JP2001182599A - Control device of direct injection engine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To effectively reduce occurrence of torque shock when different combustion states are swtiched in an in-cylinder injection engine. SOLUTION: This control device for a direct injection engine 1 with four air cylinders has an injector 7 for directly injecting fuel into a combustion engine and selectively performs stratification combustion or uniform premixed combustion according to an operation state. The control device determines whether the combustion state is transferred between the stratification combustion and the uniform premixed combustion. When it determines that the combustion state is transferred, it performs a transfer control for sequentially switching the combustion state every air cylinder. The switch of the combustion state for all air cylinders is performed for a period longer than one cycle. For example, the combustion state for each air cylinder is switched at a switch interval (e.g. 3/4 cycle or 5/4 cycle) longer than an ignition interval (1/4 cycle) between air cylinders.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a control system for a direct injection engine which performs combustion using fuel directly injected into a combustion chamber, and more particularly to switching control of a combustion mode between stratified combustion and uniform mixed combustion. About.

[0002]

2. Description of the Related Art Conventionally, in order to improve fuel efficiency, an in-cylinder injection engine in which an injector is provided in a combustion chamber and fuel is directly injected into a cylinder has been proposed. In general, in a direct injection engine (also referred to as a direct injection engine), one of a combustion mode of stratified combustion with an emphasis on fuel efficiency or a uniform mixed combustion with an emphasis on output is selectively executed according to the operating state of the engine. . As is well known, the air-fuel ratio in stratified combustion is significantly different from the air-fuel ratio in uniform mixed combustion.
Therefore, when switching from one combustion mode to the other combustion mode, a torque shock occurs due to a change in the air-fuel ratio,
There is a problem that drivability of the direct injection engine is deteriorated.

[0003] In order to reduce such a torque shock, for example, Japanese Patent Application Laid-Open No. 9-151771 discloses a technique for transiently performing intermediate combustion control when switching the combustion mode in a direct injection engine. I have. In this intermediate control, fuel injection is performed in the compression stroke as in the case of stratified charge combustion, and the air-fuel ratio A / F is changed to the air-fuel ratio (A / F
F = 30-40) and a value (A / F = 22-) that is thinner than the air-fuel ratio (A / F = 12-22) of homogeneous combustion (homogeneous combustion).
30). Then, between the homogeneous combustion and the intermediate control, the fuel injection amount is changed stepwise in order to suppress an increase in the HC concentration in the exhaust gas. Further, between the stratified combustion and the intermediate control, the fuel injection amount is gradually changed to smooth the torque fluctuation.

[0004] However, this publication does not mention at all the switching order and switching interval of each cylinder in the intermediate control. Therefore, for example, when the combustion mode of each cylinder is switched in the same order as the ignition order, the torque fluctuation (which occurs to some extent even if the intermediate control is performed) generated for each cylinder is short (one cycle). Focus on As a result, there is a possibility that a torque shock that the driver can feel is generated.

On the other hand, Japanese Patent Application Laid-Open No. Hei 10-184430 discloses that in a lean burn engine that selectively performs operation at a lean air-fuel ratio and operation at a rich air-fuel ratio, torque shock generated during execution of a rich spike is suppressed. Techniques are disclosed. This rich spike means that NOx generated during the lean operation is occluded in the three-way catalyst, and when the predetermined amount of storage is reached, the lean operation is temporarily shifted to the rich operation. At the time of the transition control between the lean operation and the rich operation, the switching of the operating state of each cylinder is sequentially performed at a switching interval of two engine revolutions (one cycle) or more. As a result, the torque fluctuation generated in each cylinder is dispersed by the length of the switching interval between the cylinders, and the torque shock can be effectively reduced.

However, the technique disclosed in this publication relates to an MPI (Multi-point Injection) system in which an injector is provided in an intake manifold, and does not relate to a direct injection engine. In other words, the technique disclosed in this publication is for switching between lean operation and rich operation in one combustion mode called homogeneous mixed combustion, and is not related to switching between different combustion modes such as stratified combustion and uniform mixed combustion. . As is well known, stratified combustion is a combustion method in which a rich portion (near an ignition plug) and a thin portion of an air-fuel mixture are formed in a combustion chamber, and the rich portion that is easy to ignite is ignited to spread the entire combustion. Such a stratified mixture can be formed in an in-cylinder injection engine in which fuel is directly injected into the combustion chamber by an injector provided in the combustion chamber, but such a stratified state is formed by an injector provided outside the combustion chamber. I can't.

[0007]

In order to improve the drivability of the direct injection engine, it is necessary to suppress the torque shock generated when shifting to a different combustion mode at a higher level.

It is an object of the present invention to provide a control apparatus for a direct injection engine which can effectively reduce the occurrence of torque shock when switching between different combustion modes.

[0009]

According to a first aspect of the present invention, there is provided an injector having an injector for directly injecting fuel into a combustion chamber, and selectively performing stratified combustion or uniform mixed combustion according to an operation state. In the control device of the in-cylinder injection engine performed in the following, a determination unit that determines whether to shift the combustion mode between stratified combustion and uniform mixed combustion, and if the determination unit determines that the combustion mode is to be shifted, Control means for performing transition control for sequentially switching the combustion mode for each cylinder. This control means switches the combustion mode of all cylinders over a period longer than one cycle during the transition control.

A second aspect of the present invention relates to a control apparatus for a direct injection engine having an injector for directly injecting fuel into a combustion chamber and selectively performing stratified combustion or uniform mixed combustion in accordance with an operation state. Determining means for determining whether to shift the combustion mode between combustion and homogeneous mixed combustion,
And control means for performing transition control for sequentially switching the combustion mode for each cylinder when the determination section determines that the combustion mode is to be shifted. This control means switches the combustion mode of each cylinder at a switching interval longer than the ignition interval between the cylinders during the transition control.

Here, in the second invention, the switching interval may be made different depending on the cylinder. In the case of a four-cylinder in-cylinder injection engine, it is desirable to set the switching interval to an interval longer than 1/4 cycle, for example, 5/4 cycle or 3/4 cycle.

