阅读说明：本技术 直喷发动机的控制装置以及控制方法 (Control device and control method for direct injection engine ) 是由 葛西理晴 儿玉贵义 于 2017-06-15 设计创作，主要内容包括：在发动机的运转区域中,在低负荷侧的第1区域内进行均质燃烧,另一方面,在与第1区域相比处于高负荷侧的第2区域内进行如下分层燃烧,即,通过第1喷射动作而使得燃料分散至缸内,通过第2喷射动作而使得燃料偏向火花塞附近。在发动机的运转状态从第1区域向第2区域转移的区域转移时,执行基于分层燃烧的转移控制,在转移控制中,通过第2喷射动作而喷射比第2区域内的该第2喷射动作的目标量多的量的燃料,然后使得第2喷射动作的喷射量趋向目标量而减小。(In the operating region of the engine, homogeneous combustion is performed in the 1 st region on the low load side, while stratified combustion is performed in the 2 nd region on the high load side of the 1 st region, in which fuel is dispersed into the cylinder by the 1 st injection operation and fuel is biased to the vicinity of the spark plug by the 2 nd injection operation. When the operating state of the engine shifts from the 1 st region to the 2 nd region, a shift control based on stratified combustion is executed in which fuel of an amount larger than a target amount of the 2 nd injection event in the 2 nd region is injected by the 2 nd injection event and then the injection amount of the 2 nd injection event is made to decrease toward the target amount.)

1. A control method of a direct injection engine, the direct injection engine having:

a spark plug; and

a fuel injection valve provided so as to be able to directly inject fuel into the cylinder, wherein,

in the operating region of the engine, homogeneous combustion is performed in a 1 st region on a low load side, while stratified combustion is performed in a 2 nd region on a high load side compared to the 1 st region, in which fuel is dispersed into a cylinder by a 1 st injection operation of the fuel injection valve and fuel is biased to the vicinity of the spark plug by a 2 nd injection operation of the fuel injection valve,

executing transition control based on the stratified combustion when an operating state of an engine transitions from the 1 st region to the 2 nd region,

in the transition control, a larger amount of fuel than a target amount of the 2 nd injection event in the 2 nd region is injected by the 2 nd injection event, and then the injection amount of the 2 nd injection event is made to decrease toward the target amount.

2. The control method of the direct injection engine according to claim 1,

the excess air ratio of the mixed gas is set to be about 2 in both the 1 st zone and the 2 nd zone.

3. The control method of the direct injection engine according to claim 1 or 2,

the 1 st injection operation is performed in an intake stroke, and the 2 nd injection operation is performed in a compression stroke.

4. The control method of the direct injection engine according to any one of claims 1 to 3,

in the 2 nd region, an ignition timing that is later than the ignition timing in the 1 st region is set as a target ignition timing of the ignition plug,

the 2 nd injection event based on the target amount is performed immediately before the target ignition timing is reached.

5. The control method of the direct injection engine according to claim 4,

in the transition control, an interval from an injection timing of the 2 nd injection action to an ignition timing of the ignition plug is set to a constant value,

after retarding the ignition timing from the target ignition timing, the ignition timing is advanced toward the target ignition timing in accordance with a decrease in the injection amount of the 2 nd injection event.

6. The control method of the direct injection engine according to claim 4,

in the transition control, the ignition timing of the ignition plug is set to the target ignition timing,

the interval from the injection timing of the 2 nd injection event to the ignition timing is shortened in accordance with the decrease in the injection amount of the 2 nd injection event with respect to the interval immediately after the transition to the 2 nd region.

7. The control method of the direct injection engine according to claim 4,

in the transition control, the injection timing of the 2 nd injection operation is set to be constant,

the interval from the injection timing of the 2 nd injection event to the ignition timing of the ignition plug is shortened in accordance with a decrease in the injection amount of the 2 nd injection event with respect to the interval immediately after the transition to the 2 nd region.

8. The control method of the direct injection engine according to claim 4,

in the transition control, the ignition timing of the ignition plug is retarded from the ignition timing in the 1 st region toward the target ignition timing,

the interval from the injection timing of the 2 nd injection event to the ignition timing of the ignition plug is shortened in accordance with a decrease in the injection amount of the 2 nd injection event with respect to the interval immediately after the transition to the 2 nd region.

9. The control method of the direct injection engine according to any one of claims 1 to 8,

is configured to be capable of changing the compression ratio of the engine,

in the 2 nd region, a compression ratio lower than that in the 1 st region is set.

10. The control method of the direct injection engine according to any one of claims 1 to 9,

in the 2 nd region, a compression ratio higher than a compression ratio at which knocking can be suppressed in the case of operation by homogeneous combustion in the same operating state is set.

11. A control device for a direct injection engine, wherein,

the control device for the direct injection engine includes:

a spark plug;

a fuel injection valve provided so as to be able to directly inject fuel into the cylinder; and

a controller that controls operations of the ignition plug and the fuel injection valve,

the controller has:

an operating state detection unit that detects an operating state of the engine;

a combustion state control unit that controls a combustion state in the cylinder based on an operating state of the engine; and

an ignition control unit that sets an ignition timing of the ignition plug,

the combustion state control unit is configured to control the combustion state,

in the case where the operating state of the engine is in the 1 st region on the low load side, the engine is operated by homogeneous combustion, and in the case where the operating state of the engine is in the 2 nd region on the high load side of the 1 st region, the engine is operated by stratified combustion in which fuel is dispersed into the cylinder by the 1 st injection operation of the fuel injection valve and fuel is biased toward the vicinity of the ignition plug by the 2 nd injection operation of the fuel injection valve,

executing transition control based on the stratified combustion when the operating state of the engine transitions from the 1 st region to the 2 nd region,

in the transition control, fuel in an amount larger than a target amount of the 2 nd injection action in the 2 nd region is injected by the 2 nd injection action, and then the injection amount of the 2 nd injection action is made to decrease toward the target amount.

Technical Field

The present invention relates to a direct injection engine capable of switching a combustion method according to an operation region, and a control method thereof.

