continuously variable transmission and an internal combustion engine capable of detecting a hydraulic pressure value in an optimum state according to an actual signal state can be obtained.

Further, according to claim 6 of the present invention, in any one of claims 1 to 5, paying attention to the fact that the time constant of the hydraulic response changes depending on the oil temperature, the operation supplied to the CVT. At least one of an oil temperature detecting means for detecting the oil temperature of the oil and a water temperature detecting means for detecting the cooling water temperature of the engine is provided, and the delay characteristic generating means delays based on at least one of the oil temperature and the cooling water temperature. Since the characteristic time constant is changed, the response characteristic of the engine is changed according to the change of the hydraulic response characteristic due to the actual change of the oil temperature of the CVT, and the continuously variable transmission and the internal combustion engine in which the controllability is improved. There is an effect that the control device can be obtained.

According to the seventh aspect of the present invention, in any one of the first to fifth aspects, attention is paid to the fact that the oil temperature rises according to the integrated value of the engine speed, and an engine start command is generated. Since the rotation speed integration means for integrating the rotation speed from the time is provided and the delay characteristic generation means is configured to change the time constant of the delay characteristic based on the integrated value of the rotation speed, the oil temperature can be adjusted without using the temperature sensor. There is an effect that a control device for a continuously variable transmission and an internal combustion engine, which is capable of detecting the rough value of, and changing the response characteristic of the engine in accordance with the change of the hydraulic response characteristic can be obtained.

[Brief description of drawings]

FIG. 1 is a block diagram showing a first embodiment of the present invention.

FIG. 2 is a configuration diagram showing a torque transmission system of an automobile according to the first embodiment of the present invention.

FIG. 3 is a flowchart showing a processing operation according to the first embodiment of the present invention.

FIG. 4 is a characteristic diagram showing a time change of a torque response with respect to an accelerator depression amount in the first embodiment of the present invention.

FIG. 5 is a block diagram showing a second embodiment of the present invention.

FIG. 6 is a flowchart showing a processing operation according to the second embodiment of the present invention.

FIG. 7 is a flowchart showing a processing operation according to the second embodiment of the present invention.

FIG. 8 is a flowchart showing a processing operation according to the third embodiment of the present invention.

FIG. 9 is a block diagram showing a fourth embodiment of the present invention.

FIG. 10 is a flowchart showing a processing operation according to Embodiment 4 of the present invention.

FIG. 11 is a block diagram showing a fifth embodiment of the present invention.

FIG. 12 is a flowchart showing a processing operation according to the fifth embodiment of the present invention.

FIG. 13 is a block diagram showing a sixth embodiment of the present invention.

FIG. 14 is a flowchart showing a processing operation according to the sixth embodiment of the present invention.

FIG. 15 is a block diagram showing Embodiment 7 of the present invention.

FIG. 16 is a flowchart showing a processing operation according to the seventh embodiment of the present invention.

FIG. 17 is a block diagram showing a conventional control device for a continuously variable transmission and an internal combustion engine.

FIG. 18 is a configuration diagram showing a conventional torque transmission system of an automobile.

[Explanation of reference numerals] 1 speed change mechanism, 1a drive belt, 2 primary side pulley,
2a primary side fixed piece, 2b primary side movable piece, 2c primary side hydraulic chamber, 3 secondary side pulley, 3a secondary side fixed piece,
3b Secondary side movable piece, 3c Secondary side hydraulic chamber, 4 Primary side valve, 5 Primary side oil passage, 6 Secondary side valve, 7 Secondary side oil passage, 9 Engine, 9a Fuel injection device, 10 Oil pressure sensor, 51 accelerator pedal, 52 throttle,
55 CVT (continuously variable transmission), 57A-57F EC
U, 57d Engine controller, 57e Fuel injection amount corrector, 57f CVT controller, 57g Gear ratio calculator, 57h Engine torque estimator (steady torque calculating means, transient torque estimating means), 57i Torque margin calculator, 57j Transmittable torque Calculation means, 57k hydraulic pressure response model (hydraulic response estimation means), 57m hydraulic pressure estimation observer (two-input type hydraulic pressure estimation means), 57n engine speed accumulator, Co oil temperature, d4 constant, F fuel,
Fc fuel injection amount target value, J fuel injection signal, Ne engine speed, P1 first hydraulic pressure, P2 second hydraulic pressure,
Po Measured oil pressure, Poc oil pressure target value, Pos
Estimated hydraulic pressure, RT Gear ratio, SNe total number of revolutions, T
e Engine torque, Tes Estimated transient response value, T
m transmittable torque, To output torque, ΔT torque margin (comparative calculation value), S103, S217 step of giving delay characteristic to fuel injection command value, step of S211 comparing torque margin with predetermined value, step of S311 torque margin to time constant For obtaining S403 and S504
Step of giving delay characteristic to estimated oil pressure value, S603
Step of changing the time constant from the oil temperature, S705 Step of changing the time constant from the rotation speed integrated value.