In the first and second aspects of the present invention, when switching the combustion mode of one cylinder, the control means controls the air-fuel ratio in the uniform mixed combustion and a first threshold set to a leaner side than the air-fuel ratio. Value, the air-fuel ratio is gradually changed, and between the air-fuel ratio in stratified combustion and a second threshold value set on the richer side than the air-fuel ratio, the air-fuel ratio is gradually changed, and Between the first threshold value and the second threshold value, it is preferable to perform control for changing the air-fuel ratio in a stepwise manner.

Further, in the above configuration, based on at least one of the engine speed and the accelerator opening,
Detecting means for detecting the change speed of the operating region may be provided. The control means executes the shift control when the change speed of the operation region detected by the detection means is equal to or less than the determination threshold.

Further, a fuel injection control means for controlling the fuel injection amount of the injector may be added to the above configuration. In this case, during stratified charge combustion, the fuel injection control means starts fuel injection in the compression stroke and ends fuel injection immediately before ignition. Then, during uniform mixed combustion, fuel injection is started in a stroke earlier than in stratified combustion.

[0015]

FIG. 1 is an overall configuration diagram of an example of a four-cylinder horizontally opposed in-cylinder injection engine as an example. Among the cylinders # 1 to # 4 of the in-cylinder injection engine 1, the cylinders # 1 and # 3
Are arranged in the left bank, and the cylinders # 2 and # 4 are arranged in the right bank. Each intake port of the engine 1 is provided with an intake valve 2, and these intake ports communicate with an intake manifold 3. on the other hand,
Each exhaust port of the engine 1 is provided with an exhaust valve 4, and these exhaust ports communicate with an exhaust manifold 5. An ignition plug 6 for igniting the air-fuel mixture is provided at the center of the combustion chamber of each of the cylinders # 1 to # 4 in the cylinder head. An injector 7 for directly injecting fuel (gasoline) into the combustion chamber is provided near the intake valve 2 in each combustion chamber. The in-cylinder injection engine 1 requires the injector 7
Is supplied with high-pressure fuel.

The air cleaner 8 provided in the intake passage has:
It is connected to an air chamber 9 communicating with the intake manifold 3. Between the air cleaner 8 and the air chamber 9, an electric throttle valve 1 for adjusting the amount of intake air is provided.
0 is interposed. The throttle valve 10 is operated by an electric motor 11 and does not have a structure mechanically linked to an accelerator pedal 18. The opening of the throttle valve 10 (throttle opening) is controlled by a control device 12 (hereinafter, referred to as a microcomputer) mainly composed of a microcomputer.
"ECU").

On the other hand, the exhaust manifold 5 communicates with the three-way catalytic converter 13 via an exhaust passage, and has a NOx storage catalytic converter 14 downstream thereof.
Is provided. And these catalytic converters 1
Exhaust gas purified through the exhaust pipes 3 and 14 is discharged through the muffler 15.

An exhaust gas recirculation passage 16 is provided between the exhaust manifold 5 and the air chamber 9. The exhaust gas recirculation passage 16 has an exhaust gas recirculation valve 17 (hereinafter referred to as “E”) for adjusting the amount of exhaust gas recirculated to the intake passage.
GR valve). The EGR valve 17 is driven by a built-in stepper motor,
The opening is set by an output signal from the ECU 12. By adjusting the opening of the EGR valve 17 (EGR opening), an inert gas that does not contribute to combustion is appropriately mixed into the intake air flowing through the intake passage. As a result, the combustion temperature can be reduced, and the emission amount of NOx contained in the exhaust gas can be reduced.

The ECU 12 performs calculations relating to the fuel injection amount, fuel injection timing, ignition timing, throttle opening, EGR opening, and the like according to a control program stored in the ROM. Then, a control signal corresponding to the calculated control amount is output to various actuators. Sensor signals from various sensors including the sensors 20 to 24 are input to the ECU 12 in order to detect an operation state. The intake air amount sensor 20 is provided immediately downstream of the air cleaner 8 and is a hot wire type or hot film type sensor that detects the intake air amount Q. The throttle opening sensor 21 is a sensor that detects the throttle opening θt. The accelerator opening sensor 22 is constituted by a potentiometer or the like, and is a sensor that detects the amount of depression of the accelerator pedal 18 (accelerator opening θa) that indicates the required load of the driver. The engine speed sensor 23 is a sensor that detects the engine speed Ne, and may be a crank angle sensor disposed near the outer periphery of a crank rotor that rotates integrally with the crankshaft of the engine 1. The air-fuel ratio sensor 24 is provided immediately upstream of the three-way catalytic converter 13 and is a sensor for detecting an exhaust air-fuel ratio A / F (equivalent to an exhaust equivalent ratio Φex) from exhaust gas flowing through an exhaust passage. As the air-fuel ratio sensor 24, for example, a linear O2 sensor can be used.

The fuel injection amount by the injector 7 is basically set to a value at which combustion corresponding to the target air-fuel ratio according to the operating state is performed (air-fuel ratio control). This target air-fuel ratio is set to a value according to the combustion mode.
It is appropriately set within the range of 30 to 40 for normal stratified combustion, and 12 to 18 for normal homogeneous mixed combustion. However, when a disturbance factor is added, the exhaust air-fuel ratio A / F (actual air-fuel ratio) does not match the target air-fuel ratio. Therefore, the exhaust air-fuel ratio A / A detected by the air-fuel ratio sensor 24
Air-fuel ratio feedback control based on F (actual air-fuel ratio) is performed. Specifically, the exhaust air-fuel ratio A / F is detected from the output voltage V of the linear O2 sensor, which is the air-fuel ratio sensor 24.
Since the output voltage V of the sensor and the exhaust air-fuel ratio are in a one-to-one relationship, a table summarizing the relationship between the two is created through experiments, simulations, or the like.
It is stored in the ROM in U12. Then, feedback correction is performed on the fuel injection amount such that the deviation between the exhaust air-fuel ratio specified with reference to the table and the target air-fuel ratio is reduced. As a result, the exhaust air-fuel ratio can be made to converge to the target air-fuel ratio, and the control accuracy of the air-fuel ratio can be improved.