Background

In order to further reduce the environmental load, the demand for improving the fuel consumption of internal combustion engines is increasing. The thinning of the mixture gas is a known countermeasure for improving the fuel consumption of the internal combustion engine. JPH10-231746 discloses a direct injection engine configured to be able to switch the combustion method according to the operating range, wherein the combustion method is switched from stratified combustion to homogeneous combustion according to an increase in the engine load when accelerating from a low-speed low-load range. In the operation based on the homogeneous combustion, the fuel is injected in the intake stroke, and in the operation based on the stratified combustion, the fuel is injected in the compression stroke. In the region of operation by stratified combustion, particularly in the region on the high load side, fuel is injected in both the intake stroke and the compression stroke (paragraphs 0036 and 0037).

Disclosure of Invention

The inventors of the present invention have made studies to perform an operation in which the air excess ratio of the air-fuel mixture is set to a value higher than the stoichiometric air-fuel ratio equivalent value in the entire operation region of the engine, and the operation is performed by homogeneous combustion in the operation region on the low load side, while fuel injection is performed a plurality of times in one combustion cycle in the operation region on the high load side, fuel is dispersed in the cylinder by the 1 st injection action, and the fuel is unevenly dispersed and combusted in the vicinity of the spark plug by the 2 nd injection action performed after the 1 st injection action (hereinafter, referred to as "stratified combustion", and may be particularly referred to as "weakly stratified combustion" in order to distinguish from stratified combustion in the case where fuel injection is performed only in the compression stroke).

Here, in the operation based on the stratified combustion, it is preferable to limit the injection amount in the 2 nd injection action to a small amount from the viewpoint of suppressing the emission of NOx. Further, when switching from homogeneous combustion to stratified combustion with respect to an increase in engine load, if the injection amount of the 2 nd injection event is limited to a small amount immediately after the switching, a sufficient amount of fuel is not injected as the injection amount of the 2 nd injection event, and combustion may become unstable. On the other hand, in order to avoid instability of combustion, if only the injection amount of the 2 nd injection action is increased, not only the NOx emission amount may be increased, but also the combustion may become excessively rapid.

The purpose of the present invention is to appropriately switch from homogeneous combustion to stratified combustion in a direct injection engine that performs homogeneous combustion in an operating region on the low load side and stratified combustion in an operating region on the high load side without impairing combustion stability.

One aspect of the present invention provides a control method for a direct injection engine.

A method according to an aspect of the present invention is a method for controlling a direct injection engine including: a spark plug; and a fuel injection valve provided so as to be able to directly inject fuel into the cylinder. In the operating region of the engine, homogeneous combustion is performed in the 1 st region on the low load side, while stratified combustion is performed in the 2 nd region on the high load side of the 1 st region, in which fuel is dispersed into the cylinder by the 1 st injection operation and fuel is biased to the vicinity of the spark plug by the 2 nd injection operation. Further, at the time of a region transition of the operating state of the engine from the 1 st region to the 2 nd region, transition control based on stratified combustion is executed in which fuel of an amount larger than a target amount of the 2 nd injection action in the 2 nd region is injected by the 2 nd injection action and then the injection amount of the 2 nd injection action is made to decrease toward the target amount.

Another aspect of the present invention provides a control device for a direct injection engine.

Drawings

Fig. 1 is a configuration diagram of a direct injection engine according to an embodiment of the present invention.

Fig. 2 is a structural diagram of the variable compression ratio mechanism of the engine.

Fig. 3 is an explanatory diagram showing an example of the operation region map of the engine.

Fig. 4 is an explanatory diagram showing the fuel injection timing and the ignition timing according to the operation region.

Fig. 5 is an explanatory view showing a gravity center line of a spray beam of the fuel injection valve.

Fig. 6 is an explanatory diagram showing a positional relationship between the spray and the spark plug.

Fig. 7 is a flowchart showing the entire flow of combustion control (including control at the time of zone transition) according to an embodiment of the present invention.

Fig. 8 is an explanatory diagram showing an example of changes in the excess air ratio, the compression ratio, and the fuel consumption rate with respect to the engine load.

Fig. 9 is an explanatory diagram showing a specific example of control (transition control) performed at the time of the area transition.

Fig. 10 is an explanatory diagram showing another example of the transition control.

Fig. 11 is an explanatory diagram showing another example of the transition control.

Fig. 12 is an explanatory diagram showing another example of the transition control.

Fig. 13 is an explanatory diagram showing a modification of the change in the compression ratio with respect to the engine load.

Detailed Description

Embodiments of the present invention will be described below with reference to the drawings.

(Engine Integrated Structure)

Fig. 1 is a configuration diagram of a direct injection engine (spark ignition engine, hereinafter referred to as "engine") 1 according to an embodiment of the present invention.

The main body of the engine 1 is formed by a cylinder block 1A and a cylinder head 1B, and a block or a cylinder is formed as a space surrounded by the cylinder block 1A and the cylinder head 1B. Fig. 1 shows only 1 cylinder, but the engine 1 may be a multi-cylinder type direct injection engine having a plurality of cylinders.

The piston 2 is inserted into the cylinder block 1A so as to be vertically reciprocable along a cylinder center axis Ax, and the piston 2 is coupled to a crankshaft, not shown, via a connecting rod 3. The reciprocating motion of the piston 2 is transmitted to the crankshaft via the connecting rod 3, and converted into rotational motion of the crankshaft. A chamber 21a is formed in the top face 21 of the piston 2, and the smooth flow of air drawn into the cylinder through the intake port 4a is suppressed from being obstructed by the piston top face 21.

The cylinder head 1B has a lower surface defining a roof-type combustion chamber Ch. A combustion chamber Ch is formed as a space surrounded by the lower surface of the cylinder head 1B and the piston top surface 21. In the cylinder head 1B, as passages for communicating the combustion chambers Ch with the outside of the engine, a pair of intake passages 4 are formed on one side of the cylinder center axis Ax, and a pair of exhaust passages 5 are formed on the other side. An intake valve 8 is provided at a port portion (intake port) 4a of the intake passage 4, and an exhaust valve 9 is provided at a port portion (exhaust port) 5a of the exhaust passage 5. Air taken into the intake passage 4 from outside the engine is taken into the cylinder while the intake valve 8 is open, and exhaust gas after combustion is discharged into the exhaust passage 5 while the exhaust valve 9 is open. A throttle valve, not shown, is provided in the intake passage 4, and the flow rate of air taken into the cylinder is controlled by the throttle valve.