Claims (7)

[Claims] 1. An in-cylinder direct injection type internal combustion engine having a fuel injection device for directly injecting fuel into each cylinder; a continuously variable transmission for transmitting output torque of the internal combustion engine to wheels; A control means for controlling parameters and first and second hydraulic pressures for the continuously variable transmission, wherein the continuously variable transmission has two pulleys, one end of which is axially movable, and the two pulleys. A drive belt stretched between them, a first hydraulic means for applying the first hydraulic pressure for changing the gear ratio to one of the two pulleys, and the above-mentioned means for preventing the drive belt from slipping. A control device for a continuously variable transmission and an internal combustion engine, comprising: a second hydraulic means for applying a second hydraulic pressure to the other of the two pulleys; and the control means includes a delay characteristic generating means, Characteristic generation means, When the fuel injection amount target value for the fuel injection device changes, a difference between the fuel injection amount target value and the fuel injection amount command value that determines the actual fuel injection amount is greater than the second hydraulic response time constant. A control device for a continuously variable transmission and an internal combustion engine, which is characterized by having a delay characteristic with a large time constant. 2. The control means calculates a steady torque of the internal combustion engine from the fuel injection amount and the rotational speed of the internal combustion engine, and a generated torque of the internal combustion engine from the calculated value of the steady torque. A transient torque estimating means for estimating a transient response value, a hydraulic pressure measuring means for measuring the second hydraulic pressure, a gear ratio detecting means for detecting a gear ratio of the continuously variable transmission, the second hydraulic pressure and the A transmissible torque calculating means for calculating a transmissible torque of the continuously variable transmission based on a gear ratio, and a torque margin calculating means for obtaining a comparative calculation value of the estimated value of the transient response value and the calculated value of the transmissible torque. The control device for the continuously variable transmission and the internal combustion engine according to claim 1, wherein the delay characteristic generating means is enabled only when the comparison calculation value becomes equal to or less than a predetermined value. 3. The control device for a continuously variable transmission and an internal combustion engine according to claim 2, wherein the delay characteristic generating means sets the time constant of the delay characteristic as a function of the comparison calculation value. 4. The control means includes a hydraulic pressure response estimating means, and the hydraulic pressure response estimating means sets a response time constant of the second hydraulic pressure to a time constant based on a hydraulic pressure target value for the second hydraulic pressure. The hydraulic response estimated value having a first-order lag characteristic is obtained, and the transferable torque computing means uses the hydraulic response estimated value as the second hydraulic pressure.
A control device for a continuously variable transmission and an internal combustion engine according to item 1. 5. The control means includes a two-input type hydraulic pressure estimating means, and the hydraulic pressure estimating means calculates the second hydraulic pressure based on a value obtained by multiplying the second hydraulic pressure by a constant and the hydraulic pressure target value. A hydraulic pressure estimated value having a first-order lag characteristic whose time constant is a sum of a response time constant of hydraulic pressure and the constant is obtained, and the transmissible torque computing means uses the hydraulic pressure estimated value as the second hydraulic pressure. The control device for a continuously variable transmission and an internal combustion engine according to claim 2. 6. The control means includes at least one of an oil temperature detecting means for detecting an oil temperature of hydraulic oil supplied to the continuously variable transmission and a water temperature detecting means for detecting a cooling water temperature of the internal combustion engine. It is included, The said delay characteristic generation means changes the time constant of the said delay characteristic based on at least one of the said oil temperature and the said cooling water temperature, The claim 1 to 5 characterized by the above-mentioned.
13. A control device for a continuously variable transmission and an internal combustion engine according to any one of items 1 to 10. 7. The control means includes a rotational speed integration means for integrating the rotational speeds of the internal combustion engine from the time when a start command is issued, and the delay characteristic generation means includes the delay based on the integrated value of the rotational speeds. The control device for a continuously variable transmission and an internal combustion engine according to any one of claims 1 to 5, wherein the time constant of the characteristic is changed.