In the air-fuel ratio control described later, an equivalent ratio is used instead of the air-fuel ratio as a control variable. Here, the equivalent ratio is the reciprocal of the excess air ratio λ, and is expressed as a stoichiometric air-fuel ratio (for example, 14.5) / actual air-fuel ratio. If the equivalence ratio is determined, the air-fuel ratio is also uniquely specified, so that either the air-fuel ratio or the equivalence ratio may be used as a control variable of the air-fuel ratio. The target equivalence ratio corresponding to the above-described target air-fuel ratio is specified by using a method such as a table reference based on the driving state requested by the driver's required load (for example, accelerator opening θa), engine speed Ne, and the like. . In addition, since the air-fuel ratio is largely different between the uniform mixed combustion and the stratified combustion, the equivalent ratio also greatly differs depending on the combustion mode. Therefore, in the present embodiment, tables for calculating the target equivalence ratio are separately prepared for uniform mixed combustion and stratified combustion.

FIG. 2 is a flowchart showing a routine for determining the transition of the combustion mode. The ECU 12 repeatedly executes this determination routine in a predetermined execution cycle (for example, 10 ms) to determine the transition of the combustion mode. Then, the transition control of the combustion mode is executed by another routine according to the content of the transition execution flag PFM that is the determination result. First, in step 1, the ECU 12 determines the engine speed Ne specified by the engine speed sensor 23 and the accelerator opening θ specified by the accelerator opening sensor 22.
a (corresponding to the driver's required load).

In the following step 2, a target combustion mode (stratified combustion or uniform mixed combustion) according to the operating state of the engine is selected with reference to the combustion mode determination map shown in FIG. As can be seen from the map in the figure, in the target combustion mode, stratified charge combustion is set in the low-load, low-speed region, and uniform mixed combustion is set in other regions. Here, the stratified combustion is a combustion method in which fuel injection by the injector 7 is started in the compression stroke and ended immediately before ignition, and the rear end of the fuel spray is ignited by the ignition plug 6 to burn the air-fuel mixture. The stratified combustion uses only the air around the fuel and can obtain stable combustion with an extremely small amount of fuel relative to the amount of charged air, and thus is suitable for low-load and low-speed engine operation. On the other hand, in the homogeneous mixed combustion, the fuel is injected in a stroke earlier than the stratified combustion (for example, the end of the exhaust stroke or the intake stroke), the injected fuel is sufficiently diffused in the cylinder, and after the injected fuel and the air are uniformly mixed, It is a combustion system that ignites. Uniform mixed combustion has a high air utilization factor and can improve the output of the engine, and is therefore suitable for high-load high-speed operation.

Then, in step 3, it is determined whether or not the transition execution flag PFM is "1". The transition execution flag PFM is initially set to “0”,
“1” means that the transition control of the combustion mode is being performed.
If a negative determination is made in step 3 (PFM =
"0"), the process proceeds to step 4, and it is determined whether or not the currently performed combustion mode matches the target combustion mode determined in step 2. If an affirmative determination is made in step 4, the process proceeds to the return without changing the shift execution flag PFM (= "0"), and the execution of this routine in the current execution cycle ends. While the current combustion mode matches the target combustion mode, step 1
Since the series of steps from to 4 is repeated,
The transition execution flag PFM remains “0”. Therefore, during that period, normal combustion control is continued.

Thereafter, if the operating state changes and the current combustion mode is different from the target combustion mode determined in step 2, the result of the determination in step 4 changes from affirmative to negative, so that the transition execution flag PFM is set to ""1" is set (step 5). As a result, normal combustion control is temporarily interrupted, and transition control for transitioning the combustion mode is instructed. Details of the transition control will be described later.

When the shift execution flag PFM is set to "1" in step 5, the determination result of step 3 changes from negative to affirmative in subsequent execution cycles, so that the process proceeds from step 3 to step 6. This step 6
Is related to the end determination of the transition control, and it is determined whether or not the reference equivalent ratio Φ has reached the target equivalent ratio Φaftr in the combustion mode after the transition. Target equivalent ratio Φaf after this shift
tr is, if the combustion mode after the transition is stratified combustion, for example,
It is set appropriately at 0.36 to 0.48 (30 to 40 in air-fuel ratio conversion). When the combustion mode after the transition is homogeneous mixed combustion, the target equivalent ratio Φaftr is appropriately set, for example, at 0.81 to 1.21 (12 to 18 in air-fuel ratio conversion).

The reference equivalent ratio Φ is calculated according to a reference equivalent ratio calculation routine shown in FIG. Although the details will be described later, by repeating this calculation routine, the reference equivalent ratio Φ changes with time (decreases from uniform mixing to stratification and increases from stratification to uniform mixing) during the transition control. The target equivalent ratio Φaf after the shift
Until tr is reached (meaning the end of transition control)
Since the result of the determination in step 6 is negative, the transition execution flag PFM is maintained at "1", and the transition control is continued. Then, when the reference equivalent ratio Φ reaches the post-shift target equivalent ratio Φaftr, the result of the determination in step 6 changes from negative to affirmative, so that the shift execution flag PFM is reset to “0” (step 7). As a result, the transition control ends, and the control returns to the normal control.

When the transition execution flag PFM is set to "1" (step 5), the transition control from the current combustion mode to the target combustion mode is started. In this transition control, the switching of the combustion mode in each cylinder is sequentially performed on the assumption that the equivalent ratio (final equivalent ratio φfinal) is within the range in which the switching of the combustion mode can be performed (for example, 18 to 27 in air-fuel ratio conversion). Is done. The reason for setting the range of the reference equivalence ratio Φ as the condition for switching the combustion mode is to suppress the rapid change of the air-fuel ratio due to the switching of the combustion mode as much as possible and to reduce the torque shock.