In the cylinder head 1B, further between the intake port 4a and the exhaust port 5a, an ignition plug 6 is provided on the cylinder center axis Ax, and a fuel injection valve 7 is provided between the pair of intake ports 4a, 4a on the side of the cylinder center axis Ax. The position of the ignition plug 6 is preferably in the vicinity of the cylinder center axis Ax and is not limited to the cylinder center axis Ax. The fuel injection valve 7 is configured to be able to receive fuel supply from a high-pressure fuel pump, not shown, and to be able to directly inject the fuel into the cylinder. The fuel injection valve 7 is a multi-hole type fuel injection valve, and is disposed on the intake port 4a side of the cylinder center axis Ax so as to inject fuel in a direction obliquely intersecting the cylinder center axis Ax, in other words, so that a spray beam center line AF described later intersects the cylinder center axis Ax at an acute angle. In the present embodiment, the fuel injection valve 7 is provided at a position surrounded by the ignition plug 6 and the intake ports 4a, 4 a. The fuel injection valve 7 is not limited to this arrangement, and may be provided on the opposite side of the ignition plug 6 with respect to the intake port 4 a.

The tumble control valve 10 is provided in the intake passage 4, and the opening area of the intake passage 4 is substantially reduced by the tumble control valve 10, thereby strengthening the flow of air in the cylinder. In the present embodiment, the air sucked into the cylinder through the intake port 4a forms a tumble flow toward the opposite side of the intake port 4a with respect to the cylinder center axis Ax as the flow of the air, in other words, a tumble flow that passes through the in-cylinder space formed on the exhaust port 5a side in the direction from the lower surface of the cylinder head 1B toward the piston top surface 21, and the tumble flow is intensified by the tumble control valve 10. The intensification of the in-cylinder flow can be achieved by changing the shape of the intake passage 4 without being limited to the provision of the tumble control valve 10. For example, the intake passage 4 may be formed in a more upright state so that the air flows into the cylinder at a more gentle angle with respect to the cylinder center axis Ax, or the center axis of the intake passage 4 may be formed in a more linear state so that the air flows into the cylinder with a stronger force.

An exhaust gas purification device (not shown) is interposed in the exhaust passage 5. In the present embodiment, the catalyst having the oxidation function and the catalyst having the NOx storage reduction function are incorporated in the exhaust gas purification apparatus, and the exhaust gas after combustion discharged to the exhaust passage 5 is purified of Hydrocarbons (HC) by oxygen remaining in the exhaust gas, and then released to the atmosphere after storing the NOx component. As will be described later, in the present embodiment, combustion is performed with the excess air ratio λ of the mixture gas set to a value in the vicinity of 2 over the entire operating region of the engine 1, but in a region on the lean side where the excess air ratio λ is higher than the stoichiometric air-fuel ratio equivalent value, the amounts of discharged carbon monoxide (CO) and nitrogen oxides (NOx) decrease, while HC tends to maintain a constant amount of discharged exhaust. By increasing the excess air ratio λ and setting the air-fuel ratio to be substantially higher than the theoretical value, the NOx emission itself can be suppressed, the capacity of the storage catalyst can be suppressed, and the HC release into the atmosphere can be suppressed.

(Structure of variable compression ratio mechanism)

Fig. 2 is a configuration diagram of a variable compression ratio mechanism provided in the engine 1.

In the present embodiment, the compression ratio of the engine 1 is mechanically changed by changing the top dead center position of the piston 2 by the variable compression ratio mechanism.

The variable compression ratio mechanism couples the piston 2 and the crankshaft 15 via an upper link 31 (connecting rod 3) and a lower link 32, and changes the compression ratio by adjusting the posture of the lower link 32 with a control link 33.

The upper connecting rod 31 is connected to the piston 2 at an upper end by a piston pin 34.

The lower link 32 has a coupling hole at the center thereof, and is coupled to the crankshaft 15 so as to be swingable about the crank pin 15a by inserting the crank pin 15a of the crankshaft 15 into the coupling hole. The lower link 32 is connected at one end to the lower end of the upper link 31 by a connecting pin 35, and at the other end to the upper end of the control link 33 by a connecting pin 36.

The crankshaft 15 has a crank pin 15a, a crank journal 15b, and a balance weight 15c, and is supported by the crank journal 15b with respect to the engine body. The crank pin 15a is disposed at a position eccentric with respect to the crank journal 15 b.

The control link 33 is connected at an upper end to the lower link 32 by a connecting pin 36, and at a lower end to a control shaft 38 by a connecting pin 37. The control shaft 38 is disposed parallel to the crankshaft 15, and a coupling pin 37 is provided at a position eccentric with respect to the center. The control shaft 38 is formed with gears on the outer periphery. The gear of the control shaft 38 is engaged with a pinion 40 driven by an actuator 39, and the pinion 40 is rotated by the actuator 39, so that the control shaft 38 can be rotated, and the posture of the lower link 32 can be changed by the movement of the connecting pin 37.

Specifically, by rotating the control shaft 38 so that the position of the connecting pin 37 is relatively lowered with respect to the center of the control shaft 38, the posture or inclination of the lower link 32 can be changed (in the state shown in fig. 2, the lower link 32 is rotated rightward) so that the position of the connecting pin 35 is relatively raised with respect to the center of the crank pin 15a, and the compression ratio of the engine 1 can be mechanically increased. On the other hand, by rotating the control shaft 38 so that the position of the connecting pin 37 is relatively raised with respect to the center of the control shaft 38, the posture or inclination of the lower link 32 can be changed so that the position of the connecting pin 35 is relatively lowered with respect to the center of the crank pin 15a (in the state shown in fig. 2, the lower link 32 is rotated leftward), and the compression ratio of the engine 1 can be mechanically lowered.

In the present embodiment, the compression ratio is lowered with respect to an increase in the engine load by the variable compression ratio mechanism.

(construction of control System)

The operation of the engine 1 is controlled by an engine controller 101.

In the present embodiment, the engine controller 101 is configured as an electronic control unit, and is configured by a microcomputer having a central processing unit, various storage devices such as a ROM and a RAM, an input/output interface, and the like.

Detection signals of an air flow meter, an air-fuel ratio sensor, and the like, which are not shown, are input in addition to detection signals of the acceleration sensor 201, the rotational speed sensor 202, and the cooling water temperature sensor 203 to the engine controller 101.

The acceleration sensor 201 outputs a signal corresponding to the amount of operation of the accelerator pedal by the operator. The operation amount of the accelerator pedal is an index of a load requested to the engine 1.

The rotation speed sensor 202 outputs a signal corresponding to the rotation speed of the engine 1. As the rotational speed sensor 202, a crank angle sensor may be used, and the rotational speed may be detected by converting a unit crank angle signal or a reference crank angle signal output from the crank angle sensor into a rotational speed per unit time (engine speed).

The cooling water temperature sensor 203 outputs a signal corresponding to the temperature of the engine cooling water. The temperature of the engine lubricating oil may be employed instead of the temperature of the engine cooling water.

The engine controller 101 stores map data in which various operation control parameters of the engine 1 such as a fuel injection amount and the like are assigned to operation states such as a load, a rotational speed, and a cooling water temperature of the engine 1, detects the operation state of the engine 1 at the time of actual operation of the engine 1, sets the fuel injection amount, the fuel injection timing, the ignition timing, the compression ratio and the like based on the map data by referring to the map data, outputs a command signal to a drive circuit of the ignition plug 6 and the fuel injection valve 7, and outputs the command signal to the actuator 39 of the variable compression ratio mechanism.

(outline of Combustion control)

In the present embodiment, the engine 1 is operated with the excess air ratio λ of the mixture gas set to a value near 2. The "air excess ratio" is a value obtained by dividing the air-fuel ratio by the theoretical air-fuel ratio, and includes an air excess ratio of 2 and its vicinity when the air excess ratio is in the vicinity of "2", and in the present embodiment, the air excess ratio is in the range of 28 to 32 in terms of the air-fuel ratio, and preferably an air excess ratio of 30 in terms of the air-fuel ratio is used. The "air excess ratio of the air-fuel mixture" refers to the air excess ratio in the entire cylinder, and specifically refers to a value obtained by dividing the actually supplied air amount by the minimum air amount (mass) theoretically required for combustion of the fuel supplied to the engine 1 in each combustion cycle.

Fig. 3 shows an operation region map of the engine 1 according to the present embodiment.

In the present embodiment, the air excess ratio λ of the air-fuel mixture is set to be in the vicinity of 2 over the entire region where the engine 1 is actually operated, regardless of the engine load. The region in which the engine is operated with the excess air ratio λ set to a value in the vicinity of 2 is not limited to the entire operation region of the engine 1, and may be a partial operation region. For example, the air excess ratio λ may be set to a value around 2 in a low load region and a medium load region of the entire operating region, and may be set to a stoichiometric air-fuel ratio equivalent value (═ 1) by switching the air excess ratio λ in a high load region.

In the present embodiment, in the 1 st region Rl where the engine load is equal to or less than the predetermined value over the entire operating region of the engine 1 in the operating region where the excess air ratio λ is set at around 2, the 1 st predetermined value λ 1 around 2 is set as the excess air ratio λ, and a homogeneous mixed gas obtained by diffusing fuel is formed over the entire cylinder and is combusted. On the other hand, in the 2 nd region Rh in which the engine load is higher than the predetermined value, the air excess ratio λ is set to the 2 nd predetermined value λ 2 in the vicinity of 2, the fuel-rich air-fuel mixture (the 1 st air-fuel mixture) is deflected in the vicinity of the ignition plug 6, and a stratified air-fuel mixture in which the air-fuel mixture (the 2 nd air-fuel mixture) leaner than the 1 st air-fuel mixture is dispersed is formed around the deflected air-fuel mixture, and combustion is performed.

In order to form the stratified mixture, in the present embodiment, the fuel having the air excess ratio of the 2 nd predetermined value (λ ═ λ 2) is injected a plurality of times in one combustion cycle. A part of the fuel for each combustion cycle is injected at the 1 st timing from the intake stroke to the first half of the compression stroke by the 1 st injection action, and at least a part of the remaining fuel is injected at a timing retarded from the 1 st timing, specifically, at the 2 nd timing immediately before the ignition timing of the ignition plug 6 in the second half of the compression stroke by the 2 nd injection action. In the present embodiment, the ignition timing is set in the compression stroke, and therefore the 2 nd timing also becomes the timing in the compression stroke.

Fig. 4 shows the fuel injection timing IT and the ignition timing Ig corresponding to the operating region.

In the 1 st region Rl (low load region) where the operation is performed by the homogeneous combustion, the fuel is supplied for each combustion cycle by 1 injection operation performed in the intake stroke. The engine controller 101 sets the fuel injection timing ITl in the intake stroke, and outputs an injection pulse that continues for a period corresponding to the fuel injection amount from the fuel injection timing ITl to the fuel injection valve 7. The fuel injection valve 7 is driven to open by an injection pulse to inject fuel. In the 1 st region Rl, the ignition timing Igl is set in the compression stroke.

In contrast, in the 2 nd region Rh (high load region) operated by the stratified combustion, the fuel is injected for each combustion cycle divided into 2 times of the intake stroke and the compression stroke. About 90% of the fuel of the entire fuel injection amount is injected by the 1 st injection action as the 1 st injection action, and the remaining 10% of the fuel is injected by the 2 nd injection action as the 2 nd injection action. The injection quantity of the 2 nd injection operation is not limited to a quantity corresponding to 10% of the entire fuel injection quantity, and may be as small as possible in terms of the operation characteristics of the fuel injection valve 7. The engine controller 101 sets the 1 st timing ITh1 in the intake stroke and the 2 nd timing ITh2 in the compression stroke as the fuel injection timings, and outputs an injection pulse that continues for a period corresponding to each fuel injection amount to the fuel injection valve 7. The fuel injection valve 7 is driven to open by the injection pulse, and fuel is injected at the 1 st timing ITh1 and the 2 nd timing ITh2, respectively. The ignition timing Igh is set in the compression stroke also in the 2 nd region Rh, but is set to be later than the ignition timing Igl of the 1 st region Rl.