FIG. 7 is an explanatory diagram showing the order of switching the combustion mode for each of the cylinders # 1 to # 4 of the engine 1.
The ignition order of the four cylinders # 1 to # 4 is set in advance, and the cylinder # 1 → the cylinder #
Ignition, that is, combustion is performed in the order of 3 → cylinder # 2 → cylinder # 4. Further, the circles shown in the figure indicate the combustion mode before the transition, and the crosses indicate the combustion mode (the target combustion mode) after the transition.
In switching between the cylinders, first, the combustion mode of the cylinder # 1 is switched to the target combustion mode at the time of combustion of the cylinder # 1 in the first cycle (one cycle is two engine revolutions in a four-cycle engine). The timing at which the combustion mode of cylinder # 1 switches is denoted by t # 1. Next, the combustion mode of the cylinder # 3 is switched during the combustion of the cylinder # 3 in the second cycle, and the combustion mode of the cylinder # 2 is switched during the combustion of the cylinder # 2 in the third cycle. Each cylinder #
The timings at which the combustion modes 3 and # 2 are switched are denoted by t # 3 and t # 2, respectively. Finally, the combustion mode of the cylinder # 4 is switched during the combustion of the cylinder # 4 in the fourth cycle (the switching timing t # 4), and the switching of the combustion mode for all the cylinders # 1 to # 4 is completed. Then, as described above, when the reference equivalent ratio Φ reaches the post-transition target equivalent ratio Φaftr, the transition execution flag PFM is reset to “0”, and the transition control ends (steps 6 and 6 in FIG. 2).
7).

In the switching order shown in FIG. 7, the switching interval of the combustion mode for each of the cylinders # 1 to # 4 is 5/4 cycle. By setting the switching interval longer than the cylinder ignition interval (1/4 cycle), it is possible to effectively reduce the occurrence of torque shock that the driver can feel. This point will be described in comparison with the combustion mode switching order shown in FIG. FIG.
In the switching order shown in (1), the combustion mode of each of the cylinders # 1 to # 4 is switched in the same order (same interval) as the ignition order. Therefore, the switching interval is 1/4 cycle, and the switching of the combustion mode of all cylinders # 1 to # 4 is 1 cycle.
Complete between cycles. The torque fluctuation itself generated in each cylinder can be reduced to some extent by transition control described later. However, in the switching order shown in FIG. 6, since torque fluctuations in all cylinders # 1 to # 4 concentrate in a short period of one cycle, a torque shock that the driver can feel is likely to occur. On the other hand, in the switching order shown in FIG. 7, each cylinder # has a switching interval (5/4 cycle) longer than the ignition interval (1/4 cycle).
The combustion modes # 1 to # 4 are sequentially switched. The period required from the first switching in cylinder # 1 to the completion of the last switching in cylinder # 4 is 15/4 cycle. Therefore, the torque fluctuation of each of the cylinders # 1 to # 4 is dispersed by an amount corresponding to an increase in the switching interval, in other words, by an increase in the period required for switching all the cylinders # 1 to # 4. As a result, the torque shock can be reduced more effectively in the order shown in FIG. 7 than in the switching order shown in FIG.

Next, specific transition control of the combustion mode will be described. FIG. 3 is a flowchart showing a routine for calculating the reference equivalent ratio Φ. This calculation routine is based on EC
It is repeatedly executed in a predetermined execution cycle (for example, 10 ms) by U12. The reference equivalent ratio Φ is used to calculate a final equivalent ratio Φfinal which is a control variable of the fuel injection pulse width Ti. Therefore, the fuel injection pulse width Ti (corresponding to the fuel injection amount of the injector 7) calculated from the final equivalent ratio Φfinal basically changes according to the reference equivalent ratio Φ (with exceptions). The reference equivalence ratio Φ is also used when making a transition end determination in the determination routine of FIG. 2 (step 6).

First, at step 11, it is determined whether or not the transition control flag PFM is "1". The content of the transition control flag PFM is set by the determination routine shown in FIG. If a negative determination is made in step 11, that is, during the normal control period, the target equivalent ratio Φbfr before the transition is set as the reference equivalent ratio Φ (step 12). On the other hand, if the determination in step 11 is affirmative, that is, if the transition execution flag PFM is “1”, the flow proceeds to step 13 to determine which combustion mode the transition is to (uniform mixed combustion → stratification). Combustion or stratified combustion → homogeneous mixed combustion). When the transition from the homogeneous mixture combustion to the stratified combustion is performed, an affirmative determination is made in step 13, and
Subtraction processing of the reference equivalent ratio Φ is performed (step 14).
That is, a value obtained by subtracting the predetermined control constant α from the current reference equivalent ratio Φ is set as a new reference equivalent ratio Φ.
On the other hand, when shifting from stratified combustion to homogeneous mixed combustion,
A negative determination is made in step 13, and an addition process of the reference equivalent ratio Φ is performed (step 15). That is, a value obtained by adding a predetermined control constant β to the current reference equivalent ratio Φ is set as a new reference equivalent ratio Φ. When the shift execution flag PFM is reset to “0” again by the shift determination routine (end of the shift control), the determination result of step 11 changes from affirmative to negative, so that the reference equivalent ratio Φ is set to the target after shift. Is set to the equivalent ratio Φaftr (step 1
2).

As can be understood from the above description, by repeating the calculation routine shown in FIG. 3, at the time of transition from uniform mixed combustion to stratified combustion, the reference equivalent ratio Φ is calculated from the target equivalent ratio Φbfr of uniform mixed combustion. It gradually decreases (the air-fuel ratio increases). Further, at the time of transition from stratified combustion to uniform mixed combustion, the reference equivalent ratio Φ gradually increases from the target equivalent ratio Φbfr of stratified combustion (the air-fuel ratio decreases).

FIG. 4 is a flowchart showing a routine for calculating the fuel injection pulse width Ti. This calculation routine is executed by a predetermined execution cycle (for example, 10
ms). Note that this calculation routine is executed separately for each of the cylinders # 1 to # 4. As a result, the final equivalent ratio Φfinal (# 1) to Φf
inal (# 4) is calculated, and the fuel injection pulse width Ti is also calculated for each cylinder.