The excess air ratio λ (1 st predetermined value λ 1) set in the 1 st region Rl on the low load side and the excess air ratio λ (2 nd predetermined value λ 2) set in the 2 nd region Rh on the high load side can be set as appropriate in consideration of the thermal efficiency of the engine 1. The 1 st predetermined value λ 1 and the 2 nd predetermined value λ 2 may be different values from each other, but may be equal values. In the present embodiment, equal values are set (λ 1 ═ λ 2).

(description of Fuel spray)

Fig. 5 shows the spray beam gravity center line AF of the fuel injection valve 7.

As described above, the fuel injection valve 7 is a multi-hole type fuel injection valve, and has 6 injection holes in the present embodiment. The spray beam gravity center line AF is defined as a straight line connecting the front end of the fuel injection valve 7 and the spray beam center CB, and the injection direction of the fuel injection valve 7 is determined as a direction along the spray beam gravity center line AF. The "spray beam center" CB is the center of a virtual circle formed by connecting the tips of the spray beams B1 to B6 at the time when a predetermined time has elapsed since the injection of the spray beams B1 to B6 by the fuel injected from the injection holes.

Fig. 6 shows the positional relationship between the spray (spray beams B1 to B6) and the tip (plug gap G) of the ignition plug 6.

In the present embodiment, the spray beam center of gravity AF is inclined with respect to the central axis of the fuel injection valve 7, and the angle formed by the cylinder central axis Ax and the spray beam center of gravity AF is made larger than the angle formed by the cylinder central axis Ax and the central axis of the fuel injection valve 7. This allows the spray to approach the spark plug 6 and the spray beam (for example, spray beam B4) to pass through the vicinity of the plug gap G. The spray beam passing through the vicinity of the plug gap G is not limited to 1 beam, and may be a plurality of beams, and for example, 2 beams may sandwich the plug gap G.

By passing the spray beam through the vicinity of the plug gap G in this way, in the 2 nd region Rh on the high load side, the mixed gas in the vicinity of the ignition plug 6 can be caused to flow by the kinetic energy of the spray of the fuel injected immediately before the ignition timing Igh is reached, and the fuel contained in the mixed gas in the vicinity of the ignition plug 6 can be made dense, so that the spark plug discharge passage for ignition can be sufficiently extended even after the tumble flow is attenuated or broken, and the ignitability can be ensured. The "spark plug discharge path" refers to an arc generated in the plug gap G at the time of ignition.

(explanation based on the flowchart)

Fig. 7 shows an overall flow of the combustion control according to the present embodiment by a flowchart. The combustion control includes control performed at the time of zone transition (hereinafter referred to as "transition control") according to the present embodiment.

Fig. 8 shows changes in the air excess ratio λ, the compression ratio CR, and the fuel consumption rate ISFC with respect to the engine load.

The combustion control according to the present embodiment will be described with reference to fig. 7 with appropriate reference to fig. 8. The engine controller 101 is programmed to execute the control routine shown in fig. 7 at predetermined intervals.

In the present embodiment, in addition to the switching of the combustion modes (homogeneous combustion, stratified combustion) described above, the compression ratios CRl, CRh of the engine 1 are changed by the variable compression ratio mechanism in accordance with the operating regions Rl, Rh.

In S101, the accelerator opening APO, the engine rotation speed Ne, the cooling water temperature Tw, and the like are read as the operating state of the engine 1. The operating state such as the accelerator opening APO is calculated by an operating state calculation flow separately executed based on detection signals of the acceleration sensor 201, the rotational speed sensor 202, the cooling water temperature sensor 203, and the like.

In S102, it is determined whether or not the operating region of the engine 1 is the 1 st region Rl on the low load side based on the read operating state. Specifically, when the accelerator opening APO is equal to or less than a predetermined value defined for each engine rotation speed Ne, it is determined that the operation region is the 1 st region Rl, the process proceeds to S103, and the engine 1 is operated by homogeneous combustion in the order of S103 to S105. On the other hand, when the accelerator opening APO is higher than the predetermined value for each of the engine rotational speeds Ne, it is determined that the operating region is the 2 nd region Rh on the high load side, the routine proceeds to S106, and the engine 1 is operated by the weak stratified combustion in the order of S106 to S111. In the present embodiment, the transition control is realized by the processing shown in S107 to S109.

In S103, the compression ratio CRl for the 1 st region Rl is set. In the 1 st region Rl, the compression ratio CRl is set to a value as large as possible in a range where knocking does not occur. In the present embodiment, as shown in fig. 8, a target compression ratio having a tendency to decrease with an increase in the engine load is set in advance, and the variable compression ratio mechanism is controlled based on the target compression ratio such that the compression ratio CRl is made to decrease as the engine load increases. However, without being limited to this, the engine 1 may be provided with a knock sensor, and in the case where the occurrence condition of knocking is detected based on the target compression ratio set to a constant value, the variable compression ratio mechanism may be used to lower the compression ratio CRl to suppress knocking.

In S104, the fuel injection amount FQl and the fuel injection timing ITl for the 1 st region Rl are set. Specifically, the fuel injection amount FQl is set based on the load, the rotation speed, and the like of the engine 1, and the fuel injection timing ITl is set. The fuel injection amount FQl and the like are set, for example, in the following manner.

The basic fuel injection amount FQbase is calculated based on the accelerator opening APO and the engine rotation speed Ne, and the fuel injection amount FQ for each combustion cycle is calculated by correcting the basic fuel injection amount FQbase according to the cooling water temperature Tw or the like. Then, the calculated fuel injection amount FQ (FQl) is substituted into the following equation to be converted into an injection period or an injection pulse width Δ t, and the fuel injection timing IT1 is calculated. The basic fuel injection quantity FQbase and the fuel injection timing ITl can be calculated by searching a map that is predetermined appropriately by experiments or the like.