First, at step 21, it is calculated whether or not the transition execution flag PFM is "1". At the time of the normal control, from the negative determination in step 21 to step 2
Proceeding to 2, the reference equivalence ratio Φ is set as the final equivalence ratio Φfinal. The reference equivalent ratio Φ is calculated based on the calculation routine shown in FIG.

On the other hand, if an affirmative determination is made in step 21, the process proceeds to step 23, in which the present combustion mode for the cylinders # 1 to # 4 (target cylinders) to be switched is determined. As described above, the combustion mode is switched by the cylinder # 1 (timing t # 1) → the cylinder # 3 (timing t # 3) → the cylinder # 2 (timing t # 2) → the cylinder # 4.
(Timing t # 4), and the switching interval of each cylinder is 5/4 cycle. Therefore, cylinder # 1
Note that during a part of the period (t # 1 to t # 4) from the switching timing t # 1 to the switching timing t # 4 of the cylinder # 4, the combustion mode differs between the cylinders # 1 to # 4. I want to be.

If a negative determination is made in step 23, that is, if uniform mixed combustion is being performed in the target cylinder, the flow advances to step 25 to determine whether or not the reference equivalence ratio Φ is larger than the first threshold value Φlim1. Is determined. The first threshold value Φlim1 is a value appropriately set based on the upper limit value (uniform lean limit air-fuel ratio) of the air-fuel ratio at which uniform mixed combustion can be performed. For example, a uniform lean limit air-fuel ratio of 20
In this case, the equivalent ratio corresponding to the air-fuel ratio is about 0.73 (uniform lean limit equivalent ratio). That is, if the equivalence ratio is smaller than 0.73, the mixture is too lean to perform appropriate homogeneous mixed combustion. Therefore, a value obtained by adding a predetermined margin MRG to the uniform lean limit equivalent ratio is set as the first threshold value Φlim1. Then, on the basis of system control, if the equivalence ratio is smaller than the first threshold value Φlim1, it is treated that uniform mixed combustion cannot be executed.

If an affirmative determination is made in step 25, that is, if the reference equivalent ratio Φ is larger than the first threshold value Φlim1, the process proceeds to step 27, where the final equivalent ratio Φ
The value of the reference equivalent ratio Φ is set to final. That is,
When the reference equivalence ratio Φ is within the range in which the uniform mixed combustion can be performed, the value of the reference equivalence ratio Φ is used as it is as the final equivalence ratio Φfinal which is a control variable of the fuel injection pulse width Ti. On the other hand, if the reference equivalent ratio Φ is equal to or smaller than the first threshold value Φlim1, the process proceeds from step 25 to step 28, and the value of the first threshold value Φlim1 is set to the final equivalent ratio Φfinal. That is, when the reference equivalence ratio Φ is too small to execute the uniform mixed combustion, the final equivalence ratio Φfinal is held at the first threshold Φlim1, which is the lean limit value that can maintain the combustion mode.

If the determination in step 23 is affirmative, that is, if stratified charge combustion is being performed in the target cylinder, the process proceeds to step 24, where it is determined whether the reference equivalent ratio Φ is smaller than the second threshold value Φlim2. It is determined whether or not. The second threshold value Φlim2 is a value appropriately set based on the lower limit value (stratified rich limit air-fuel ratio) of the air-fuel ratio at which stratified combustion can be performed. For example, when the stratified rich limit air-fuel ratio is 25, the equivalent ratio corresponding to this air-fuel ratio is 0.58 (stratified rich limit equivalent ratio). That is, if the equivalence ratio is larger than 0.58, the air-fuel ratio is too small to perform appropriate stratified combustion. Therefore, a value obtained by subtracting a predetermined margin MRG from the stratified rich limit equivalent ratio is set as the second threshold value Φlim2. Then, on the basis of system control, stratified combustion is handled as being infeasible if the equivalence ratio is larger than the second threshold value Φlim2.

The first threshold value Φlim1 and the second threshold value Φlim2 are set in consideration of both the suppression of torque fluctuation and the stability of combustion. As will be described later, when the combustion mode is switched, the final equivalent ratio Φfinal switches stepwise between the first threshold value Φlim1 and the second threshold value Φlim2. Changes to
Therefore, if the main focus is on suppressing the torque fluctuation due to the air-fuel ratio change, the change amount of two threshold values | Φlim1−Φli
It is preferable that m2 | is as small as possible, that is, the margin MRG is as small as possible. However, if the margin MRG is too small, the actual air-fuel ratio in each combustion mode may be below the stratified rich / uniform lean limit air-fuel ratio in relation to the accuracy of the combustion control. Therefore, these thresholds Φlim1 and Φlim2 should be set to values that appropriately satisfy both the stability of combustion and the suppression of torque fluctuation.

If an affirmative determination is made in step 24, that is, if the reference equivalent ratio Φ is smaller than the second threshold value Φlim2, the final equivalent ratio Φfinal is added to the reference equivalent ratio Φfinal.
Is set (step 27). That is, when the reference equivalence ratio Φ is in a range in which stratified combustion can be performed, the final equivalence ratio Φfinal, which is a control variable of the fuel injection pulse width Ti, is used.
Is used as it is. On the other hand, if the reference equivalent ratio Φ is equal to or larger than the second threshold Φlim2, the value of the second threshold Φlim2 is set to the final equivalent ratio Φfinal (step 26). That is, when the reference equivalence ratio Φ is too large to perform stratified combustion, the final equivalence ratio Φfinal is held at the second threshold Φlim2 which is a limit value on the rich side that can maintain the combustion mode.

When the final equivalent ratio Φfinal is calculated in steps 22, 26, 27 and 28, the following step 29
, The basic fuel injection pulse width Tp is calculated. The basic fuel injection pulse width Tp is determined by the engine speed sensor 23
Based on the engine speed Ne and the intake air amount Q specified by the intake air amount sensor 20 and the intake air amount Q, respectively, it is calculated according to the following equation. In the equation, K is an injector characteristic correction coefficient (set individually for each of the cylinders # 1 to # 4).

## EQU1 ## Tp = K.times.Q / Ne

Then, the fuel injection pulse width Ti is calculated from the following equation based on the final equivalent ratio Φfinal calculated in the above step, the basic fuel injection pulse width Tp calculated in step 29, and the like (step 30).

[Formula 2] Ti = Tp × Φfinal × LAMBDA + Ts

Here, LAMBDA is an air-fuel ratio feedback correction coefficient, and by appropriately setting this correction coefficient, the exhaust equivalent ratio Φex calculated from the air-fuel ratio sensor 24 is obtained.
(Equivalent equivalence ratio) is controlled so as to converge to the target equivalence ratio Φbfr. Ts is an invalid injection pulse width determined by the battery voltage.

Then, the fuel injection pulse width Ti calculated in step 30 is output to the injector 7 of the target cylinder (step 31). Accordingly, a fuel amount corresponding to the fuel injection pulse width Ti is directly injected from the injector 7 into the cylinder. By performing the above-described fuel injection control, control is performed such that the exhaust equivalent ratio Φex (exhaust air-fuel ratio A / F) matches the target equivalent ratio (target air-fuel ratio) even during the transition control.

Next, a specific transition control will be described by taking as an example a case where the transition from the uniform mixed combustion to the stratified combustion is performed. FIG. 5 is a timing chart in this transition control. At timing t0, the transition execution flag PFM is switched to “1” by the transition determination routine shown in FIG.
The transition control is started. Since the throttle opening θt changes from the current throttle opening θbfr to the target throttle opening θaftr (θbfr <θaftr) for stratified charge combustion, the intake air amount Q also increases due to the primary delay. In addition, from the timing t0 to the timing t1 at the end of the transition control, the reference equivalent ratio Φ gradually decreases according to the reference equivalent ratio calculation routine shown in FIG. 3 (steps 11, 13, and 14). Until the reference equivalence ratio Φ reaches the first threshold Φlim1, the final equivalence ratios Φfinal (# 1) to Φfinal (# 4) of all the cylinders # 1 to # 4 similarly decrease,
The corresponding fuel is injected from the injector 7 of each cylinder (steps 21, 23, 25, 27, 29 to 3).
1). As the intake air amount Q increases, the basic fuel injection pulse width Tp shown in the above equation 1 increases, but the final equivalent ratio Φfinal decreases. Therefore, the fuel injection pulse width Tp shown in Expression 2 is substantially constant (or slightly increased). Therefore, the actual air-fuel ratio gradually increases toward the air-fuel ratio of stratified combustion.

The timing t # 1 is the first ignition timing of the cylinder # 1 to be switched in a state where the combustion mode can be switched (the air-fuel ratio is 18 to 27). At this timing t # 1, the combustion mode of the cylinder # 1 switches to stratified combustion, and the final equivalent ratio Φfinal (# 1) becomes the reference equivalent ratio Φ
(Φlim1 if Φ> Φlim1) from the second threshold Φlim2
Changes step by step. This is because the calculation procedure of the final equivalent ratio Φfinal (# 1) changes from step 27 (or 28) to step 26 when the determination result of step 23 changes from negative to positive at the switching timing t # 1. . Then, thereafter, the final equivalent ratio Φ of the cylinder # 1
final (# 1) is the second threshold Φlim until the reference equivalence ratio Φ reaches the second threshold Φlim2 (timing t2).
It is held at 2 (step 26). On the other hand, cylinder # 1
For the other cylinders # 2 to # 4, the uniform mixed combustion is continued after the timing t # 1. Therefore, the reference equivalent ratio Φ
Reaches the first threshold value Φlim1, the final equivalent ratios Φfinal (# 2) to Φfinal (# 4) of the cylinders # 2 to # 4 are held at the first threshold value Φlim1 (step 28). ).

The timing t # 3 after the timing t # 1 is 2
This is the ignition timing of the switching target cylinder # 3.
Until timing t # 3, the final equivalent ratio Φfinal of cylinder # 3
(# 3) is held at the first threshold value Φlim1 (step 28). Then, at this timing t # 3, the combustion mode of the cylinder # 3 is switched to stratified combustion, and the final equivalent ratio Φfinal (# 3) is changed from the first threshold value Φlim1 set in step 28 to step It changes stepwise to a second threshold value Φlim2 set at 26. Then, thereafter, the final equivalent ratio Φfinal (# 3) of the cylinder # 3
Is held at the second threshold value Φlim2 until timing t2 (step 26). On the other hand, for the cylinders # 2 and # 4, the uniform mixed combustion is continued even after the timing t # 3. Therefore, the final equivalent ratio Φfinal of cylinders # 2 and # 4
(# 2) and Φfinal (# 4) are held at the first threshold Φlim1 (step 28).

Until the ignition timing t # 2 of the third target cylinder, cylinder # 2, the final equivalent ratio Φfinal (#
2) is held at the first threshold value Φlim1 (step 28). Then, at this timing t # 2, the cylinder #
The combustion mode 2 is switched to stratified combustion, and the final equivalent ratio Φfinal (# 2) changes stepwise from the first threshold Φlim1 to the second threshold Φlim2. Then, similarly to the case of the cylinders # 1 and # 3, the final equivalent ratio Φfi of the cylinder # 2
nal (# 2) is held at the second threshold value Φlim2 (step 26). On the other hand, for the cylinder # 4, the uniform mixed combustion is continued even after the timing t # 2. Therefore, the final equivalent ratio Φfinal (# 4) of the cylinder # 4 is equal to the first threshold Φlim
It is held at 1 (step 28).

Until the ignition timing t # 4 of the last target cylinder, cylinder # 4, the final equivalence ratio Φfinal (# 4) of cylinder # 4
Is held at the first threshold value Φlim1 (step 28). At this timing t # 4, the final equivalent ratio Φfinal (# 4) of the cylinder # 4 changes stepwise to the second threshold value Φlim2 and is held at the second threshold value Φlim2 (step 26). ). Thereby, all cylinders # 1 to # 4
Is completed.