In the above equation (1), FQ represents the fuel injection amount, ρ represents the fuel density, a represents the total injection nozzle area, Cd represents the nozzle flow rate coefficient, Pf represents the fuel injection pressure or the fuel pressure, and Pa represents the cylinder internal pressure.

In S105, the ignition timing Igl for the 1 st region R1 is set. In the 1 st region Rl, the ignition timing Igl in the compression stroke is set. Specifically, the ignition timing Igl is set to MBT (optimal ignition timing) or a timing in the vicinity thereof.

In S106, the compression ratio CRh for the 2 nd region Rh is set. In the 2 nd region Rh, the compression ratio CRh is set to be lower than that of the 1 st region Rl. Further, as in the 1 st region Rl, a target compression ratio having a tendency to decrease with an increase in engine load is set in advance, and the variable compression ratio mechanism is controlled based on the target compression ratio to decrease the compression ratio CRh, but when a knock sensor is provided, the variable compression ratio mechanism may decrease the compression ratio CRh to suppress knocking when occurrence of knocking is detected based on the target compression ratio set to a constant value (lower than a value set in the 1 st region Rl).

Here, in the present embodiment, the compression ratio CRh for the 2 nd region Rh is set to a compression ratio higher than a compression ratio at which knocking can be suppressed when the engine is operated by homogeneous combustion based on the same operating state (engine load). Fig. 8 shows compression ratios capable of suppressing knocking in the case of homogeneous combustion with a two-dot chain line. In this way, in the present embodiment, the compression ratio CRh for the 2 nd region Rh is a compression ratio higher by a constant value than the compression ratio in the case of homogeneous combustion shown by the two-dot chain line. Regarding the 2 nd region Rh, "setting the compression ratio CRh to a compression ratio lower than the 1 st region Rl" means "lower than the 1 st region Rl" as an overall tendency exhibited by the entire engine load.

Fig. 8 shows a change in the air excess ratio λ. In the present embodiment, the excess air ratio λ decreases from λ ═ 2 in the 1 st region Rl with respect to an increase in the engine load, and when the excess air ratio λ shifts from the 1 st region Rl to the 2 nd region Rh, the excess air ratio λ increases to a value slightly greater than 2, and then decreases toward λ ═ 2 in the 2 nd region Rh. Such an operation that the excess air ratio λ shows with respect to an increase in the engine load does not meet the positive design intention of changing the excess air ratio λ itself. The reason why the air excess ratio λ decreases in the 1 st region Rl is to ensure adjustment of ignitability with respect to a decrease in the compression ratio CRl for the purpose of suppressing knocking, in other words, increase correction of fuel in a range in which the effect of thinning by the mixed gas is not impaired. The increase in the excess air ratio λ when shifting from the 1 st region Rl to the 2 nd region Rh is adjusted based on the stratified mixture, so that ignitability can be improved and combustion can be achieved based on a high excess air ratio λ.

In S107, it is determined whether or not the transition control is being executed. Whether or not the transition control is being executed, in other words, whether or not the transition control is completed, can be determined from the injection amount of the 2 nd injection action (hereinafter, sometimes referred to as "2 nd transition injection amount") FQt2 performed in the transition control.

In the present embodiment, after the transition control is started, fuel in an amount larger than the injection amount FQh2 in the normal state of the 2 nd injection operation in the 2 nd region Rh is injected by the 2 nd injection operation, and the 2 nd transition injection amount FQt2 is reduced by the engine 1 each time the engine enters the next cycle to gradually approach the injection amount FQh2 in the normal state. Therefore, the transition control is determined to be completed because the 2 nd transition injection amount FQt2 matches the normal injection amount FQh2 in the 2 nd region Rh. After the transfer control is completed, the engine controller 101 starts normal control. Here, the normal injection amount FQh2 corresponds to the "target amount in the 2 nd region" of the 2 nd injection action.

In S108, the injection amount for the 1 st injection operation (hereinafter, sometimes referred to as "1 st transfer injection amount") FQt1 and the 2 nd transfer injection amount FQt2 during the transfer control are set, and the fuel injection timings ITt1 and ITt2 for the transfer control are set. Specifically, as in the normal operation described later, the fuel injection quantity FQ for each combustion cycle corresponding to the operating state of the engine 1 is calculated, and the portion of the calculated fuel injection quantity FQ having a predetermined ratio is set as the 1 st transition injection quantity FQt1, and the remaining portion is set as the 2 nd transition injection quantity FQt 2. Then, the injection timing ITt1 of the 1 st injection operation and the injection timing ITt2 of the 2 nd injection operation are calculated by substituting the 1 st and 2 nd transfer injection amounts FQt1 and FQt2 into the above expression (1) and converting the amounts into injection periods or injection pulse widths Δ t1a and Δ t2a, respectively.

The ratio Ra of the 1 st transfer injection amount FQt1 to the fuel injection amount FQ for transfer control is calculated as a ratio obtained by subtracting the correction value Δ R from the ratio R (e.g., 90%) set in the normal state (Ra ═ R — Δ R). Further, by setting a large correction value Δ R immediately after the start of the shift control, in other words, immediately after the shift from the 1 st region Rl to the 2 nd region Rh, and decreasing the correction value Δ R each time the number of executions of the shift control is increased once, the 1 st shift injection amount FQt1 is gradually increased from the fuel injection amount immediately after the start of the control, and the 2 nd shift injection amount FQt2 can be made close to the injection amount FQh2 at the normal time.

In the present embodiment, the correction value Δ R is set to a value that varies in the range of 0 to 0.1, and the correction value is set immediately after the start of the shift controlΔ R is set to 0.1(Ra ═ 0.8), the 2 nd transfer injection amount FQt2 is set to 20% of the entire fuel injection amount FQ, and the correction value Δ R is decreased to 0 in accordance with the increase in the number of control executions, whereby the 2 nd transfer injection amount FQt2 is decreased to 10% of the entire fuel injection amount FQ. Then, at the timing when correction value Δ R reaches 0, it is determined that the transition control is completed. If the 2 nd injection operation fails in the middle of the transition control and fuel is not injected, the transition control may be interrupted and the transition to the normal control may be made. In this case, the 2 nd transfer injection amount FQt2 set in the flow before the one cycle of the time when the 2 nd injection operation failed is set_n-1The injection amount FQh2 at the normal time is set.