When the reference equivalence ratio Φ reaches the second threshold Φlim2 at the timing t2, step 2
The determination result of No. 4 changes from negative to positive. Accordingly, the procedure for calculating the final equivalent ratios φfinal (# 1) to φfinal (# 4) of each cylinder changes from step 26 to step 27.
After timing t2, the final equivalent ratio Φfinal (# 1) to Φfinal
(# 4) decreases with the reference equivalent ratio Φ. When the reference equivalence ratio Φ reaches the target equivalent ratio Φaftr after the transition (timing t1), the transition execution flag PFM is reset to “0”, and the transition control ends.

As described above, in this embodiment, during the transition control of the combustion mode, the combustion mode of the four cylinders # 1 to # 4 is sequentially switched over a period longer than one cycle. That is, the combustion mode of each cylinder is switched at a switching interval (5/4 cycle) longer than the ignition interval (1/4 cycle) between adjacent cylinders in the ignition order. Therefore, the torque fluctuations of the respective cylinders are dispersed by an amount corresponding to an increase in the switching interval, in other words, by an increase in the period required for the switching of all the cylinders.
It is possible to reduce torque shock. Further, at the time of the transition control, the torque fluctuation generated in each cylinder is suppressed as much as possible by controlling the fuel injection amount of each cylinder. In other words, the first and second thresholds Φlim1 and Φlim2 are set so that the stepwise change of the final equivalent ratio Φfinal before and after the switching of the combustion mode becomes as small as possible. Thereby, torque shock can be reduced more effectively, and drivability of the direct injection engine can be further improved.

In the above embodiment, while the torque shock can be suppressed, the transition control period becomes longer. As a result, the driver may experience an acceleration lag depending on the driving state, which may lead to a decrease in drivability. Therefore, it is preferable to selectively execute the transition control method according to the magnitude of the change speed of the operation region specified based on the accelerator opening θa or the like by the following method.

FIG. 11 is a flowchart showing a combustion mode transition mode determination routine. First, step 41
, The accelerator opening θa detected by the accelerator opening sensor 22 is read. Next, in step 42, in order to determine the magnitude of the change speed of the operation region,
Accelerator opening change speed Δθa / Δt is calculated. This can be calculated as a change amount of the accelerator opening θa per predetermined time.

Then, in step 43, it is determined whether or not the accelerator opening change speed Δθa / Δt is greater than a predetermined determination threshold TH. If an affirmative determination is made in step 43 (Δθa / Δt> TH), that is, if the change speed of the operating region is high, step 44
The normal shift control is instructed. In this case, the transition control is performed according to the switching order shown in FIG. As described above, in this switching order, each of the cylinders # 1 to # 4
Is concentrated in a short period of one cycle, so that a torque shock that the driver can feel is likely to occur. However, in such an operating state, it is necessary to shift the combustion mode quickly, and the fluctuation in the required torque of the driver itself is large (the driver itself also expects a certain degree of torque shock). Therefore,
It is not necessary to consider much the problem of torque shock associated with the transition of the combustion mode.

On the other hand, if a negative determination is made in step 43 (Δθa / Δt ≦ TH), that is, if the speed of change in the operating range is low, the process proceeds to step 45, where torque shock mitigation shift control is instructed. In this case, the transition control is performed according to the switching order shown in FIG. As described above, in this switching order, the torque fluctuation of each of the cylinders # 1 to # 4 is dispersed over several cycles, so that the torque shock can be reduced. Therefore, it is preferable to perform the torque shock mitigation shift control in a driving state in which the driver is likely to feel the torque shock sensitively, such as when the vehicle is running at a constant vehicle speed.

The speed of change in the operating state can be specified by using the engine speed Ne (or both sensor outputs θa and Ne may be used together) instead of the accelerator opening θa. In this case, when the engine speed Ne is higher than a predetermined threshold value, it is determined that the change speed of the operating region is high, and when it is equal to or less than the threshold value, it is determined that the change speed is low.

The order of switching the combustion modes of the cylinders # 1 to # 4 is not limited to the method described in the above embodiment. The present invention can be widely applied to various methods including the following modifications.

FIG. 8 is an explanatory diagram showing a modification of the order of switching the combustion mode. First, at the time of combustion of cylinder # 1 in the first cycle, the combustion mode of cylinder # 1 is switched to the target combustion mode. Further, during the combustion of cylinder # 4 in the same cycle, the combustion mode of cylinder # 4 is switched. Cylinders # 1 and # 4 adjacent to each other in the order of switching the combustion mode
The switching interval between them is 3/4 cycle. Then, the combustion mode of the cylinder # 2 is switched during the combustion of the cylinder # 2 in the subsequent second cycle (the adjacent cylinders # 4 and # 4).
The interval between # 2 is also 3/4 cycle). Finally, the combustion mode of the cylinder # 3 is switched during the combustion of the cylinder # 3 in the third cycle (the interval between the adjacent cylinders # 2 and # 3 is also 3/4).
Cycle), the switching of the combustion mode for all cylinders # 1 to # 4 is completed.

In the example described above, the combustion mode is switched in the order of cylinder # 1 → cylinder # 4 → cylinder # 2 → cylinder # 3, and the switching interval between adjacent cylinders is in the order shown in FIG. 3 shorter than the switching interval (5/4 cycle)
/ 4 cycle. Therefore, the period required for switching all the cylinders # 1 to # 4 can be shortened as compared with the case of FIG.

FIG. 9 is an explanatory diagram showing another modification of the order of switching the combustion mode. First, at the time of combustion of cylinder # 1 in the first cycle, the combustion mode of cylinder # 1 is switched to the target combustion mode. In addition, during the combustion of cylinder # 2 in the same cycle, the combustion mode of cylinder # 2 is switched. The switching interval between adjacent cylinders # 1 and # 2 is 1 /
Two cycles. Then, the combustion mode of the cylinder # 3 is switched during the combustion of the cylinder # 3 in the subsequent second cycle (the switching interval between the adjacent cylinders # 2 and # 3 is 3/4 cycle). Finally, the combustion mode of the cylinder # 4 is switched during the combustion of the cylinder # 4 in the second cycle (the interval between the adjacent cylinders # 3 and # 4 is １ ／ cycle), and the combustion for all the cylinders # 1 to # 4 is performed. The form switching is completed.