The fuel injection timings ITt1 and ITt2 for transition control may be set based on the injection timings ITh1 and ITh2 of the 1 st and 2 nd injection operations in the normal state.

In S109, the ignition timing Igt for transition control is set. In the present embodiment, the ignition timing Igt for transition control is set with reference to the normal ignition timing Igh.

In S110, the normal fuel injection amounts FQh1, FQh2 and the fuel injection timings ITh1, ITh2 for the 2 nd region Rh are set. Specifically, as in the 1 st region Rl, the basic fuel injection amount FQbase corresponding to the operating state of the engine 1 is calculated, and correction according to the cooling water temperature Tw or the like is performed, whereby the fuel injection amount FQ per combustion cycle is calculated. Then, a portion of the calculated fuel injection quantity FQ at a predetermined ratio (for example, 90%) is set as the injection quantity FQh1 of the 1 st injection operation, and the remaining portion is set as the injection quantity FQh2 of the 2 nd injection operation. Then, the injection amounts FQh1, FQh2 of the 1 st and 2 nd injection actions are substituted into the above expression (1) and converted into injection periods or injection pulse widths Δ t1, Δ t2, respectively, and the injection timing ITh1 of the 1 st injection action and the injection timing ITh2 of the 2 nd injection action are calculated. The distribution of the fuel injection amounts FQh1, FQh2 and the calculation of the fuel injection timings ITh1, ITh2 in the normal state can be performed by searching a map that is appropriately defined in advance by an experiment or the like, similarly to the basic fuel injection amount FQbase.

In S111, the normal ignition timing Igh for the 2 nd region Rh is set. In the 2 nd region Rh, the fuel injected by the 2 nd injection operation (fuel injection timing ITh2) is used as a spark to cause combustion in the entire cylinder, and the interval from the ignition timing Igh and the fuel injection timing ITh2 to the ignition timing Igh is set so that the peak value of the generated heat can be reached at a timing slightly exceeding the compression top dead center. Specifically, the ignition timing Igh is set to a timing in the compression stroke that is later than the ignition timing Igl in the 1 st region Rl, and is set immediately before the compression top dead center in the present embodiment.

In the present embodiment, the engine controller 101 constitutes a "controller", and the ignition plug 6, the fuel injection valve 7, and the engine controller 101 constitute a "control device for a direct injection engine". The function of the "operating state detecting unit" is realized by the processing of S101, the function of the "combustion state control unit" is realized by the processing of S102, S104, S107, S108, and S110, and the function of the "ignition control unit" is realized by the processing of S105, S109, and S111 in the flowchart shown in fig. 7.

Fig. 9 to 12 show the specific contents of the transition control according to the present embodiment by using a time chart.

The setting of the injection timing ITt2 and the ignition timing Igt for the 2 nd injection operation in the transfer control will be described with reference to fig. 9 to 12. In the present embodiment, the injection timing ITt1 of the 1 st injection action is set to the injection timing ITh1 of the 1 st injection action in the normal state immediately after the start of the transition control.

In the example shown in fig. 9, the interval Δ Cr from the injection timing ITt2 of the 2 nd injection action to the ignition timing Igt is made constant with respect to the crank angle over the entire control period from the start to the end of the transition control. On the other hand, the ignition timing Igt is retarded from the normal ignition timing Igh, which is the target ignition timing in the 2 nd region Rh, and then advanced to approach the normal ignition timing Igh in accordance with the decrease in the 2 nd transfer injection amount FQt 2. Since the interval Δ Cr from the injection timing ITt2 of the 2 nd injection event to the ignition timing Igt is constant, the injection timing ITt2 of the 2 nd injection event is also advanced in accordance with the advance of the ignition timing Igt.

In the example shown in fig. 10, the ignition timing Igt of the ignition plug 6 is set to the ignition timing Igh from the normal time immediately after the start of the shift control, and is maintained at a constant crank angle position throughout the control period of the shift control. On the other hand, the interval Δ Cr from the injection timing ITt2 of the 2 nd injection action to the ignition timing Igt is made shorter by the decrease in the 2 nd transfer injection amount FQt2 with respect to the larger interval immediately after the start of the transfer control. Since the ignition timing Igt is constant, the injection timing ITt2 of the 2 nd injection event lags behind the shortening of the interval Δ Cr.

In the example shown in fig. 11, the injection timing ITt2 of the 2 nd injection operation is set to the injection timing ITh2 in the normal state immediately after the start of the transition control, and is held at a constant crank angle position throughout the control period of the transition control. On the other hand, the interval Δ Cr from the injection timing ITt2 to the ignition timing Igt of the ignition plug 6 is made shorter by the decrease in the 2 nd transfer injection amount FQt2 with respect to the larger interval immediately after the start of the transfer control. The injection timing ITt2 is constant, so the ignition timing Igt at the retard-side crank angle position immediately after the start of the transition control is advanced in accordance with the shortening of the interval Δ Cr.

In the example shown in fig. 12, the ignition timing Igt of the ignition plug 6 is gradually retarded from the ignition timing Igl for the 1 st region Rl toward the target ignition timing (the normal-time ignition timing Igh) in the 2 nd region Rh, and at the same time, the interval Δ Cr from the injection timing ITt2 of the 2 nd injection action to the ignition timing Igt is made shorter in accordance with the decrease in the 2 nd transfer injection amount FQt2 with respect to the larger interval immediately after the start of the transfer control. By the shortening of the interval Δ Cr, the retard amount per control execution period of the injection timing ITt2 is made larger than the ignition timing Igt.

The above is the content of the combustion control according to the present embodiment, and the effects obtained by the present embodiment are summarized below.

(Explanation of action and Effect)

In the 1 st region Rl on the low load side, homogeneous combustion is performed, while in the 2 nd region Rh on the high load side, stratified combustion is performed by switching the combustion method, whereby the anti-knocking performance of combustion is improved, and therefore, knocking can be suppressed without excessively depending on the retardation of the ignition timing. Thereby, particularly, high thermal efficiency can be achieved in the entire operating region by improvement of the thermal efficiency in the 2 nd region Rh.