In the above-described example, the combustion mode is switched in the order of cylinder # 1, cylinder # 2, cylinder # 3, and cylinder # 4. The switching interval between adjacent cylinders is changed depending on the cylinder (1/2 cycle or 3/4 cycle).
By changing the switching interval for each cylinder in this manner, the period required for controlling the transition of the combustion mode can be further reduced as compared with the case of FIG.

The above description relates to a four-cylinder engine, but it is obvious that the present invention can also be applied to an in-cylinder injection engine having more cylinders (a six-cylinder engine, an eight-cylinder engine, etc.). . For example, when the present invention is applied to a six-cylinder engine, the ignition interval between the cylinders is 1/6 cycle. Therefore, the switching interval of each cylinder may be set longer than 1/6 cycle (for example, 7/6 cycle interval). .

[0064]

As described above, according to the present invention, when the combustion mode is shifted between stratified combustion and uniform mixed combustion, the switching interval between adjacent cylinders in the switching order is
It is set longer than the ignition interval between cylinders. Therefore, the torque fluctuation generated in each cylinder is dispersed, so that the torque shock that the driver can feel can be effectively suppressed. As a result, it is possible to further improve the drivability of the direct injection engine.

[Brief description of the drawings]

FIG. 1 is an overall configuration diagram of a direct injection engine.

FIG. 2 is a flowchart showing a transition determination routine of a combustion mode.

FIG. 3 is a flowchart showing a routine for calculating a reference equivalent ratio.

FIG. 4 is a flowchart showing a routine for calculating a fuel injection pulse width.

FIG. 5 is a timing chart in a case where a transition is made from uniform mixed combustion to stratified combustion.

FIG. 6 is an explanatory diagram showing the order of switching the combustion mode in the normal shift control.

FIG. 7 is an explanatory diagram showing an order of switching combustion modes in torque shock mitigation shift control.

FIG. 8 is an explanatory diagram showing a modification of the order of switching the combustion mode.

FIG. 9 is an explanatory diagram showing another modification of the order of switching the combustion mode.

FIG. 10 is a diagram showing a combustion mode determination map.

FIG. 11 is a flowchart showing a combustion mode transition mode determination routine;

[Explanation of symbols]

1 In-cylinder injection engine, 2 Intake valve, 3 Intake manifold, 4 Exhaust valve, 5 Exhaust manifold, 6 Spark plug, 7 Injector, 8 Air cleaner, 9 Air chamber, 10
Electric throttle valve, 11 electric motor,
12 control device (ECU), 13 three-way catalytic converter, 14 NOx storage catalytic converter,
15 muffler, 16 exhaust gas recirculation passage, 17 exhaust gas recirculation (EGR) valve, 18 accelerator pedal, 20 intake air amount sensor, 21
Throttle opening sensor, 22 accelerator opening sensor,
23 engine speed sensor, 24 air-fuel ratio sensor

 ──────────────────────────────────────────────────続 き Continued on the front page F term (reference) 3G084 BA05 BA13 BA17 CA03 CA09 DA04 DA11 DA12 DA15 EB08 EB11 EB16 EC02 FA07 FA10 FA26 FA33 FA39 3G301 HA04 HA07 HA16 JA04 KA08 LA03 LB04 MA01 MA12 NA07 NB11 NC02 ND03 ND12 NE15 NE03 NE03 NE03 NE15 NE22 PA01Z PA11Z PD02Z PE01Z PE04Z

Claims (8)

[Claims] An in-cylinder injection engine control apparatus having an injector for directly injecting fuel into a combustion chamber and selectively performing stratified combustion or uniform mixed combustion according to an operation state. Determining means for determining whether or not to shift the combustion mode, and control means for performing a shift control for sequentially switching the combustion mode for each cylinder when the determination means determines that the combustion mode is to be shifted. The control device of the direct injection engine, wherein the control unit switches the combustion mode of all cylinders over a period longer than one cycle during the transition control. 2. A control device for an in-cylinder injection engine having an injector for directly injecting fuel into a combustion chamber and selectively performing stratified combustion or uniform mixed combustion according to an operation state. Determining means for determining whether or not to shift the combustion mode, and control means for performing a shift control for sequentially switching the combustion mode for each cylinder when the determination means determines that the combustion mode is to be shifted. A control device for a direct injection engine, wherein the control means switches the combustion mode of each cylinder at a switching interval longer than an ignition interval between cylinders during the transition control. 3. The control device for a direct injection engine according to claim 2, wherein the switching interval is different for each cylinder. 4. The control apparatus according to claim 2, wherein the in-cylinder injection engine is a four-cylinder engine, and the switching interval is longer than 1/4 cycle. 5. The control apparatus for a direct injection engine according to claim 4, wherein the switching interval is 5/4 cycle or 3/4 cycle. 6. When switching the combustion mode of a single cylinder, the control means sets the air-fuel ratio between the air-fuel ratio in uniform mixed combustion and a first threshold value set leaner than the air-fuel ratio. The air-fuel ratio is gradually changed between the air-fuel ratio in the stratified combustion and a second threshold value set to be richer than the air-fuel ratio, and the air-fuel ratio is gradually changed. 3. The control device for a direct injection engine according to claim 1, wherein control is performed to change the air-fuel ratio in a stepwise manner between the second threshold value and the second threshold value. 4. 7. A control device according to claim 1, further comprising a detecting unit configured to detect a change speed of the operating region based on at least one of an engine speed and an accelerator opening, wherein the control unit controls the change of the operating region detected by the detecting unit. 7. The control device for a direct injection engine according to claim 1, wherein the shift control is executed when the speed is equal to or less than a determination threshold. 8. A fuel injection control means for controlling a fuel injection amount of the injector, wherein the fuel injection control means starts fuel injection in a compression stroke and performs fuel injection immediately before ignition during stratified combustion. The control device for a direct injection engine according to any one of claims 1 to 7, wherein the fuel injection is started in a stroke earlier than the stratified charge combustion when the mixed combustion is completed.