Further, when the region where the region Rl is shifted to the 2 nd region Rh is shifted, the shift control by the stratified combustion is executed, the fuel of an amount larger than the target amount (the normal injection amount FQh2) of the 2 nd injection operation in the 2 nd region Rh is injected by the 2 nd injection operation, and then the injection amount FQt2 of the 2 nd injection operation is made smaller toward the target amount, whereby the 2 nd injection operation of a smaller amount can be reliably executed, the required amount of fuel can be injected while ensuring the combustion stability, and the combustion method can be switched without impairing the combustion stability.

2, since the air excess ratio λ of the air-fuel mixture is set to a value near 2 in both the 1 st region Rl and the 2 nd region Rh, combustion with high thermal efficiency can be achieved, and fuel efficiency can be reduced.

In the 2 nd region Rh, the ignition timing Igh of the ignition plug 6 is delayed from the ignition timing Igl in the 1 st region Rl, so that the peak timing of the heat generated by combustion can be appropriately set based on the positional relationship with the piston 2, specifically, the compression top dead center can be set to a slightly excessive crank angle position. By performing the 2 nd injection operation based on the target amount immediately before the ignition timing Igh is reached, the mixed gas in the vicinity of the ignition plug 6 can be caused to flow by the kinetic energy of the fuel spray injected by the 2 nd injection operation, ignition can be performed while turbulence remains, and formation of an initial flame can be promoted to stabilize combustion.

In the transition control, the interval Δ Cr from the injection timing ITt2 of the 2 nd injection action to the ignition timing Igt is set to a constant value (fig. 9), whereby combustion can be stably generated. Then, by retarding the ignition timing Igt from the ignition timing in the normal state (target ignition timing) Igh and then advancing the ignition timing to approach the target ignition timing Igh in accordance with the decrease in the injection amount FQt2 of the 2 nd injection event, excessively rapid combustion can be avoided for the increase in the fuel injection amount FQt2 with respect to the target amount FQh 2.

In this way, retarding the ignition timing Igt with respect to the increase in the injection amount FQt2 in the 2 nd injection event can avoid excessively abrupt combustion. The suppression of combustion based on the delay of the ignition timing Igt is not limited to the example shown in fig. 9, and the suppression may be achieved by setting the injection timing ITt2 of the 2 nd injection event to a constant value and by shortening the interval Δ Cr from the injection timing ITt2 of the 2 nd injection event to the ignition timing Igt in accordance with the decrease in the injection amount FQt2 of the 2 nd injection event with respect to the interval immediately after the transition to the 2 nd region Rh (fig. 11).

Further, the suppression of combustion with respect to the increase in the injection quantity FQt2 of the 2 nd injection event is not limited to the delay of the ignition timing Igt, and may be achieved by changing the interval Δ Cr from the injection timing ITt2 of the 2 nd injection event to the ignition timing Igt as shown in fig. 10 and 12. Specifically, while the ignition timing Igt is set to a constant value, the interval Δ Cr from the injection timing ITt2 to the ignition timing Igt may be shortened in accordance with a decrease in the injection amount FQt2 of the 2 nd injection operation with respect to the interval immediately after the transition to the 2 nd region Rh (fig. 10), or the ignition timing Igt may be retarded from the ignition timing Igl in the 1 st region Rl toward the target ignition timing Igh in the 2 nd region Rh, and the interval Δ Cr from the injection timing ITt2 to the ignition timing Igt may be shortened in accordance with a decrease in the injection amount FQt2 of the 2 nd injection operation with respect to the interval immediately after the transition to the 2 nd region Rh (fig. 12).

The 5 th region Rh of the high load side can change the compression ratio CR of the engine 1, and the compression ratio CR (CRh) is made lower than the 1 st region Rl of the low load side in the 2 nd region Rh, whereby knocking can be suppressed without depending on the delay of the ignition timing.

Here, if the compression ratio CR is made to decrease, not only the thermal efficiency decreases, but also ignitability deteriorates due to the decrease in the in-cylinder temperature, and combustion becomes unstable. On the other hand, ignitability can be ensured by reducing the air excess ratio λ of the mixed gas and relatively increasing the amount of fuel in the mixed gas. However, in this case, not only the effect of improving fuel consumption due to the thinning of the mixed gas is offset, but also the NOx emission amount may increase.

In the present embodiment, since the explosion resistance of combustion is improved by performing stratified combustion in the 2 nd region Rh, explosion can be suppressed at a higher compression ratio than in the case of homogeneous combustion, and the fuel consumption rate can be reduced. Fig. 8 shows a case where the fuel consumption rate ISFC can be reduced by performing stratified combustion with respect to the 2 nd region Rh as compared with the case of homogeneous combustion (the fuel consumption rate in the case of homogeneous combustion is shown by a two-dot chain line). Further, since the ignitability can be ensured without lowering the excess air ratio λ by stratification of the mixed gas, high thermal efficiency can be maintained.

In the present embodiment, as shown in fig. 8, the compression ratio CR is increased in a stepwise manner when shifting from the 1 st region Rl to the 2 nd region Rh with respect to an increase in the engine load (however, in actual operation, there is a delay in the operation of the variable compression ratio mechanism according to the characteristics of the actuator 39, the link mechanisms 31, 32, 33, and the like). The compression ratio CRh for the 2 nd region Rh is not limited to such a setting, and may be continuously changed with respect to an increase in the engine load. For example, as shown in fig. 13, in the 2 nd region Rh, the compression ratio CRh is made to vary in such a manner that the difference from the compression ratio (shown by the two-dot chain line) that can suppress knocking in the case of homogeneous combustion is made large with respect to an increase in the engine load.

While the embodiments of the present invention have been described above, the above embodiments are merely illustrative of some application examples of the present invention, and the technical scope of the present invention is not limited to the specific configurations of the above embodiments. Various changes and modifications can be made to the above-described embodiment within the scope of the matters described in the claims.